centrifugal

CENTRIFUGAL COMPRESSORS A Basic Guide

Dr. Meherwan P.Boyce, P.E.

Disclaimer: The recommendations, advice, descriptions, and the methods in this book are presented solely for educational purposes. The author and publisher assume no liability whatsoever for any loss or damage that results from the use of any of the material in this book. Use of the material in this book is solely at the risk of the user. Copyright © 2003 by PennWell Corporation 1421 South Sheridan Road Tulsa, Oklahoma 74112 (800) 752-9764 [email protected] www.pennwell.com www.pennwell-store.com Cover design by Amy Spehar Book design by Robin Remaley Library of Congress Cataloging-in-Publication Data Boyce, Meherwan P. Centrifugal compressors : a basic guide / by Meherwan P. Boyce p. cm ISBN 0-87814-801-9 ISBN 978-0-87814-801-1 1. Compressors. I. Title. TJ267.5.C5 B59 2002 621.5’1--dc21 2002029815 All rights reserved. No part of this book may be reproduced, stored in a retrieval system, or transcribed in any form or by any means, electronic or mechanical including photocopying or recording, without the prior permission of the publisher. Printed in the United States of America. 2 3 4 5 6 7 8

12 11 10 09 08

Table of Contents LIST OF FIGURES. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ix LIST OF TABLES. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxxi LIST OF ACRONYMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxxiii LIST OF CONSTANTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxxv CHAPTER 1:

Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1

CHAPTER 2:

Aerothermodynamics of Compressors . . . . . . . . . . . . 35

CHAPTER 3:

Compressor Definition and Standards Defining Performance and Mechanical Equipment . . . . . . . . . . 69

CHAPTER 4:

Design Characteristics . . . . . . . . . . . . . . . . . . . . . . . . 123

CHAPTER 5:

Diffuser Design Characteristics . . . . . . . . . . . . . . . . . 183

CHAPTER 6:

Off-Design Performance Characteristics. . . . . . . . . . 205

CHAPTER 7:

Surge Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 225

CHAPTER 8:

Gas Turbines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 281

CHAPTER 9:

Steam Turbines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 359

CHAPTER 10: Electric Motors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 395 CHAPTER 11: Rotor Dynamics, Bearings, Lubrication Couplings, and Gears . . . . . . . . . . . . . . . . . . . . . . . . . 413 CHAPTER 12: Instrumentation Controls . . . . . . . . . . . . . . . . . . . . . . 491 CHAPTER 13: Compressor Performance Testing . . . . . . . . . . . . . . . 535 CHAPTER 14: Maintenance Techniques . . . . . . . . . . . . . . . . . . . . . . 569 INDEX . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 643

| vii |

1

Introduction

A compressor is a device that pressurizes a working fluid. One of the basic aims of using a compressor is to compress the fluid and deliver it at a pressure higher than its original pressure. Compression is required for a variety of purposes, some of which are listed below: •

To provide air for combustion

•

To transport process fluid through pipelines

•

To provide compressed air for driving pneumatic tools

•

To circulate process fluid through a certain process

Different types of compressors are shown in Figure 1-1. The positivedisplacement compressors are used for intermittent flow in which successive volumes of fluid are confined in a closed space to increase their pressures. The other broad class of compressors is the rotary type for continuous flow. In this type of compressor, rapidly rotating parts (impellers) accelerate fluid to a high speed; this velocity is then converted into additional pressure by gradual deceleration in the diffuser or volute, which surrounds the impeller. The positive-displacement type of compressors can be further classified as either reciprocating or rotary type, as shown in Figure 1-1. The reciprocating compressor has a piston having a reciprocating motion within a cylinder. The rotary positivedisplacement compressors have rotating elements whose positive action

|1|

| Centrifugal Compressors | results in compression and displacement. The rotary positivedisplacement can be further subdivided into sliding vane, liquid piston, straight-lobe and helical-lobe type compressors. The continuous-flow type compressors, as shown in Figure 1-1, can be classified under dynamic or ejector type, entrain the inflowing fluid using a high velocity gas or steam jet, and then convert the velocity of the mixture to pressure in a diffuser. The dynamic compressors have rotating elements, which accelerate the inflowing fluid, and convert the velocity head into pressure head, partially in the rotating elements and partially in the stationary diffusers or blade. The dynamic type can be further subdivided into centrifugal, axial-flow, and mixed-flow compressors. The main flow of gas in the centrifugal compressor is radial. The flow of gas in an axial compressor is axial, and the mixed-flow compressor combines some characteristics of centrifugal and axial compressors.

Compressor Positive Displacement Reciprocating Sliding Vane

Continuous Flow Ejector

Rotary

Liquid Piston

Straight Lobe

Helical

Centrifugal

Dynamic Axial

Mixed Flow

Figure 1-1: Principal types of compressors.

COMPRESSOR SELECTIONS It is not always obvious what type of compressor is needed for an application. Of the many types of compressors mostly used in the process industry, some of the more significant are the centrifugal, axial, rotary, and the reciprocating compressors. They fall into three categories, as shown in Figure 1-2. For very high flows and low-pressure ratios, an axial-flow compressor is best. Axial-flow compressors usually have a higher efficiency, as seen in Figure 1-3, but a smaller operating region than a centrifugal machine. Centrifugal compressors operate most efficiently at medium flow rates and high-pressure ratios. Rotary and reciprocating

|2|

| Introduction | compressors (positive-displacement machines) are best used for low flow rates and high-pressure ratios. The positive displacement compressors, more commonly known as the reciprocating compressor, were the most widely used compressors in the process and pipeline industries up to and through the 1960s.

30

POSITIVE DISPLACEMENT

Pressure Ratio

20

CENTRIFUGAL COMPRESSOR 10

AXIAL FLOW COMPRESSOR

1 10 2

10 3

10 4

Flow

10 5

10 6

(CFM)

Figure 1-2: Performance characteristics of different types of compressors.

In the 1960s, the centrifugal flow compressors became popular because their efficiency was comparable to that of the reciprocating compressor, and because of its much lower maintenance costs. Today the centrifugal compressor is the main compressor in the process and pipeline industries. Due to its compact size and its comparable light weight, the centrifugal compressor is used extensively in the offshore industry. The centrifugal and axial flow compressors fall into the category known as the turbomachinery group of machines. The turbomachine group of machines mostly consists of high speed rotating machines, with steady flow characteristics. Centrifugal compressors are an integral part of the chemical process industries. They are used extensively because of their smooth operation, large tolerance-to-process fluctuations, and higher reliability than other types of compressors. The centrifugal compressor may be known as a fan,

|3|

| Centrifugal Compressors | blower, booster, or exhauster. Broadly speaking, fans are low-pressure compressors and blowers are medium-pressure compressors. Boosters and exhausters are named for their application. Before giving an in-depth discussion about centrifugal compressors, let us look at different types of compressors and their applications.

Figure 1-3: Variation of adiabatic efficiency with specific speed for three types of compressors.

In turbomachinery the centrifugal flow and axial flow compressors, which are continuous flow compressors, are the ones used for compressing the air. Positive displacement compressors such as the, reciprocating, gear type, or lobe type to name just a few, are widely used in the industry for many applications. In this book, we are examining the centrifugal compressor in its many applications from the large lower pressure process compressor to the high-pressure centrifugal compressors used in small gas turbine applications. The characteristics of these compressors are given in Table 1-1. The pressure ratio of the axial and centrifugal compressors has been classified into three groups: industrial, aerospace, and research. The industrial pressure ratio is low for the reasons that the operating range needs to be large. The operating range is defined as the range between the surge point and the choke point. Figure 1-4 shows the

|4|

| Introduction | operating characteristics of a compressor. The surge point is the point when the flow is reversed in the compressor. The choke point is the point when the flow has reached a Mach=1.0, the point where no more flow can get through the unit, a “stone wall.” When surge occurs, the flow is reversed and so are all the forces acting on the compressor, especially the thrust forces, which can lead to total destruction of the compressor. Thus, surge is a region that must be avoided. Choke conditions cause a large drop in efficiency but do not lead to destruction of the unit. TYPES OF COMPRESSORS Positive Displacement Centrifugal Axial

PRESSURE RATIO Industrial Aerospace Research Up to 30 -

EFFICIENCY

OPERATING RANGE

75%-82%

Large 25% Narrow 3%-10%

1.2-1.9

2.0-7.0

13

75%-87%

1.05-1.3

1.1-1.45

2.1

80%-91%

Table 1-1: Compressor characteristics. It is important to note that the operating range is narrowed with the increase in pressure ratio and the number of stages.

Figure 1-4: Operating characteristics of a compressor.

|5|

| Centrifugal Compressors | AXIAL-FLOW COMPRESSORS Axial-flow compressors are used mainly as the gas compressor in large gas turbines, due to their smaller frontal area for the same high flow. They have also been used as process compressors where efficiency is important, the flow high, and the pressure ratio low. They are used sparingly in the process industry because they are much more expensive than their centrifugal counterpart, they have a narrower operation margin (surge to choke margin), and they are very susceptible to foreign object damage. An axial-flow compressor compresses its working fluid by first accelerating the fluid, and then diffusing it to obtain a pressure increase. The fluid is accelerated by a row of rotating airfoils or blades (the rotor) and diffused by a row of stationary blades (the stator). The diffusion in the stator converts the velocity increase gained in the rotor to a pressure increase. One rotor and one stator make up a stage in a compressor. A compressor usually consists of multiple stages. One additional row of fixed blades (inlet guide vanes) is frequently used at the compressor inlet to ensure that air enters the first-stage rotors at the desired angle. In addition to the stators, an additional diffuser at the exit of the compressor further diffuses the fluid and controls its velocity. In an axial compressor, air passes from one stage to the next, with each stage raising the pressure slightly. By producing low-pressure increases on the order of 1.1:1-1.4:1, very high efficiencies can be obtained. Using multiple stages permits overall pressure increases up to 40:1. The rule of thumb for a multiple stage gas turbine compressor is that the energy rise per stage would be constant, rather than the pressure rise per stage. Figure 1-5 shows a multistage high-pressure axial flow turbine rotor. The turbine rotor depicted in this figure has a low-pressure compressor followed by a high-pressure compressor. There are also two turbine sections; the reason there is a large space between the two turbine sections is that this is a reheat turbine, and the second set of combustors are located between the high pressure and the low pressure turbine sections. The compressor produces 30:1 pressure in 22 stages. As with other types of rotating machinery, an axial compressor can be described by a cylindrical coordinate system. The Z-axis is taken as running the length of the compressor shaft, the radius r is measured outward from the shaft, and the angle of rotation θ is the angle turned by the blades in Figure 1-6. This coordinate system will be used throughout this discussion of compressors in this book.

|6|

| Introduction |

Figure 1-5: A high pressure ratio turbine rotor. (Courtesy ALSTOM)

Figure 1-6: Coordinate system for axial flow compressor.

|7|

| Centrifugal Compressors | Figure 1-7 shows the pressure, velocity, and total enthalpy variation for flow through several stages of an axial compressor. It is important to note here that the changes in the total conditions for pressure, temperature, and enthalpy occur only in the rotating component where energy is inputted into the system. As seen in Figure 1-6, the length of the blades and the annulus area, which is the area between the shaft and shroud, decreases through the length of the compressor. This reduction in flow area compensates for the increase in fluid density as it is compressed, permitting a constant axial velocity.

Figure 1-7: Variation of flow and thermodynamic properties in an axial flow compressor.

CENTRIFUGAL FLOW COMPRESSORS Centrifugal compressors are an integral part of the petrochemical industry, finding extensive use because of their smooth operation, large tolerance of process fluctuations, and their higher reliability compared to other types of compressors. They are also used in small gas turbines. The centrifugal compressors range in size from pressure ratios of 1.3:1 per stage in the process industries, to 3-7:1 per stage in small gas turbines, and as high as 13:1 on experimental models. This means that the compressor pressure ratio must be between. This is considered a highly loaded

|8|

| Introduction | compressor. With pressure ratios that exceed 5:1, flows entering the diffuser from the impeller are supersonic in their mach number (M>1.0). This requires special design of the diffuser. The centrifugal compressor has a limited stable operating range. The capacity varies from 45% to 90% of rated capacity. This may affect the economics of operating at partial load. The centrifugal type compressor should be selected for the worst possible conditions, but at the same time, meet other design requirements. The operating speed of the centrifugal compressor is higher than that for other compressors. For aircraft and space applications, the rpm can range from 50,000 to 100,000. Most commercial units run below 20,000 rpm. With the trend toward increasing the rpm, problems due to bearing lubrication, vibration, and balancing are becoming more significant at higher speeds. Centrifugal compressors are well suited for direct connection to gas or steam turbine drives which have variable-speed control. Due to absence of inertia forces, centrifugal compressors require smaller and less expensive foundations. These machines have a high availability factor, frequently operate for 2 to 3 years without shutdown, and require less maintenance than the reciprocating type. In a typical centrifugal compressor, the fluid is forced through the impeller by rapidly rotating impeller blades. The velocity of the fluid is converted to pressure, partially in the impeller and partially in the stationary diffusers. Most of the velocity leaving the impeller is converted into pressure energy in the diffuser as shown in Figure 1-8. It is normal practice to design the compressor so that half the pressure rise takes place in the impeller and the other half in the diffuser. The diffuser consists essentially of vanes, which are tangential to the impeller. These vane passages diverge to convert the velocity head into pressure energy. The inner edge of the vanes is in line with the direction of the resultant airflow from the impeller, as shown in Figure 1-9. In the centrifugal or mixed-flow compressor, the air enters the compressor in an axial direction and exits in a radial direction into a diffuser. This combination of rotor (or impeller) and diffuser comprises a single stage. The air initially enters a centrifugal compressor at the inducer, as shown in Figure 1-8. The inducer, usually an integral part of the impeller, is very much like an axial-flow compressor rotor. Many earlier designs kept the inducer separate. The air then goes through a 90º turn and exits into a diffuser, which usually consists of a vaneless space followed by a vaned diffuser. This is especially true if the compressor exit is supersonic, as is the case with high-pressure ratio compressors. The vaneless space is

|9|

| Centrifugal Compressors |

Figure 1-8: Aerodynamic and thermodynamic properties in a centrifugal compressor stage.

Figure 1-9: Flow in a vaned diffuser.

| 10 |

| Introduction | used to reduce the velocity leaving the rotor to a value lower than Mach number =1 (M 90°, the vanes are forwardcurved or forward-swept. They have different characteristics of theoretical head-flow relationship to each other. In Figure 1-10, the forward-curved blade has the highest theoretical head. In actual practice, the head characteristics of all the impellers are similar to the backwardcurved impeller. Most applications use backward curved blades since they

| 11 |

| Centrifugal Compressors |

Figure 1-10: Theoretical head characteristics as a function of the flow in a centrifugal impeller.

TYPES OF IMPELLERS Radial Blades

Backward Curved Blades

Forward Curved Blades

ADVANTAGES 1. Reasonable compromise between low energy transfer and high absolute outlet velocity 2. No complex bending stress 3. Ease in manufacturing 1. Low outlet kinetic energy 2. Low diffuser inlet Mach Number 3. Surge margin is widest of the three High Energy Transfer

Table 1-2: Impeller designs—advantages and disadvantages.

| 12 |

DISADVANTAGES Surge Margin is Narrow

1. Low energy transfer 2. Complex Bending Stress 3. Difficulty in Manufacturing 1. High-outlet Kinetic Energy 2. High Diffuser Inlet Mach Number 3. Complex Bending Stress 4. Difficulty in Manufacturing

| Introduction | have the lowest velocity leaving the impeller, thus the diffuser has a much smaller dynamic head to convert. Also, backward curved blades have a much larger operational margin. Table 1-2 shows the advantages and disadvantages of various impeller designs. Diffusers form an important part of a centrifugal compressor, and usually are the most difficult to design. The function of the diffuser in a compressor is the conversion of dynamic or kinetic head generated by the impeller to pressure energy. This conversion is essential for obtaining the required pressure rise of the compressor and also for achieving good efficiency in the gas transmission along the supply pipe. In a gas turbine engine using a centrifugal compressor, the air is required to negotiate through several narrow passages and bends before arriving at the combustion chamber. The bends in these narrow passages will reduce the total energy of airflow; this energy loss can be reduced when an efficient diffuser, and hence low velocities, occur in these passages and bends. A low velocity is also essential in the combustion chamber for achieving high combustion efficiencies. A diffuser is, hence, a component of critical importance when optimum efficiency is a requirement in turbomachinery. Figure 1-8 shows the static pressure and velocity changes in a centrifugal compressor and diffuser. The diffuser assembly may be an integral part of the compressor casing or a separately attached assembly. In each instance, it consists of a number of vanes formed tangentially to the impeller as seen in Figure 1-9. The vane passages are divergent to convert the kinetic energy into pressure energy, and the inner edges of the vanes are in line with the direction of the resultant airflow from the impeller. The clearance between the impeller and the diffuser is an important factor, as too small a clearance will set up aerodynamic buffeting impulses that could be transferred to the impeller and create unsteady airflow and vibration. A P P L I C AT I O N O F C E N T R I F U G A L C O M P R E S S O R S The centrifugal compressor has many applications, requiring it to have many performance characteristics. Centrifugal compressors used in gas turbines are required to have a high-pressure ratio and have a narrow operating range. Centrifugal compressors operating in the process industry have a need for a large operating range and thus operate at a small pressure ratio. Table 1-3 summarizes some important applications of centrifugal compressors.

| 13 |

| Centrifugal Compressors | INDUSTRY OR APPLICATION Gas Turbines Power Drive Iron and Steel Blast furnace

Mining and Metallurgy

Natural Gas

Refrigeration

Utilities

Miscellaneous

Bessemer Converter Cupola Coke Oven Power Furnaces

Production Distribution Processing Chemical

SERVICE OR PROCESS Compression Compression Combustion Off gas Oxidation

TYPICAL GAS HANDLED Air Air Air Blast furnace gas Air

Combustion Compression For Tools and Machinery Copper and nickel Purification Pelletizing (Iron Ore Concentration) Re-pressuring oil wells Transmission Natural Gasoline Separation Refrigeration Various Processes

Air Coke Oven Gas Air Air

Industrial & Air conditioning Commercial Steam Soot blowing Generators Combustion Cyclone furnaces City Gas Manufacturing Distribution Sewage Agitation Treatment Industrial Power for tools Power and Machines Paper Fourdrinier vacuum Making Material Conveying handling Gas Engines Supercharging

Table 1-3: Applications of centrifugal compressors.

| 14 |

Air Natural Gas Natural Gas Natural Gas Propane and methane Butane, Propane, Ethylene, Ammonia Special Refrigerants Special Refrigerants Air Air Air Fuel gas Fuel gas Air Air Air and water vapor

Air

| Introduction | CENTRIFUGAL COMPRESSORS IN GAS TURBINES Many small gas turbines that produce below 5 MW incorporate centrifugal compressors, or combinations of centrifugal and axial compressors, as well as radial-inflow turbines. A small turbine will often consist of a single-stage centrifugal compressor producing a pressure ratio as high as 6:1, a single side combustor where temperatures of about 1,800°F (982°C) are reached, and radial-inflow turbines. Figure 1-11 shows a schematic of such a typical turbine. Air is induced through an inlet duct to the centrifugal compressor, which, rotating at a high speed, imparts energy to the air. On leaving the impeller air with increased pressure and velocity, it passes through a high-efficiency diffuser, which converts the velocity energy to static pressure. The compressed air, contained in a pressure casing, flows at low speed to the combustion chamber, which is a side combustor. A portion of the air enters the combustor head, mixes with the fuel, and burns continuously. The remainder of the air enters through the wall of the combustor and mixes with the hot gases. Good fuel atomization and controlled mixing ensure an even temperature distribution in the hot gases, which pass through the volute to enter the radial inflow turbine nozzles. High acceleration and expansion of the gases through the nozzle guide vane passages and turbine combine to impart rotational energy, which is used to drive the external load and auxiliaries on the cool side of the turbine.

Figure 1-11: A small radial flow gas turbine cutaway showing the turbine rotor.

| 15 |

| Centrifugal Compressors | The efficiency of a small turbine is usually much lower than a larger unit because of the limitation of the turbine inlet temperature and the lower component efficiencies. Turbine inlet temperature is limited because the turbine blades are not cooled. Radial-flow compressors and impellers inherently have lower efficiencies than their axial counterparts. These units are rugged, and their simplicity in design assures many hours of trouble-free operation. A way to improve the lower overall cycle efficiencies, 18%-23%, is to use the waste heat from the turbine unit. High thermal efficiencies (30%-35%) can be obtained, since nearly all the heat not converted into mechanical energy is available in the exhaust, and most of this energy can be converted into useful work. These units, when placed in a combined heat power application, can reach efficiencies of the total process as high as 60%-70%. Figure 1-12 shows an aeroderivative small gas turbine. This unit has three independent rotating assemblies mounted on three concentric shafts. This turbine has a three-stage axial flow compressor followed by a centrifugal compressor, each driven by a single stage axial flow compressor. Power is extracted by a two-stage axial flow turbine and delivered to the inlet end of the machine by one of the concentric shafts. The combustion system comprises a reverse flow annular combustion chamber, with multiple fuel nozzles and a spark igniter. This aeroderivative engine produces 4.9 MW and has an efficiency of 32%.

Axial Flow Compressor

Centrifugal Compressor Concentric Shafts

Axial Flow Turbine

Figure 1-12: A small aeroderivative gas turbine. ST30 marine and industrial gas turbine engine. (Courtesy Pratt & Whitney Canada Corporation)

| 16 |

| Introduction | Micro-turbines usually refer to units of less than 350 kW. These units are usually powered by either diesel fuel or natural gas. The micro turbines can be either axial flow or centrifugal-radial inflow units. The initial cost, efficiency, and emissions will be the three most important criteria in the design of these units. As seen in Figure 1-13, today’s micro turbines are using radial flow turbines and compressors, due to the compactness and ruggedness of these types of compressors and turbines.

Figure 1-13: A compact microturbine schematic. (Courtesy Capstone Corporation)

CENTRIFUGAL COMPRESSORS IN THE PROCESS INDUSTRY The common method of classifying process-type centrifugal compressors is based on the number of impellers and the casing design. Table 1-4 shows three types of centrifugal compressors. For each type of compressor, approximate maximum ratings of pressure, capacity, and brake horsepower are also shown.

| 17 |

| Centrifugal Compressors | CASING TYPE

APPROXIMATE MAXIMUM RATINGS Approximate Approximate Approximate Pressure Inlet Capacity Power psig (Bar) cfm (cmm) Horsepower (kW) 1. Integral Gear Type Compressor Single Stage 45 psig 250,000 cfm 3,000 HP (3.0 bar) (7,079 cmm) (2,241 kW) Multistage More than 600 psig 60,000 cfm 10,000 HP (40 bar) (102 cmm) (7,470 kW) 2. Horizontally Split Casings Single Stage 15 psig 650,000 cfm 10,000 HP (double suction) (1.03 bar) (18,406 cmm) (7,457 kW) Multistage 1000 psig 200,000 cfm 35,000 HP (69 bar) 5,663 cmm (26,100 kW) 3. Barrel Type Compressor Pipeline 1200 psig 25,000 cfm 20,000HP (82 bar) (708 cmm) (14,914 kW) Multistage More than 5500 psig 20,000 cfm 15,000 HP (379 bar) (566 cmm) (11,185 kW) Table 1-4: Industrial centrifugal compressor classification based on casing design.

Figure 1-14: An integral geared centrifugal compressor, showing intercoolers below the compressor base plate. (Courtesy Atlas Copco Comptec, Inc.)

| 18 |

| Introduction |

Figure 1-15: Horizontally split centrifugal compressor, with closed-face impellers. (Courtesy Man-Turbo)

Integrated gear type units have impellers, which are usually mounted on the extended motor shaft, and similar sections are mounted to obtain the desired number of stages. Casing material is either steel or cast iron. These machines require minimum supervision and maintenance and are quite economic in their operating range. The integrated gear casing design is used extensively in supply of lighter gases, such as CO, CO2, H2, and air. Figure 1-14 is a typical integral gear multistage centrifugal compressor. The horizontally split type have casings split horizontally at the midsection and the top as shown in Figure 1-15. The bottom halves are bolted and doweled together as shown in Figure 1-16. This design type is preferred for large multistage units. The internal parts such as shaft, impellers, bearings, and seals are readily accessible for inspection and repairs by removing the top half. The casing material is cast iron or cast steel. There are various types of barrel or centrifugal compressors. Lowpressure types with overhung impellers are used for combustion processes, ventilation, and conveying applications. Multistage barrel casings are used for high-pressures in which the horizontally split joint is

| 19 |

| Centrifugal Compressors |

Figure 1-16: Horizontal casing centrifugal compressor. (Courtesy MAN Turbomaschinen AG Schweiz)

inadequate. Figure 1-17 shows the barrel compressor in the background and the inner bundle from the compressor in front. Once the casing is removed from the barrel, it is horizontally split, as shown in Figure 1-18. Compressor trains can be a combination of axial flow compressors as well as centrifugal compressors. Figure 1-19 shows a long train, which is a typical arrangement for nitric acid plants. Here the train is driven by a 5.9 MW steam turbine, an axial air compressor, followed by a centrifugal compressor for the nitrous gas, and a tail gas 11MW axial flow expander.

| 20 |

| Introduction |

Figure 1-17: Horizontal casing centrifugal compressor. (Courtesy Dresser Rand Corporation)

Figure 1-18: The inner bundle of a barrel centrifugal compressor opened (Courtesy MAN Turbomaschinen AG Schweiz)

| 21 |

| Centrifugal Compressors |

Figure 1-19: The long train is a typical arrangement for acid nitric plants (process UHDE): steam turbine (43.5 t/h; 5.9MW) as driver, axial air compressor (AV56-14; 139300Nm3/h), centrifugal compressor (R71-3; 122500Nm3/h) for nitrous gas and tail gas expander (E56-4; 111300Nm3/h, 11MW). (Courtesy MAN Turbomaschinen AG Schweiz)

C O M P R E S S O R I N T E R N A L C O N F I G U R AT I O N To properly design a centrifugal compressor, one must know the operating conditions – the type of gas, its pressure, temperature, and molecular weight. One must also know the corrosion properties of the gas so that proper metallurgical selection can be made. Gas fluctuations due to process instabilities must be pinpointed so that the compressor can operate without surging. The process compressors are all designed to meet the rigorous API specifications. The API specifications, which govern the compressors from a mechanical point of view, are API Std 617, Centrifugal Compressors for Petroleum, Chemical and Gas Industry Services, 6th Edition, February 1995, and the API Std 672, Packaged, Integrally Geared Centrifugal Air Compressors for Petroleum, Chemical, and Gas Industry Services, 3rd Edition, September 1996. The performance of these compressors is based on the ASME, Performance Test Code on Compressors and Exhausters, ASME PTC 10 1997, American Society of Mechanical Engineers 1997. There are two types of centrifugal compressors used in the process industry one for air and lighter gases and the other for process gases. The process air compressors are usually compressors with open-faced impellers, each stage having a single stage pressure ratio of up to 3:1, as seen in Figure 1-20. The process gas compressors have close-faced impellers, as shown in Figure 1-21, with a

| 22 |

| Introduction |

Figure 1-20: An integral geared centrifugal compressor rotor assembly showing intercoolers below the compressor base plate. (Courtesy Atlas Copco Comptec, Inc.)

Figure 1-21: Two sets of closed-face impellers used in a centrifugal compressor. Note the wide size differential. (Courtesy Dresser Rand Corporation)

| 23 |

| Centrifugal Compressors | very low-pressure ratio up to 1.1-1.3, and thus have large operating margins (surge-to-choke margins). The two rotor assemblies shown in Figure 1-21 show the large variation in flow and pressure that these compressors cover. Figure 1-22 shows centrifugal compressor rotors and axial flow compressors.

Figure 1-22: Axial flow compressors and centrifugal compressor. (Courtesy Sulzer Corporation)

Centrifugal compressors for industrial applications have relatively low-pressure ratios per stage. This condition is necessary so that the compressors can have a wide operating range while stress levels are kept at a minimum. Because of the low-pressure ratios for each stage, a single machine may have a number of stages in one “barrel” to achieve the desired overall pressure ratio. Figure 1-23 shows some of the many configurations. The first two schematics shown in Figure 1-23 are integrally geared centrifugal compressors. The rest of the schematics are for a large process gas compressor.

| 24 |

| Introduction |

Figure 1-23: Various configurations of centrifugal compressors.

| 25 |

| Centrifugal Compressors | Some factors to be considered when selecting a configuration to meet plant needs are: •

Intercooling between stages can considerably reduce the power consumed.

•

Back-to-back impellers allow for a balanced rotor thrust and minimized overloading of the thrust bearings.

•

Cold inlet or hot discharge at the middle of the case reduces oilseal and lubrication problems.

•

Single inlet or single discharge reduces external piping problems.

•

Balance planes that are easily accessible in the field can appreciably reduce field-balancing time.

•

Balance piston with no external leakage will greatly reduce wear on the thrust bearings.

•

Hot and cold sections of the case that are adjacent to each other will reduce thermal gradients and thus reduce case distortion.

•

Horizontally split casings are easier to open for inspection than vertically split ones, reducing maintenance time.

•

Overhung rotors present an easier alignment problem because shaft-end alignment is necessary only at the coupling between the compressor and driver.

•

Smaller, high-pressure compressors that do the same job will reduce foundation problems but will have greatly reduced operational range.

These integrally geared centrifugal compressors have open-faced impellers, which are cantilevered. These compressors are used in many services, such as compressing air, CO, CO2, and H2. They can also be used for the simultaneous processing of different gases and for energy recovery and process control applications. The integrally geared compressor has many varied configurations. These range from single-stage units to multistage units. In these compressors the power is usually transmitted through the bull gear to the pinion gear on which the impeller is assembled, as shown in Figure 1-24. Figure 1-25 shows an integral geared single stage

| 26 |

| Introduction |

Figure 1-24: An integral geared single stage centrifugal compressor, with a cast scroll exit, for mechanical vapor recompression service. (Courtesy Atlas Copco Comptec, Inc.)

Figure 1-25: Schematic of an integral geared centrifugal compressor. (Courtesy Atlas Copco Comptec, Inc.)

| 27 |

| Centrifugal Compressors |

Figure 1-26: Schematic of a typical rotor/bull gear configuration for a six stage compressor. (Courtesy Atlas Copco Comptec, Inc.)

centrifugal compressor, with a cast scroll exit, for mechanical vapor recompression service. In multiple stage integrally geared centrifugal compressors, the power is usually transmitted through the bull gear to the two pinion gears for a four stage compressor, as shown in Figure 1-26, which is a schematic of such an arrangement. Multistaging with up to six stages on one gearbox, as shown in Figure 1-27, is a uniquely cost effective configuration. Placing an expander on one pinion to recover waste energy is another unique way to have an efficient compressor. These compressors can use multiple gases in the same compressor.

| 28 |

| Introduction |

Figure 1-27: An integral geared centrifugal compressor, showing intercoolers below the compressor base plate. (Courtesy Atlas Copco Comptec, Inc.)

Figure 1-28 shows a typical multistage, integrally geared compressor. Note the three intercoolers under the base plate. This is to provide intercooling between stages, thus reducing the power needed to compress the gases. The large industrial compressors have many configurations, as shown in Figure 1-23. The most common type is a straight, flow-through compressor with all compressor impellers facing the same direction, as shown in Figure 1-28. This type of compressor has a high thrust load. The back-to-back pressure compressor, as shown in Figure 1-29, is a double flow inlet with many advantages. It reduces the inlet relative mach number, and also reduces the thrust created due to the balanced rotor thrust. If knockout drums are needed between stages, the back-to-back compressor, as shown in Figure 1-30, has a great advantage, since it does not contaminate the flow entering the stage after the knockout drum due to leakage from the previous higher-pressure stage. This action does not pollute the cleaner gas, as it would occur in straight flow-through compressors. Back-to-back impellers allow for a balanced rotor thrust and minimized overloading of the thrust bearings.

| 29 |

| Centrifugal Compressors |

Figure 1-28: Straight flow-through centrifugal compressor. (Courtesy Dresser Rand Corporation)

Figure 1-29: Back-to-back compressor with double flow inlet. (Courtesy Dresser Rand Corporation)

Side loaded compressors are also often used in process plants. These types of compressors have side loadings, as seen in Figure 1-31. In this compressor, the gases are brought in at three sections. This type of compressor requires a complex surge system, and, during startup, there are loading problems that must be taken into account to reduce the torque requirements.

| 30 |

2nd Section Entry

1st Section Discharge

1st Section Entry

Rotation

2nd Section Discharge

| Introduction |

Figure 1-30: Back-to-back centrifugal compressor. (Courtesy Dresser Rand Corporation)

Figure 1-31: Side-loaded compressor. (Courtesy Enterprise LLP)

| 31 |

| Centrifugal Compressors | I M P E L L E R FA B R I C AT I O N Centrifugal compressor impellers are either shrouded or unshrouded. Open, shrouded impellers that are mainly used in single-stage applications are made by investment casting techniques, or by three-dimensional milling, as seen in Figure 1-32. Such impellers are used, in most cases, for the high-pressure ratio stages. The shrouded impeller (as shown in Figure 1-33) is commonly used in the process compressor because of its lowpressure ratio stages. The low tip stresses in this application make it a feasible design. Figure 1-34 shows several fabrication techniques. The most common type of construction is seen in A and B where the blades are fillet-welded to the hub and shroud. In B, the welds are full penetration. The disadvantage in this type of construction is the obstruction of the aerodynamic passage. In C, the blades are partially machined with the covers and then butt-welded down the middle. For backward lean-angled blades, this technique has not been very successful, and there has been difficulty in achieving a smooth contour around the leading edge. D illustrates a slot-welding technique and is used where blade-passage height is too small (or the backward lean-angle too high) to permit conventional fillet welding. In E, an electron-beam technique is still in its infancy, and work needs to be done to perfect it. Its major disadvantage is that electron-beam welds should preferably be stressed in tension but, for the configuration of E, they are in shear. Configurations G through J use rivets. Where the rivet heads protrude into the passage, aerodynamic performance is reduced. Materials for fabricating these impellers are usually low-alloy steels, such as AISI 4140 or AISI 4340. For most application, AISI 4140 is satisfactory. AISI 4340 is used for larger impellers requiring higher strengths. For corrosive gases, AISI 410 stainless steel (about 12% chromium) is used. Monel K-500 is employed in halogen gas atmospheres and oxygen compressors because of its resistance to sparking. Titanium impellers have been applied to chlorine service. Aluminum-alloy impellers have been used in great numbers, especially at lower temperatures (below 300°F). With new developments in aluminum alloys, this range is increasing. Aluminum and titanium are sometimes selected because of their low density. This low density can cause a shift in the critical speed of the rotor, which may be advantageous.

| 32 |

| Introduction |

Figure 1-32: Open-faced impeller fabrication. (Courtesy MAN Turbomaschinen AG Schweiz)

| 33 |

| Centrifugal Compressors |

Figure 1-33: Closed-faced impeller. (Courtesy MAN Turbomaschinen AG Schweiz)

Figure 1-34: Several fabrication techniques for centrifugal impellers.

| 34 |

2

Aerodynamics of Compressors

This chapter examines the overall performance characteristics of centrifugal compressors. This material will familiarize the reader with the behavior of these machines, classified under the broad term turbomachinery. Pumps and compressors are used to produce pressure; turbines produce power. These machines have some common characteristics. The main element is a rotor with blades or vanes, and the path of the fluid in the rotor may be axial, radial, or a combination of both. There are three methods of studying the elements of turbomachinery operation. First, by examining forces and velocity diagrams, it is possible to discover some general relationships between capacity, pressure, speed, and power. Second, comprehensive experimentation can be undertaken to study relationships between different variables. Third, without considering the actual mechanics, one can use dimensional analysis to derive a set of factors whose grouping can shed light on overall behavior. The analysis presented in this chapter shows the typical performance diagrams one can expect from turbomachines. Aerothermodynamics is fundamental to the aerodynamics and thermodynamics of all types of turbomachines. Turbomachines are considered to be steady-flow machinery, operating with gases that have a low viscosity, and are rotating at a high velocity. The equations and processes described in this chapter govern all types of turbomachines with emphasis on the centrifugal compressor.

| 35 |

| Centrifugal Compressors | D I M E N S I O N A L A N A LY S I S In attempting to predict the behavior of a proposed design in engineering, several approaches are possible. One approach is to use empirical equations based upon previous experience in similar situations. This approach to a problem is quite satisfactory for situations that do not differ greatly from those used to develop the empirical information. The empirical approach must be used in some cases where no other method is available. The shortcoming of this method is that it may not be used with any degree of confidence in situations that are entirely new or different. In design situations where no previous experience in the proposed area is available, the engineer may choose to perform laboratory experiments to develop the needed information. In such cases the use of dimensional analysis is most helpful. It permits the use of experimental equipment of a scale, which differs from actual equipment and also permits a reduction in the number of measurements required. This method has been used in the field of turbomachinery to scale models to predict the behavior of full-scale compressors or turbines. The method has limitations in that not all desired conditions can be simulated in the laboratory and not all physical phenomenon can be easily scaled. This is especially true when it comes to the structural and mechanical integrity of the scaled up system. Nevertheless, the method is powerful and very commonly used, but must be used with great care. Another method for predicting behavior of a proposed engineering system is by the use of analogies. An analogy is a similarity in some respects between things otherwise unlike. Some engineering systems behave in similar ways to an entirely different type of system. For example, the flow of electricity and the flow of heat in a solid conductor are mathematically similar. The engineer may study the behavior of a certain selected electrical system and predict the behavior of a similar thermal system by analogy. This method is particularly useful since electrical circuits are easily assembled and electrical measurements may be easily made. The analytical approach often yields valuable results. The physical laws governing the engineering system may be applied and predictions made from these laws. The reliability of this method is limited only by the number of simplifying assumptions made and by the mathematical techniques used to obtain desired information. This method is particularly useful in predicting behavior of physical systems where no experience is

| 36 |

| Aerothermodynamics of Compressors | available, and where the analogies cannot be applied. In fluid mechanics and heat transfer, the physical laws that are important in predicting behavior may be classified as follows: •

Conservation of mass

•

Conservation of momentum (Navier–Stokes Equation)

•

Conservation of energy (First Law of Thermodynamics)

•

Second law of thermodynamics

•

Phenomenological equations

A partial list of some phenomenological laws includes: •

Equation of state (for gases)

•

Newton’s viscosity law

•

Hooke’s law (for elastic solids)

•

Fourier’s heat conduction law

•

Stefan-Boltzmann’s radiation law

•

Newton’s law of cooling

•

Fick’s law of diffusion

•

Ohm’s law (for electrical conductors)

A solution to a problem may be obtained by formulating the above laws as they apply to control volumes or systems and deriving suitable equations. In some cases these laws are expressed in terms of differential equations, valid at each point in the substance under study. By appropriate simplification of the equations and by application of suitable boundary and initial conditions, the equations may be solved for the desired information. The solution in some cases may be approximate in nature; for example, certain differential equations yield only to numerical solutions. Turbomachines can be compared with each other by dimensional analysis. This analysis produces various types of geometrically similar parameters. Dimensional analysis is a procedure where variables representing a physical situation are reduced into dimensionless groups.

| 37 |

| Centrifugal Compressors | These dimensionless groups can then be used to compare performance of various types of machines with each other. Dimensional analysis as used in turbomachines can be employed to: (1) compare data from various types of machines—it is a useful technique in the development of blade passages and blade profiles, (2) select various types of units based on maximum efficiency and pressure head required, and (3) predict a prototype’s performance from tests conducted on a smaller scale model or at lower speeds. Dimensional analysis leads to various dimensionless parameters, which are based on the dimension’s Force (F), Mass (M), Length (L), Temperature (T), time (t) and Heat (Q). Based on these elements, one can obtain various independent parameters such as density (ρ), viscosity (µ), speed (N), diameter (D), and velocity (V), as shown in Table 2-1. The independent parameters lead to forming various dimensionless groups, which are used in the fluid mechanics of turbomachines. The specific speed compares the head and flow rate in geometrically similar machines at various speeds.

Ns =

N Q H

3 4

(2-1)

where H is the adiabiatic head, Q is the volume rate, and N the speed. The specific diameter compares head and flow rates in geometrically similar machines at various diameters. 1

DH 4 Ds = Q

(2-2)

The flow coefficient is the capacity of the flow rate expressed in dimensionless form.

φ=

Q ND 3

(2-3)

Many compressor applications, especially centrifugal compressors, are also given by the following relationship:

φ = Vm U2

(2-4)

where Vm is the meridional value and U2 is the tip speed of the impeller.

| 38 |

| Aerothermodynamics of Compressors |

DIMENSIONAL SYSTEM Mass (m) Force (f) Length (L) Time (t) Temperature (T) Work (w) Heat (q) Volume Velocity (v) Acceleration (a) Frequency (N) Area (a) Coefficient of thermal expansion (β) Density (ρ) Dimensional constant (gc) Specific heat at constant pressure (cp) at constant volume (cv) Heat transfer coefficient (h) overall (U) Work rate (W) Heat flow rate (q) Kinematic viscosity (v) Mass flow rate (m) Pressure (p) Angular velocity (ω) Volume flow rate Thermal conductivity (k) Thermal Diffusivity Viscosity, absolute (µ)

FMLTtQ*

BRITISH ENGINEERING UNITS

METRIC ENGINEERING UNITS

lbm lbf ft sec °R lbf-ft Btu ft3 ft/sec ft/sec2 1/sec ft2

kg newton m sec °K dyne-cm kJ m3 m/sec cm/sec2 1/sec cm2

1/°R

1/°K

M F L t T FL Q L3 Lt -1 Lt -2 t -1 t2 1 T M T ML T

lbm /ft3

kg/m3

lbm /lfbfsec2

m/sec2

_Q_ MT

Btu/lbm°R

kJ/kg°K

_Q_ tL2T

Btu/sec ft2°R

kJ/sec m2°K

FL t FL t L2 t M t F t t -1 L3t -1 Q t 2 L t -1 M Lt

ft lbf /sec

kJ/sec

Btu/sec

kJ/sec

ft2/sec

cm2/sec

lb/sec

kg/sec

lbf /ft2

bar

rad/sec ft 3/sec

rad/sec m3/sec

ft/lbf /ft-sec-°R

dyne-cm/cm-sec-°K

ft 2/sec

m2/sec

lbm /sec-ft

kg/sec-m

TABLE 2-1: Dimensions of Major Variables.

| 39 |

| Centrifugal Compressors | The pressure coefficient is the pressure or pressure rise expressed in dimensionless form ϕ=

H N2D2

(2-5)

For many applications, the pressure coefficient can be written as: P2 - P1 ρU 2 2

ϕ=

(2-6)

The temperature coefficient is given as θ=

T2 - T1 U2 2 gc p

(2-7)

The previous equations are some of the major dimensionless parameters. For the flow to remain dynamically similar, all the parameters must remain constant; however, constancy is not possible in a practical sense, so one much make choices. Reducing the Momentum Equation (Navier-Stokes) to non-dimensional form obtained the Reynolds Number and the Froude Number. Reducing the Energy Equation of a viscous compressible flow to a nondimensional form introduces the non-dimensional parameters such as Mach Number, Reynolds Number, and the Prandtl Number. The significance of the Reynolds number and the Froude Number can be seen from the requirements of dynamic similarity of an incompressible viscous flow. The Reynolds number is the ratio of the inertia forces to the viscous forces. Re =

rVD ν

(2-8)

Where ρ is the density of the gas, V the velocity, D the diameter of the impeller, and v the viscosity of the gas. The significance of the Reynolds number can be seen in the requirement of similarity of flows of an incompressible viscous fluid, where not only the bodies must be geometrically similar but also the Reynolds number in the two flow patterns must be equal in magnitude. This is a major requirement in the testing of Compressors.

| 40 |

| Aerothermodynamics of Compressors | The Froude Number is a very useful parameter in open channel flow when the flow is mainly due to gravity; it is the ratio of the dynamic force to the gravitational force: 2

Fr = V gL

(2-9)

where g is the gravitational force and L is the length In the case of flows in similar bodies in various compressible fluids, there are three additional similarity parameters: Mach Number, ratio of specific heats, and the Prandtl Number. The Mach Number is the ratio of velocity to the acoustic speed (a) of a gas at a given temperature M = V/a

(2-10)

The ratio of the specific heats is cp γ =c v

(2-11)

The Prandtl Number is a ratio of the momentum diffusivity or kinematic viscosity to the mass thermal diffusivity. Pr =

Kinematic viscosity Thermal diffusivity

=

µ / ρp γ / ρc p

=

µ cp γ

(2-12)

For a complete similarity of flows of viscous compressible fluids, the flows must have the same Reynolds Number, Mach Number, Froude Number, Prandtl Number, and the ratio of specific heats. Flow coefficients and pressure coefficients can be used to determine various off-design characteristics. The Reynolds number affects the flow calculations for skin friction and velocity distribution. When using dimensional analysis in computing or predicting performance-based tests performed on smaller-scale units, it is not physically possible to keep all parameters constant. The variations of the final results depend on the scale-up factor and the difference in the fluid medium. It is important in any type of dimensionless study to understand the limit of the parameters and that the geometrical scaleup of similar parameters must remain constant. Many scale-ups have developed major problems because stress, vibration, and other dynamic factors were not considered.

| 41 |

| Centrifugal Compressors | NOZZLES AND DIFFUSERS It is in the nozzle that a change in enthalpy is converted into a change in kinetic energy, and in the diffuser, a change in kinetic energy is converted into a pressure head. In the nozzle the velocity is increased and the static pressure and temperature are reduced, while in the diffuser the velocity is decreased and the static pressure and temperature are increased. There are no changes in total pressure and temperature in the nozzle or diffuser. The two equations of some importance in steady state flow-through nozzles are the energy equation and the continuity equation and these are shown below. For a stationary nozzle The Energy Equation: h1 +

V2 V12 = h2 + 2 2g c J 2g c J

(2-13)

and The Continuity Equation: . m = A1 V1 ρ1 = A2 V2 ρ2

(2-14)

where A V ρ J gc h

= Area = Velocity = Density = Mechanical equivalent of heat = Gravitational constant = Enthalpy . m = Mass flow the flow per unit area can be written as follows: . m γ P M = γ +1 (2-15) A R T γ - 1 2 2(γ - 1) (1+ M ) 2 where the Mach Number (M ) is defined as: V (2-16) M = a It is important to note that the Mach Number is based on Static Temperature.

| 42 |

| Aerothermodynamics of Compressors | The acoustic velocity (a) in a gas is defined by the following relationship:

a2 ≡

∂P ∂ρ

) s= c

(2-17)

For an adiabatic process (s = entropy = constant), the acoustic speed can be written as follows:

a=

γ g c RTs MW

(2-18)

where Ts = static Temperature cp For an isentropic adiabatic process; γ = c v where cp and cv are the specific heats of the gas at constant pressure and volume respectively and can be written as: c p - cv = R

(2-19)

where

cp =

γR γ -1

and

cv = R γ -1

(2-20)

It is important to note that the pressure measured can be either total or static, however, only total temperature can be measured. The relationship between total and static conditions for pressure and temperature are as follows: 2 (2-21) To = Ts + V 2c p where Ts = static temperature, and V= gas stream velocity and 2 Po = Ps + ρ V 2 gc

(2-22)

Equations 2-17, and 2-18 can be written in terms of the Mach Number as follows:

To γ-1 2 = (1 + M ) Ts 2 and

(2-23)

γ

Po  γ - 1 2  γ - 1 = 1+ M  Ps  2 

(2-24)

| 43 |

| Centrifugal Compressors | The point in the nozzle where the minimum area occurs and the Mach Number = 1 is known as the throat. This reduces the above relationships as follows: . 1 m γ P = (2-25) γ +1 A R T γ - 1 2(γ - 1) (1 + ) 2

T* = To ( γ 2+1 ) γ

(2-26)

P* = Po ( γ 2+1 ) γ - 1 where γ = 1.329 for steam γ = 1.12 for wet steam γ = 1.3 for superheated steam γ = 1.4 for air at 60°F

(2-27)

To reach the Mach number above, M=1.0, a convergent-divergent nozzle would have to be used. The speed of sound (acoustic velocity) results from three-dimensional effects, and accurate results occur from the one-dimensional effect of small disturbances. Large disturbances such as shock waves propagate at a much higher velocity. These small disturbances, which are pressure waves, propagate in a gaseous medium, in this case superheated steam. Figure 2-1 shows the aero-thermal properties of the flow in a convergent-divergent nozzle. The flow leaving the nozzle is supersonic. A convergent-divergent nozzle is designed to handle an expansion between certain expansion states. If the backpressure of the discharge region is less than the design pressure at the discharge boundary, an underexpansion occurs. A free expansion occurs after the steam leaves the nozzle, and since the steam is supersonic, an expansion wave will occur. If the backpressure of the discharge region is higher than the design pressure at the discharge boundary, an over-expansion occurs. A standing shock wave occurs in the divergent area of the nozzle. Across this shock there is a sharp irreversible increase in pressure, an increase in entropy, and a decrease in velocity. The convergent-divergent nozzle may have any cross sectional shape to fit the application. The elements of the surface of the divergent nozzle are generally straight for ease of manufacture of nozzles used in steam turbines. The nozzle is proportioned after the nozzle throat area is determined. The throat area is the smallest area in the nozzle and is where the

| 44 |

| Aerothermodynamics of Compressors | flow reaches a Mach Number equal to one. The flare of the sides of the divergent section should be within good fluid dynamic limits, so that no separation of the flow occurs, which is an inclusive angle between 12°-15°.

Figure 2-1: An ideal convergent divergent nozzle.

TURBOMACHINERY The turbocompressors discussed in this section transfer energy by dynamic means from a rotating member to the continuously flowing fluid. The four basic governing relationships in any turbomachine are the following equations: •

Equation of State

•

Energy Equation

•

Continuity Equation

•

Momentum Equation

| 45 |

| Centrifugal Compressors | In this chapter these four equations are examined and relationships obtained, which would be applicable to the thermodynamic and fluid mechanic relationships in centrifugal compressors: E Q U AT I O N O F S TAT E P / ρ/ζγ = Constant where

(2-28)

P = pressure ρ = Density cp γ = for an isentropic adiabatic process where __ cp are the specific heats of the gas at constant pressure and volume respectively and can be written as: cp - cv = R (2-29)

where cp =

γR γ -1

and

cv = R γ -1

(2-30)

IDEAL GAS Ideal gas obeys the equation of state PV = MRT or P/ρ = MRT, where P denotes the pressures, V the volume, ρ the density, M the mass, T the temperature of the gas, and R the gas constant per unit of mass independent of pressure and temperature. In most cases, the ideal gas laws are sufficient to describe the flow within 5% of actual conditions. When the perfect gas compressibility factor Z can be introduced:

Z (P, T ) =

PV RT

(2-31)

Figure 2-2 shows the relationship between the compressibility factor and pressure and temperature, couched in terms of reduced pressure and temperature:

Pr =

P , Pc

Tr =

T Tc

(2-32)

Pc and Tc are the pressure and temperature of the gas at the critical point.

| 46 |

| Aerothermodynamics of Compressors |

Figure 2-2: Compressibility factor chart for a simple fluid.

Static pressure is the pressure of the moving fluid. The static pressure of a gas is the same in all directions and is a scalar point function. It can be measured by drilling a hole in the pipe and keeping a probe flush with the pipe wall. Total pressure is the pressure of the gas brought to rest in a reversible adiabatic manner. It can be measured by a pitot tube placed in the flow stream. The gas is brought to rest at the probe tip. The relationship between total and static pressure is given in the following relationship: Pt = Ps +

ρV 2 2 gc

(2-33)

Where ρV 2/2gc is the dynamic pressure head that denotes the velocity of the moving gas.

| 47 |

| Centrifugal Compressors | Static temperature is the temperature of the flowing gas. This temperature rises because of the random motion of the fluid molecules. The static temperature can only be measured by a measurement at rest relative to the moving gas. The measurement of the static temperature is a difficult, if not impossible, task. Total temperature is the temperature rise in the gas if its velocity is brought to rest in a reversible adiabatic manner. Total temperature can be measured by the insertion of a thermocouple, RTD, or thermometer in the fluid stream. The relationship between the total temperature and static temperature can be given: Tt = Ts +

V2 2cp gc

(2-34)

COMPRESSIBILITY EFFECT The effect of compressibility is important in high Mach number machines. Mach number is the ratio of velocity to the acoustic speed of a gas at a given temperature M = V/a. Acoustic speed is defined as the ratio change in pressure of the gas with respect to its density if the entropy is held constant:  ∂P  a 2 ≡    ∂ρ  s=c

(2-35)

With incompressible fluids, the value of the acoustic speed tends toward infinity. For isentropic flow, the equation of state for a perfect gas can be written: P/ργ = constant Therefore, ln P - γ ln ρ = constant (2-36) Differentiating the previous equation, the following relationship is obtained: dP dρ (2-37) -γ =0 ρ P For an isentropic flow, the acoustic speed can be written: (2-38) a2 = dP/dρ Therefore, a2 = γP/ρ (2-39) Substituting the general equation of state and the definition of the acoustic velocity, the following equation is obtained: a2 = γgc R Ts (2-40)

| 48 |

| Aerothermodynamics of Compressors | Where TS (static temperature) is the temperature of the moving gas stream. Since the static temperature cannot be measured, the value of static temperature must be computed using the measurements of static pressure and total pressure and temperature. The relationship between static temperature and total temperature is given by the following relationship: Tt V2 =1+ Ts 2 g c c pTs

(2-41)

where the specific heat cp at constant volume can be written:

γR γ -1 and where γ is the ratio of the specific heats cp =

γ =

cp cv

(2-42)

(2-43)

Combining Equations (2-29) and (2-30) gives the following relationship:

Tt γ -1 2 M =1+ Ts 2

(2-44)

The relationship between the total and static conditions is isentropic, therefore:

Tt  Pt  =  Ts  Ps 

γ-1 γ

(2-45)

and the relationship between total pressure and static pressure can be written: Pt  γ - 1 2  M  = 1 + Ps  2 

γ γ-1

(2-46)

By measuring the total and static pressure, and using equation (2-46), the Mach number can be calculated. The static temperature can be computed using equation (2-44) since the total temperature can be measured. Finally, using the definition of Mach number, the velocity of the gas stream can be calculated.

| 49 |

| Centrifugal Compressors | T H E E N E R G Y E Q U AT I O N The energy equation is the mathematical formulation of the law of conservation of energy, which is the First Law of Thermodynamics. It states that the rate at which energy enters the volume of a moving fluid is equal to the rate at which work is done on the surroundings by the fluid within the volume and the rate at which energy increases within the moving fluid. The energy in a moving fluid is composed of internal, flow, kinetic, and potential energy h1 +

V2 V12 +1 Q 2 = h2 + 2 +1W2 2gc J 2gc J

(2-47)

where h V W Q

= enthalpy = Absolute Velocity = Work = Heat rejection

Neglecting the Heat Rejection, Equation 2-32 can be rewritten as: W2 = h 1 +

1

 V12 V2  − h2 + 2  = H 01 − H 02 2g c J  2gc J 

(2-48)

In the above equation, the negative sign indicates that work is being inputted into the system (compressor), and the positive sign indicates that the work is being extracted from the system (turbine). The work per unit mass for a centrifugal compressor is also known as the adiabatic head (Had ) thus rewriting equation (2-48). H ad

γ −1 γ −1       P02  γ  γRT01   P02  γ      = c p (T02 − T01 ) = c p    − 1 =    − 1   P01   γ − 1   P01      

(2-49)

T H E C O N T I N U I T Y E Q U AT I O N The continuity equation is a mathematical formulation of the law of conservation of mass of a gas that is a continuum. The law of conservation of mass states that the mass of a volume moving with the fluid remains unchanged . (2-50) m = A1V1 ρ1 = A 2 V2 ρ2

| 50 |

| Aerothermodynamics of Compressors | where A V ρ J Gc h . m

= area = Velocity = Density = Mechanical equivalent of heat = Gravitational constant = Enthalpy = Mass flow

the flow per unit area can be written as follows: where the Mach number (M ) is defined as:

. m = A

M =

γ P R T

M γ +1

γ − 1 2 2 (γ −1) (1 + M ) 2

V a

(2-51)

(2-52)

It is important to note that the Mach number is based on static temperature. The acoustic velocity (a) in a gas is given by the following relationship

a2 =

∂P ∂ρ

)s = c

(2-53)

for an adiabatic process (s = entropy = constant) the acoustic speed can be written as follows:

a=

γ gc RTs

(2-54)

MW

where Ts = static Temperature Total conditions occur when the flow is brought to rest in a reversible adiabatic manner and the static conditions are the conditions of flow in a moving stream. It is important to note that the pressure measured can be either total or static, however, only total temperature can be measured. For the total

| 51 |

| Centrifugal Compressors | conditions of pressure and temperature to change, the energy must be added or extracted to the fluid stream. The relationship between total and static conditions for pressure and temperature are as follows: 2

To = Ts + V 2c p

(2-55)

where To = total temperature, Ts = static temperature, and V= gas stream velocity

and Po = Ps + ρ

V2 2 gc

(2-56)

where Po = total pressure, Ps = static pressure Equations 2-12, and 2-13 can be written in terms of the Mach number as follows:

To γ −1 2 = (1 + M ) Ts 2

(2-57)

and γ

Po  γ − 1 2  γ −1 = 1+ M  Ps  2 

(2-58)

To understand the flow in a turbomachine, the concepts of absolute and relative velocity must be grasped. Absolute velocity (V ) is gas velocity with respect to a stationary coordinate system. Relative velocity (W) is the velocity relative to the rotor. In turbomachinery, the air entering the rotor will have a relative velocity component parallel to the rotor blade, and an absolute velocity component parallel to the stationary blades. Mathematically, this relationship is written:

V =U+ W

(2-59)

where the absolute velocity (V ) is the algebraic addition of the relative velocity (W ) and the linear rotor velocity (U ). The absolute velocity can be resolved into its components, the radial or meridional velocity (Vm) and the tangential component Vθ. From Figure 2-3 the following relationships are obtained:

| 52 |

| Aerothermodynamics of Compressors | V12 = Vθ12 + Vm12 V22 = Vθ22 + Vm22 W12 = (U1 - Vθ1)2 - Vm12 W22 = (U2 - Vθ2)2 + Vm22

(2-60)

T H E M O M E N T U M E Q U AT I O N The Navier-Stokes equation in its simplified form is the momentum equation, which is a mathematical formulation of the law of conservation of momentum. It states that the rate of change in linear momentum of a volume moving with a fluid is equal to the surface forces and body forces acting on a fluid. Figure 2-3 shows the velocity components in a generalized turbomachine. The velocity vectors as shown are resolved into three mutually perpendicular components: the axial component (Vθ), the tangential component (Vθ), and the radial component (Vm ).

Figure 2-3: Flow characteristics in a centrifugal compressor.

| 53 |

| Centrifugal Compressors | By examining each of these velocities, the following characteristics can be noted: the change in the magnitude of the axial velocity gives rise to an axial force, which is taken up by a thrust bearing; the change in radial velocity gives rise to a radial force, which is taken up by the journal bearing. The tangential component is the only component that causes a force, which corresponds to a change in angular momentum; the other two velocity components have no effect on this force—except for what bearing friction may arise. By applying the conservation of momentum principle, the change in angular momentum obtained by the change in the tangential velocity is equal to the summation of all the forces applied on the rotor. This summation is the net torque of the rotor. A certain mass of fluid enters the turbomachine with an initial velocity Vθ1, at a radius rθ1, and leaves with a tangential velocity Vθ2 at a radius r2. Assuming that the mass flow rate through the turbomachine remains unchanged, the torque exerted by the change in angular velocity can be written: . τ = m(r1Vθ 1 − r2Vθ 2 )

(2-61)

The rate of change of energy transfer (ft-lbƒ /sec) is the product of the torque and the angular velocity (ω) . τϖ = m(r1ϖVθ1 − r2ϖVθ 2 )

(2-62)

Thus, the total energy transfer can be written: . E = m(U1Vθ 1 − U 2Vθ 2 )

(2-63)

where U1 and U2 are the linear velocity of the rotor at the respective radii. The previous relation per unit mass flow can be written as the Adiabatic Head: Had = (U1Vθ 1 − U 2Vθ 2 )

(2-64)

where Had is the energy transfer per unit mass flow ft-lbƒ /lbm or fluid pressure. Equation (2-64) is known as the Euler turbine equation. The equation of motion as given in terms of angular momentum can be transformed into other forms, which are more convenient to understanding some of the basic design components. To understand the flow in a turbomachine, the concepts of absolute and relative velocity described

| 54 |

| Aerothermodynamics of Compressors | earlier must be well understood. In turbomachinery the air entering the rotor will have a relative velocity component parallel to the rotor blade and an absolute velocity component parallel to the stationary blades. By placing these relationships into the Euler turbine equation, the following relationship is obtained: Combining equation 2-49 with equation 2-34

H ad =

1 2g c

[(V

1

2

)

]

− V2 + (U 1 − U 2 ) + (W1 − W2 ) 2

2

2

γ

P02 γ −1 =[ (U 1Vθ 1 − U 2Vθ 2 ) + 1]γ −1 P01 γ RT01

2

2

(2-65)

(2-66)

As stated earlier, a centrifugal compressor operates on the principle of putting work into the incoming air by acceleration and diffusion. Air enters the impeller axially, as shown in Figure 2-3 with an absolute velocity (V1) and an angle α1, which combines vectorially with the tangential velocity of the impeller (U) to produce the resultant relative velocity W1 at an angle α1. Air flowing through the passages formed by the impeller is turned from the axial direction to the radial direction given a relative velocity W2 exiting the impeller radially. Note that W2 is less than W1 due to diffusion of the flow in the impeller. This is the result of an increase in the passage area due to the fact that the impeller exit has a much larger diameter at the exit compared to the inlet. The combination of the relative exit velocity and blade velocity produce an absolute velocity V2 at the exit of the rotor. The flow in this section of the impeller enters from the inducer section and leaves the impeller in the radial direction. There are three impeller vane types at the exit, as shown in Figure 2-4. These are defined according to the exit blade angles. Impellers with exit blade angle β2 = 90° are radial vanes. Impellers with β2 < 90° are backward-curved or backward-swept vanes, and for β2 > 90°, the vanes are forward-curved or forward-swept. They have different characteristics of theoretical head-flow relationship to each other, as shown in Figure 24. Although in Figure 2-4 the forward-curved head is the largest, in actual practice the head characteristics of all the impellers are similar to the backward-curved impeller. The air then passes through the diffuser, which can have a number of sectors. The conversion of the velocity head to a pressure head occurs in the diffuser, as in the case of the forward curve

| 55 |

| Centrifugal Compressors | blades when the exit velocity is very high, and the accompanying losses are also high. Thus, forward curved blades are not used in centrifugal compressors.

Forward Curved Blades

Head

Radial Blades

Backward Curved Blades

Flow Figure 2-4:Theoretical head as a function of flow for various types of blade angle at the impeller exit.

DEGREE OF REACTION The degree of reaction in a compressor is defined as the ratio of the change of static head in the rotor to the head generated in the stage: R=

H rotor Hstage

(2-67)

The change in static head in the rotor is equal to the change in relative kinetic energy:

H rotor = and

| 56 |

1 2 2 (W2 − W1 ) 2g

(2-68)

| Aerothermodynamics of Compressors | Hstage = U 1Vθ 1 − U 2Vθ 2

(2-69)

Thus, the reaction of the stage can be written R = (W2 − W12 ) /2 (U 1Vθ 1 − U 2Vθ 2 ) 2

(2-70)

In the case of axial entry, and radial exit Vθ 1 = 0 Vθ 2 = U 2 Vm = V1 = W2 W1 = Vm /sin α

where: Vm = Meridional Velocity H stage = U 22

(2-71)

H = Vm (1 − 1 /sin2 α ) /2U 22

(2-72)

2

When designing a compressor with this type of blading, inlet guide vanes to provide zero-prewhirl and the correct velocity entrance angle to the first-stage impeller must precede the first stage. The relative velocity between the entrance to the impeller and the exit is greatly reduced. Thus, higher blade speeds and axial-velocity components are possible without exceeding the limiting value of 0.70-0.75 for the inlet Mach number. Higher blade speeds result in impellers of smaller diameter and less weight. Gases in a turbomachinery, for the most part, are considered to be newtonian fluids. A newtonian fluid behaves like a perfectly elastic material and is given by the following linear relationship of the shear stress (τ):

τ =µ

∂u ∂y

(2-73)

where u is the velocity of the gas parallel to the boundary in the x-direction. Non-newtonian flow is experienced in centrifugal compressors if there is liquid entrainment in the flow.

| 57 |

| Centrifugal Compressors | EFFICIENCIES A D I A B AT I C E F F I C I E N C Y The work in a compressor under ideal conditions occurs at constant entropy, as shown in Figure 2-5. The dotted line indicates the actual work done. The isentropic efficiency of the compressor can be written in terms of the total changes in enthalpy: η ad =

Isentropic Work Actual Work

=

(h1t − h2t )ideal (h1t − h2’ t ) actual

(2-74)

This equation can be rewritten for a thermally and calorically perfect gas in terms of total pressure and temperature as follows:

η ad

 P =  2 t  P1t 

  

γ −1 γ

 T  − 1 /  2t − 1   T1t  

(2-75)

The process between 1 and 2' is defined as a polytropic process by the following equation of state:

P = const ρn

(2-76)

Where n is a polytropic process. The adiabatic efficiency can then be represented by: η ad

γ −1 n−1     γ n     P P    2t 2t     −1 /  =  − 1  P1t    P1t         

(2-77)

P2t P2s

enthalpy

P1t P1s

entropy entropy

Figure 2-5: Enthalpy-entropy diagram for a compressor.

| 58 |

| Aerothermodynamics of Compressors | P O LY T R O P I C E F F I C I E N C Y Polytropic efficiency is another concept of efficiency often used in compressor evaluation. It is often referred to as small stage or infinitesimal stage efficiency. It is the true aerodynamic efficiency exclusive of the pressure-ratio effect. The efficiency is the same as if the fluid is incompressible and identical with the hydraulic efficiency. Rewriting equation (2-77) by assuming that P2t can be substituted by P1t+dP the following relationship is obtained: γ −1

 dP  γ 1 +  −1 P1t   ηp = n−1  dP  n 1 +  −1  P1t 

which can be expanded in a Taylor’s series, assuming that

γ −1 γ ηp = n−1 n

(2-78)

 dP  6000]

| 81 |

| Centrifugal Compressors | Unit Width = 2.5Q0.35

[Minimum Width = 50"]

The flange sizes Inlet Flange = 0.84Q0.35 Outlet Flange = 0.7Q0.35 The horsepower required, assuming about 70 percent efficiency overall, gives the following relationships: HP = 4.3 x 10-5 m. Had The discharge temperature is given by the following relationship, where ηad is the adiabatic efficiency: γ −1    P2  γ  T1  − 1 P  1    ∆T =  η ad

(3-7)

Thus, discharge temperature: T2 = Tm + ∆T If intercooling is to be done between stages or sections, these computations must be repeated for each section. The cost of this unit will vary depending on the type of load, material of the casing, and whether or not the casing is horizontally split or a barrel type. Approximate costs can be equated as a function of the flow and head as given by the relationship given here: __ Cost ($) = 90Q0.2 √Had For barrel type units, the cost is nearly double. For other than cast iron casing, a premium of about 30% is to be added to the base figure. These are just ballpark estimates; other parameters must be considered, such as type of lubrication system, couplings, coolers, etc. to come up with the final figures.

| 82 |

| Compressor Definition and Standards | COMPRESSOR CASINGS Most compressor manufacturers have standard castings or forgings of their various housings and impellers from which they make up a compressor for a particular set of operating conditions. The housing size limits the flow, which can be passed through it while the impeller size restricts the head that can be obtained. The diameter in most centrifugal compressors is directly proportional to the head, thus the impeller castings limit the maximum and minimum head. Therefore, in most cases, the manufacturer has large impeller castings, which can be machined to various sizes for various heads. Figure 3-8 shows the limits due to the housing and impeller of a typical centrifugal compressor. Thus, with the use of such a map, the operational characteristics of the unit are identified. Superimposing these lines on a gas turbine performance curve shows the impeller that can be used for various conditions.

Figure 3-8: Compressor map showing limitation of casings and impellers.

| 83 |

| Centrifugal Compressors | C O M P R E S S O R E VA L U AT I O N The first step in any application is defining the process, the prime mover driving these compressors, and the various auxiliaries needed to operate these Compressor trains in a very efficient and reliable manner. The more pertinent the information obtained during the evaluation of the proposal, the better the selection of equipment for the process. The following tables contain items the user should consider in his attempt to properly evaluate the bid. Table 3-1 lists the important points that must be supplied to the vendor for the entire compressor train, while the important points to consider in evaluating centrifugal compressors are listed in Table 3-2. These tables will enable the engineer to make a proper evaluation of each critical point and ensure that he is purchasing units of high reliability and efficiency. 1)

2)

3)

4)

5)

| 84 |

The Gas to Be Handled (Each Stream) • Composition by mol%, volume %, or weight %. To what extent does composition vary? • Corrosive effects. Limits to discharge temperature which may cause problems with the gas. Quantity to Be Handled for each Stage • Stage quantity and unit of measurement. • If by volume, show: ° Whether wet or dry. ° Pressure and temperature reference points. Inlet Conditions for each Stage • Barometer. • Pressure at compressor flange. • State whether gauge or absolute. • Temperature at compressor flange. • Relative humidity. • Ratio of specific heats. • Compressibility. Discharge Conditions • Pressure at compressor flange. • State whether gauge or absolute. • Compressibility. • State temperature reference. Interstage Conditions • Temperature difference between gas out of cooler and water into cooler. • Is there interstage removal or addition of gas? • Between what pressures may this be done? Advise permissible range. • If gas is removed, treated, and returned between stages, advise pressure loss. • What quantity change is involved? • If this changes gas composition, a resultant analysis (ratio of specific heats, relative humidity, and compressibility at specific interstage pressure and temperature) must be provided.

| Compressor Definition and Standards | 6)

Variable Conditions – State expected variation in intake conditions – pressure, temperature, relative humidity, MW, etc. – State expected variation in discharge pressure. – It is extremely important that changing conditions be related to each other. • If relative humidity varies from 50 to 100% and inlet temperature varies from 0 to 100°F, does 100% RH correspond with 100°F? • Variations in conditions are best shown in tabular form with all conditions included in each column. 7) Flow Diagram • Provide a schematic flow sheet showing controls involved. 8) Regulation • What is to be controlled – pressure, flow, or temperature? • Advise permissible variation in controlled item. • Is regulation manual or automatic? • If automatic, are operating devices and/or instruments to be included? • How many control steps are desired on a reciprocator? 9) Cooling Water – Temperature: maximum and minimum – Pressure at inlet and backpressure, if any. – Whether open or closed cooling system desired. – Source of water. – Fresh, stale, or brackish. – Silty or corrosive. 10) Driver • Specify type of driver. • Electric motor: type, current conditions, power factor, enclosure, service factor, temperature rise, ambient temperature. • Steam: inlet and exhaust pressure, inlet temperature and quality, importance of minimum water rate. • Fuel gas: gas analysis, available pressure, and low heating value of gas. • Geared: AGMA rating, if special. 11) General – Acceptability of petroleum lubricants? – Indoor or outdoor installation? – Floor space, special shape? Provide a sketch. – Soil character. – List accessories desired, and advise which are to be spared. – Pulsation dampeners or intake or discharge silencers to be supplied. 12) Specifications • Provide each bidder with three copies of any specification for the particular project. • Complete information enables all manufacturers to bid competitively on the same basis and assists the purchaser in evaluating bids. Table 3-1: Vendor requirements to be provided by the user for a compressor train.

| 85 |

| Centrifugal Compressors | • • • • • • • • • •

•

•

• • • • • • •

•

•

Number of stages Pressure ratio and mass flow (per casing) Type of gas seals (inner seal) and oil seals Type of bearings (radial) Bearing stiffness coefficients Types of thrust bearings (Tapered land, non-equalizing tilting pad, and Kingsbury) Thrust float Temperature for journal and thrust bearings (operating temperature) Critical speed diagram (Speed vs. bearing stiffness curve) Type of impeller ° Shroud or unshrouded ° Blading ° Attachment of blades to hub and shroud Attachment of impellers to shaft ° Shrink fit ° Key fit ° Other Campbell diagrams of impellers ° Number of blades (impellers) ° Number of blades (diffuser) ° Number of blades (guide vanes) Balance piston Balance planes (location ° How is it balanced (detail) Weight of rotors (assembled) ° Split casing ° Barrel Data on torsional vibration (Bending criticals) Alignment data Type of coupling between tandems Performance curves (separate casings) ° Surge margin ° Surge line ° Aerodynamic speed ° Efficiency Intercooling type of cooler ° Temperature drop ° Pressure drop ° Efficiency Horsepower curves

Table 3-2: Points to consider in a centrifugal compressor.

P L A N T L O C AT I O N A N D S I T E C O N F I G U R AT I O N The location of the plant is the principal determination of the type of compressor train best configured to meet its needs. The compressor train consists of the following major components:

| 86 |

| Compressor Definition and Standards | 1.

2.

3. 4.

The type of compressor a. Axial or Centrifugal Compressor i. Types of seals ii. Types of couplings iii. Types of bearings The types of drivers a. Gas turbine drives b. Steam turbine drives c. Electric motor drives The use of gears between the driver and the compressor The lubrication systems

DRIVER SELECTION The three main types of drives for centrifugal compressors are (1) steam turbines, (2) gas turbines, and (3) electric motors. Deciding which drive is best is not always easy. Selection depends on many factors, such as location, process, and unit size. For remote locations, gas turbines are used mostly due to their low maintenance and the ability to prepackage the units. Their light weight makes them a must for offshore platforms. Gas turbines also have an advantage because of their multi-fuel capabilities. In many cases this capability assures that the unit must be a gas turbine. For chemical plants, steam turbines are widely used due to the fact that most of these plants need steam or generate it in their processes. Thus, by using steam turbines the plant can utilize most of the excess steam and use energy more efficiently. In the combined cycle utilities plants, a combination of gas turbines and steam turbines are being utilized to obtain high overall efficiencies. In this configuration, the hot gases from the gas turbine are used to generate the steam for the steam turbine. For smaller flows, electric motors drive the compressor, usually through speed increasing gears. In many applications, electric motors are used in conjunction with steam or gas turbines as helpers or for startup conditions. Typical ranges for the various drives are shown in Figure 3-9, from which we note that the higher the flow, the lower the speed. At high flows, the compressor diameter must be large; therefore, the speed must be reduced to maintain the same stress levels in the machinery. Table 3-3 shows a comparison of different drivers as a function of the horsepower, speed, efficiency, and torque.

| 87 |

| Centrifugal Compressors |

Figure 3-9: Performance characteristics of different types of compressor drives.

Most of the prime movers used are very dependent on the location of the compressor train. The main reason is the size and the type of fuel available, as well as efficiency and, most important, reliability. Most offshore applications use gas turbine drives. These gas turbines are primarily aero-derivative types and are usually less than 30 MW. Frame type and industrial gas turbines are used in many applications onshore. The determination to have an aero-derivative type gas turbine or a frame type gas turbine is the plant location. In most cases, if the plant is located offshore on a platform, then an aero-derivative plant is required. In smaller plants (between 2-15 MW), the industrial type small turbines best suit the application. And in plants between 15-30 MW, both aero-derivative and frame types can apply. Aero-derivatives have lower maintenance and have high heat-recovery capabilities. In many cases, the type of fuel and service facilities may be the determination. Natural gas or diesel no. 2 would be suited for aero-derivative gas turbines, but heavy fuels would require a frame type gas turbine.

| 88 |

| Compressor Definition and Standards |

Driver

Available Horsepower Speed Range (60 Cycle)

Induction Motor

1 to 5000 or larger

Synchronous 100 to Motor 2000 or Larger

Steam Turbine

1 to 25,000 or Larger

Possible Speed Variation

3600/N N=1 thru 8

Constant Speed

3600/N N=2 thru 20

Constant Speed

34,000 to 1800

100% down to 30% 100% down to 50%

Efficiency

10 hp=86% 60 to 100% 150% min 100 hp=91% torque 1000 hp=94% 550 to 650% amperes 93% to 97% 40% torque 150% under 514 rpm 40% to 100% torque over 514 rpm 300 to 500% Amperes 30% to 175 to Up to 35% 300% 300%

Combustion Gas Turbine

300 hp @ 10,000 rpm 10,000 hp @ 3000 to 5000 pm

Integral Gas Engine

85 hp @ 600 rpm to 6000 hp @ 300 rpm

100% down to 60%

30% to 40%

Integral Diesel Engine

100 hp @ 600 rpm to 1500 hp @ 300 rpm

100% down to 60%

35%

Coupled Gas or Diesel Engine Steam Engine

100 hp @ 600 rpm to 6000 hp @ 300 rpm

100% down to 60%

35 to 41%

100% down to 20%

30 to 45%

55 to 4000 hp

150 to 500 rpm

Starting Torque and Stalling Amperes Torque % Full % Full Load Load

27% to 40%

Both single and split shaft require sizable starting motor or turbine. Single shaft design has poor part load torque characteristics and requires a larger starter. Two shaft turbines have good torque characteristics. Nil started About with 120% compressed air Nil started About with 120% compressed air Nil started About with 120% compressed air About About 120% 115%

Table 3-3: Driver selection.

| 89 |

| Centrifugal Compressors | Many plants use steam turbine drives, especially in onshore locations and in plants where steam is generated for process needs. There are two major classification types of steam turbines: 1. Back pressure steam turbine 2. Condensing steam turbines In most plants the compressor drives are mostly backpressure steam turbines because of the large number of drives needed, the compactness of the back pressure steam turbine, and the complexity of condensing the steam. On larger steam turbines, usually over 15 MW, the steam turbines are condensing and may also be extraction types. Electric motor drives are also used widely in onshore applications. In fact, there has been a trend to replace a number of small backpressure steam turbines (less than 5 MW) with synchronous motor drives. Care must be taken to ensure that, when the shift from steam turbines to electric motor drives is undertaken, a detailed computation be considered of the torsional effects on the drive shafts. TYPE OF FUEL The type of fuel is one of the most important aspects that govern the selection of a drive for a compressor train. Natural gas would be the choice of most operators if natural gas were available, since its effects on pollution is minimal, maintenance cost would also be the lowest, and it is used in many petrochemical facilities for feedstock. Diesel fuel is often used in many applications. Many turbines also use refinery gas and other low calorific gases. Aero-derivative gas turbines cannot operate on heavy fuels, thus if heavy fuels were a criteria, then the frame type turbines would have to be used. TYPES OF CONDENSERS Large steam turbines are usually condensing type steam turbines. Condensers can be water-cooled or air-cooled units. Water-cooled condensers are the most prevalent and have a higher degree of effectiveness. The cooling effectiveness of the condenser is very important to the performance of the low-pressure steam turbine. If the cooling effectiveness is reduced, then the backpressure of the low-pressure steam turbine is increased, and the power output of the steam turbine is reduced.

| 90 |

| Compressor Definition and Standards | Steam turbines traditionally have down (or bottom) exhaust, where the condenser is located below the steam turbine. In such an arrangement, the steam pedestal is raised about 30 ft. (10meters) above the base slab where the condenser sits. Care must be taken that none of the condensate splashes onto the turbine casing otherwise casing distortion and cracks can occur. ENCLOSURES Gas turbines usually come packaged in their own enclosures. These enclosures are designed so that they limit the noise to 70dB at 100 ft (30 meters) from the gas turbine. The entire plant, consisting of the gas turbine and the centrifugal compressor, can be either inside or outside. While open plants are less expensive than enclosed plants, some owners prefer to enclose their trains in a building and use permanent cranes for maintenance. Typically, steam turbines are not in their own enclosure. They are usually open and are connected to the centrifugal compressor. Electric motors usually come packaged in their own enclosures. These enclosures are designed so that they limit the noise to 70dB at a 100 ft (30 meters) from the drive. The motors are connected through a gearbox to the centrifugal compressor, since the electric motors are usually at speeds of 3600 rpm for a 60-cycle drive and 3000 rpm for a 50cycle drive, and the centrifugal compressors are operated at a much higher speed. P L A N T O P E R AT I O N M O D E Most of these compressor trains operate at full load at a single speed for 24 hours a day, 7 days a week, 365 days a year. Plant reliability is very important as these trains are the heart of processes. In most plants, the shutting down of a train would create substantial losses, and this would mean shutting down the whole process, which may take many days to re-stabilize. In applications where the compressor needs to be operated at different loads and different speeds, gas turbines and steam turbine drives have a special application. The gas turbine used for this type of application usually uses a separate power turbine, thus enabling it to operate at different loads and speeds without affecting the overall efficiency substantially.

| 91 |

| Centrifugal Compressors | S TA R T- U P T E C H N I Q U E S The start-up of a gas turbine is done by the use of electrical motors, diesel motors, and, in plants where there is an independent source of steam, by a steam turbine. In many cases these steam turbines or electrical motor drives are also used to supply additional power, once the turbine is above its sustaining speed. New turbines use the generator as a motor for start-up. After combustion occurs and the turbine reaches a certain speed, the motor declutches and becomes a generator. Use of a synchronous clutch between two rotating pieces of equipment is not new. It is very common in use with start-up equipment. Electrical motors go to full speed in a very rapid manner and thus apply very high torque to the shaft. Synchronous motor drives create a very high torsional force, and thus for torsional drives, it is recommended that a full torsional analysis be conducted. P E R F O R M A N C E S TA N D A R D S The purpose of the ASME Performance Test Codes is to provide standard directions and rules for the conduct and report of tests of specific equipment and the measurement of related phenomena. These codes provide explicit test procedures with accuracies consistent with current engineering knowledge and practice. The codes are applicable to the determination of performance of specific equipment. They are suitable for incorporation as part of commercial agreements to serve as a means to determine fulfillment of contract obligations. The parties to the test should agree to accept the code results as determined or, alternatively, agree to mutually acceptable limits of uncertainty established by prior agreement of the principal parties concerned. The performance tests must be run as much as possible to meet the ASME performance codes. These codes are very well written and fully delineate the tests required. Meetings should be held in advance with the vendors to decide which part of the code would not be valid and what assumptions and correction factors must be undertaken to meet the various power and efficiency guarantees. The determination of special data or verification of particular guarantees, which are outside of the scope of the codes, should be made only after written agreement of both parties to the test, especially regarding methods of measurement and computation, which should be completely described in the test report.

| 92 |

| Compressor Definition and Standards | ASME, Performance Test Code on Test Uncertainty: Instruments and Apparatus PTC 19.1, 1988 This test code specifies procedures for evaluation of uncertainties in individual test measurements, arising from both random errors and systematic errors, and for the propagation of random and systematic uncertainties into the uncertainty of test results. The various statistical terms involved are defined. The end result of a measurement uncertainty analysis is to provide numerical estimates of systematic uncertainties, random uncertainties, and the combination of these into a total uncertainty with an approximate confidence level. This is especially important when computing guarantees plant output and plant efficiency.

ASME, Performance Test Code on Compressors and Exhausters, ASME PTC 10 1997, American Society of Mechanical Engineers 1997 The object of this code is to provide a test procedure to determine the thermodynamic performance of an axial or centrifugal compressor on a specified gas of known properties, which may be determined during the compression process where, under specified conditions, no condensation or evaporation occurs and there is no injection of liquids. The code defines conditions and provides methods by which a compressor may be tested on a suitable test gas, and the results can be converted into anticipated performance of the same compressor when pumping the specific gas at design conditions. Also, guidelines are provided for testing compressors with inter-stage side stream inlets or outlets, internally cooled compressors, and un-cooled tandem-driven compressors with externally piped intercoolers. This code is written to provide explicit test procedures, which will yield the highest level of accuracy consistent with the best engineering knowledge and practice currently available. Nonetheless, no single universal value of the uncertainty is, or should be, expected to apply to every test. The uncertainty associated with any individual PTC 10 test will depend upon practical choices made in terms of instrumentation and methodology. Rules are provided to estimate the uncertainty for individual tests. The scope of this code includes instructions on test arrangement and instrumentation, test procedure, and methods for evaluation and reporting of final results.

| 93 |

| Centrifugal Compressors | Rules are provided for establishing the following quantities, corrected as necessary to represent expected performance under specified operating conditions with the specified gas: •

quantity of gas delivered

•

pressure rise produced

•

head

•

shaft power required

•

efficiency

•

surge point

•

choke point

Other than providing methods for calculating mechanical power losses, this code does not cover rotor dynamics or other mechanical performance parameters.

ASME, Performance Test Code on Gas Turbines, ASME PTC 22 1997, American Society of Mechanical Engineers 1997 The object of the code is to detail the test to determine the power output and thermal efficiency of the gas turbine when operating at the test conditions and to correct these test results to standard or specified operating and control conditions. Procedures for conducting the test, calculating the results, and making the corrections are defined. The code provides for the testing of gas turbines supplied with gaseous or liquid fuels (or solid fuels converted to liquid or gas prior to entrance to the gas turbine). Test of gas turbines with water or steam injection for emission control and/or power augmentation are included. The tests can be applied to gas turbines in combined-cycle power plants or with other heat recovery systems. Meetings should be held with all parties concerned as to how the test will be conducted and an uncertainty analysis should be performed prior to the test. The overall test uncertainty will vary because of the differences in the scope of supply, fuel(s) used, and driven equipment characteristics. The code establishes a limit for the uncertainty of each measurement required. The overall uncertainty is then calculated in accordance with the procedures defined in the code and by ASME PTC 19.1.

| 94 |

| Compressor Definition and Standards | ASME, Performance Test Code on Steam Turbines, ASME PTC 6 1996, The code provides procedures for the accurate testing of steam turbines. It is recommended for use in conducting acceptance tests of steam turbines and for any other situation in which performance levels must be determined with minimum uncertainty. The performance parameters addressed are: •

heat rate

•

generator output

•

steam flow

•

steam rate

•

feedwater flow

This code applies to steam turbines operating with a significant amount of superheat in the initial steam, such as the steam turbines in a combined cycle power plant. It also contains procedures and techniques required to determine enthalpy values within the moisture region and modifications necessary to permit testing This code contains rules and procedures for the conduct and reporting of steam turbine testing, including mandatory requirements for pretest arrangements, instruments to be employed, their application and methods of measurement, testing techniques, and methods of calculation of test results. This establishes a limit for the uncertainty of each measurement required. The overall uncertainty is then calculated in accordance with the procedures defined in the code and by ASME PTC 19.1.

ASME, Performance Test Code on Steam Condensing Apparatus, ASME PTC 12.2 1983, American Society of Mechanical Engineers 1983 This code provides rules for determining the performance of a condenser with regard to one or more of the following: •

The absolute pressure the apparatus will maintain at the steam inlet nozzle when transferring heat rejected by the prime mover at a given rate in Btu per hour, with a given flow and temperature of circulating water, and a given tube cleanliness factor.

| 95 |

| Centrifugal Compressors | •

The thermal transmittance of surface condensers for given operating conditions. A test method for determining, concurrent with other measurements, the degree of tube fouling, which is expressed as a cleanliness factor or fouling resistance, is described in the code.

•

The amount of undercooling of the condensate.

•

The amount of dissolved oxygen in the condensate.

Testing of the condenser, as per the code, does not require testing for the cleanliness factor. Parties to the test could agree to some assumed degree of cleanliness, after cleaning by a method agreed upon by all the parties involved. Failure to agree on this point prior to the test will require determination of the cleanliness factor.

ISO, Natural Gas—Calculation of Calorific Value, Density, and Relative Density, International Organization for Standardization ISO 6976-1983(E) This international standard specifies methods for the calculation of calorific value, both higher and lower heating values, density, and relative density of natural gas when values for the physical properties of the pure components and the composition of the gas by mole fraction are known. The standard also describes the determination of the precision of the calculated calorific value from the precision of the method of analysis.

Table of Physical Constants of Paraffin Hydrocarbons and other components of Natural Gas—Gas Producers Association Standard 2145-94 The objective of the standard is to provide the gas processing industry and natural gas users a convenient compilation of authoritative numerical values for the paraffin hydrocarbons and other compounds occurring in natural gas and natural gas liquids. The physical properties selected are those considered to be the most needed by engineers in analytical computations in gas plants and engines. M E C H A N I C A L PA R A M E T E R S The American Petroleum Institute (API), as part of their Mechanical Equipment Standards, has written some of the best standards from a

| 96 |

| Compressor Definition and Standards | mechanical point of view. These mechanical equipment standards are an aid in specifying and selecting equipment for general petrochemical use, however, they are general in nature and can be used as guidelines for power plants. The intent of these specifications is to facilitate the development of high-quality equipment with a high degree of safety and standardization. The user’s problems and experience in the field are considered in writing these specifications. The task force, which writes the specifications, consists of members from the user, the contractor, and the manufacturers. Thus, the task-force team brings together both experience and know-how, especially from the operation and maintenance point of view. The petroleum industry is one of the largest users of cogeneration power. Thus the specifications written are well suited for this industry, and the tips of operation and maintenance apply for all industries. This section deals with some of the applicable API and ASME standards for centrifugal compressors, gas turbine, steam turbine, and other various associated pieces of equipment as used in a centrifugal compressor train. The combination of the extensive ASME and API codes covers the centrifugal compressor train, which is the centerpiece of many petrochemical plants. The two main API standards, which govern centrifugal compressors in a petrochemical plant, are API 617, and API 672.

API Std 617, Centrifugal Compressors for Petroleum, Chemical and Gas Industry Services, 6th Edition, February 1995 This specification covers the minimum requirements for centrifugal compressors used in petroleum, chemical, and gas industry services that handle air or gas. It does not apply to fans or blowers that develop less than 34kPa (5 pounds per square inch) pressure rise above atmospheric pressure; these are covered by API Standard 673. This standard also does not apply to packaged, integrally geared centrifugal air compressors, which are covered by API Standard 672.

API Std 672, Packaged, Integrally Geared Centrifugal Air Compressors for Petroleum, Chemical, and Gas Industry Services, 3rd Edition, September 1996 This specification covers the air compressors for instrument air for the plant. The standard establishes the minimum requirements for constantspeed, packaged, integrally geared centrifugal air compressors, including their accessories. It may be applied for gas services, other than air that is

| 97 |

| Centrifugal Compressors | nonhazardous and nontoxic. This standard is not applicable to machines that develop a pressure rise of less than 5.0 psi (0.35 BAR) above atmospheric pressure.

The API standards, which govern the various Prime Movers are API 616 for Gas Turbines, API 612 and ISO 10436 for Steam Turbines, and NEMA MG 2 for Electric Motor drives API Std 616, Gas Turbines for the Petroleum, Chemical and Gas Industry Services, 4th Edition, August 1998 This standard covers the minimum requirements for open, simple and regenerative-cycle combustion gas turbine units for services of mechanical drive, generator drive, or process gas generation. All auxiliary equipment required for starting and controlling gas turbine units and for turbine protection is either discussed directly in this standard or referred to in this standard through references to other publications. Specifically, gas turbine units that are capable of continuous service firing gas or liquid fuel or both are covered by this standard. In conjunction with the API specifications, the following ASME codes also supply significant data in the proper selection of the gas turbine.

API Std 612, Special Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services 1997 This specification covers the minimum requirements for specialpurpose steam turbines. These requirements include basic design, materials, related lubrication systems, controls, auxiliary equipment, and accessories. Special-purpose turbines are horizontal turbines used to drive equipment that is relatively large in size and is in critical service. This category of steam turbines is not limited by temperature, pressure, or speed.

ISO 10436: 1993 Petroleum and Natural Gas Industries— General Purpose Steam Turbine for Refinery Service, 1st Edition. This specifies the minimum requirements for general-purpose steam turbines for use in petroleum refinery service. The requirements include basic design, materials, related lubrication systems, controls, auxiliary equipment, and accessories.

| 98 |

| Compressor Definition and Standards | NEMA MG 2 Safety Standard And Guide For Selection, Installation, And Use Of Electric Motors And Generators This standard covers the minimum requirements for electric motor drives for large compressors and pumps. The following are other mechanical standard governing the various drives.

ASME BASIC GAS TURBINES B 133.2 Published: 1977 (Reaffirmed year: 1997) This standard presents and describes features that are desirable for the user to specify in order to select a gas turbine that will yield satisfactory performance, availability, and reliability. The standard is limited to a consideration of the basic gas turbine including the compressor, combustion system, and turbine.

ASME GAS TURBINE FUELS B 133.7M Published: 1985 (Reaffirmed year: 1992) Gas turbines may be designed to burn either gaseous or liquid fuels or both, with or without changeover while under load. This standard covers both types of fuel.

ASME GAS TURBINE CONTROL AND PROTECTION SYSTEMS B133.4 Published: 1978 (Reaffirmed year: 1997) The intent of this standard is to cover the normal requirements of the majority of applications, recognizing that economic trade-offs and reliability implications may differ in some applications. The user may desire to add, delete, or modify the requirements in this standard to meet his specific needs, and he has the option of doing so in his own bid specification. The gas turbine control system shall include sequencing, control, protection, and operator information, which shall provide for orderly and safe start-up of gas turbine, control of proper loading, and an orderly shutdown procedure. It shall include an emergency shutdown capability, which can be operated automatically by suitable failure detectors or which can be operated manually. Coordination between gas turbine control and driven equipment must be provided for startup, operation, and shutdown.

| 99 |

| Centrifugal Compressors | ASME GAS TURBINE INSTALLATION SOUND EMISSIONS B133.8 Published: 1977 (Reaffirmed: 1989) This standard gives methods and procedures for specifying the sound emissions of gas turbine installations for industrial, pipeline, and utility applications. Included are practices for making field sound measurements and for reporting field data. Users and manufacturers can apply this standard to write specifications for procurement and to determine compliance with specification after installation. Information is included for guidance to determine expected community reaction to noise.

ASME MEASUREMENT OF EXHAUST EMISSIONS FROM STATIONARY GAS TURBINE ENGINES B133.9 Published: 1994 This standard provides guidance in the measurement of exhaust emissions for the emissions performance testing (source testing) of stationary gas turbines. Source testing is required to meet federal, state, and local environmental regulations. The standard is not intended for use in continuous emissions monitoring, although many of the online measurement methods defined may be used in both applications. This standard applies to engines that operate on natural gas and liquid distillate fuels. Much of this standard also will apply to engines operated on special fuels such as alcohol, coal gas, residual oil, or process gas or liquid. However, these methods may require modification or be supplemented to account for the measurement of exhaust components resulting from the use of a special fuel.

ASME PROCUREMENT STANDARD FOR GAS TURBINE ELECTRICAL EQUIPMENT B133.5 Published: 1978 (Reaffirmed year: 1997) The aim of this standard is to provide guidelines and criteria for specifying electrical equipment, other than controls, which may be supplied with a gas turbine. Much of the electrical equipment will apply only to larger generator drive installations, but, where applicable, this standard can be used for other gas turbine drives. Electrical equipment described here, in almost all cases, is covered by standards, guidelines, or recommended practices documented elsewhere. This standard is intended to supplement those references and point out the specific areas of interest for a gas turbine application. For a few of the individual items, no other standard is referenced for the entire subject, but, where applicable, a

| 100 |

| Compressor Definition and Standards | standard is referenced for a sub-item. A user is advised to employ this and other more detailed standards to improve his specification for a gas turbine installation. In addition, regulatory requirements such as OSHA and local codes should be considered in completing the final specification. Gas turbine electrical equipment covered by this standard is divided into four major areas: main power system, auxiliary power system, direct current system, and relaying. The main power system includes all electrical equipment from the generator neutral grounding connection up to the main power transformer or bus, but not including a main transformer or bus. The auxiliary power system is the gas turbine section AC supply and includes all equipment necessary to provide such station power as well as motors utilizing electrical power. The DC system includes the battery and charger only. Relaying is confined to electric system protective relaying that is used for protection of the gas turbine station itself.

ASME PROCUREMENT STANDARD FOR GAS TURBINE AUXILIARY EQUIPMENT B133.3 Published: 1981 (Reaffirmed year: 1994) The purpose of this standard is to provide guidance to facilitate the preparation of gas turbine procurement specifications. It is intended for use with gas turbines for industrial, marine, and electric power applications. The standard also covers auxiliary systems such as lubrication, cooling, fuel (but not its control), atomizing, starting, heatingventilating, fire protection, cleaning, inlet, exhaust, enclosures, couplings, gears, piping, mounting, painting, and water and steam injection.

API Std 613, Special Purpose Gear Units for Petroleum, Chemical and Gas Industry Services, 4th Edition, June 1995 Gears, wherever used, can be a major source of problem and downtime. This standard specifies the minimum requirements for specialpurpose, enclosed, precision, single- and double-helical, one- and twostage speed increasers and reducers of parallel-shaft design for refinery services. Primarily intended for gears that are in continuous service without installed spare equipment, these standards apply for gears used in the power industry.

| 101 |

| Centrifugal Compressors | API Std 677, General-Purpose Gear Units for Petroleum, Chemical, and Gas Industry Services, 2nd Edition, July 1997, Reaffirmed March 2000 This standard covers the minimum requirements for general-purpose, enclosed single- and multistage gear units that incorporate parallel-shaft helical and right angle spiral bevel gears for the petroleum, chemical, and gas industries. Gears manufactured according to this standard are limited to the following pitchline velocities: helical gears shall not exceed 12,000 feet per minute (60 meters per second) and spiral bevel gears shall not exceed 8,000 feet per minute (40 meters per second). This standard includes related lubricating systems, instrumentation, and other auxiliary equipment. Also included in this edition is new material related to gear inspection.

API Std 614, Lubrication, Shaft-Sealing, and Control-Oil Systems and Auxiliaries for Petroleum, Chemical and Gas Industry Services , 4th Edition, April 1999 Lubrication also provides cooling for various components of the turbine. This standard covers the minimum requirements for lubrication systems, oil-type shaft-sealing systems, and control-oil systems for special-purpose applications. Such systems may serve compressors, gears, pumps, and drivers. The standard includes the systems’ components, along with the required controls and instrumentation. Data sheets and typical schematics of both system components and complete systems are also provided. Chapters include General Requirements, Special Purpose Oil Systems, General Purpose Oil Systems and Dry Gas Seal Module Systems. This standard is well written and the tips detailed are good practices for all type of systems.

ANSI/API Std 610, Centrifugal Pumps for Petroleum, Heavy Duty Chemical, and Gas Industry Services, 8th Edition, August 1995 (-1995 ) The centrifugal pumps are used for boiler-feed water pumps and condensate pumps in a combined cycle power plant. The vertical pumps are also used for sump pumps and for water for cooling purposes in the condenser. The pump types covered by this standard can be broadly classified as overhung, between bearings, and vertically suspended.

| 102 |

| Compressor Definition and Standards | API Std 671, Special Purpose Couplings for Petroleum Chemical and Gas Industry Services, 3rd Edition, October 1998 This standard covers the minimum requirements for special purpose couplings intended to transmit power between the rotating shafts of two pieces of refinery equipment. These couplings are designed to accommodate parallel offset, angular misalignment, and axial displacement of the shafts without imposing excessive mechanical loading on the coupled equipment.

ANSI/API Std 670, Vibration, Axial-Position, and BearingTemperature Monitoring Systems, 3rd Edition, November 1993 This provides a purchase specification to facilitate the manufacture, procurement, installation, and testing of vibration, axial position, and bearing temperature monitoring systems for petroleum, chemical, and gas industry services. It covers the minimum requirements for monitoring radial shaft vibration, casing vibration, shaft axial position, and bearing temperatures. It outlines a standardized monitoring system and covers requirements for hardware (sensors and instruments), installation, testing, and arrangement. Standard 678 has been incorporated into this edition of Standard 670. This is a well-documented standard, and widely used in all industries. CENTRIFUGAL COMPRESSORS

API Standard 617 and 672 These standards are used by the petrochemical industry as the main standards used in specifications for driven centrifugal compressors. The concepts provide the reader with an outline of centrifugal compressor requirements in various applications. API Standard 617 was revised in February 1995 and is intended to cover centrifugal compressors, which develop more than five lbs/sq in. The API Standard 672 was revised in September 1996 and is intended to cover air compressors for instrument air. The definitions of various terminologies used in the field and in the standards are well defined. The design of the casings, inlet vanes, rotating elements, seals, bearings, and rotor dynamics are discussed.

| 103 |

| Centrifugal Compressors | The casings for most of the API 617 compressors should be horizontally (axially) split except in high-pressure compressors. Where the partial pressure of the H2 exceeds 200 psig, the casing should be vertically (radially) split. The calculation of the hydrogen partial pressure is given by the following relationship: Maximum casing working pressure (psig) for axially split compressors can be written as follows: Pcasing =

200 % H 2 / 100

(3-8)

The casings should include a one-eighth inch corrosion allowance and sufficient strength and rigidity to limit change of alignment to 0.002 inches if caused by the worst combination of pressure, torque, or allowable piping stress. The rotating elements consist of the impeller and the shaft. The shaft should be made of one-piece, heat-treated forged steel, with the shaft ends tapered for coupling fits. Interstage sleeves should be renewable and made of a material, which is corrosion resistant in the specified service. The rotor shaft sensing area observed by the noncontact probes should be concentric with the bearing journals and free from any scratches, marks, or any other surface discontinuity. The surface finish should be 16-32 micro inches root mean square, and the area should be demagnetized and treated. Electromechanical runout should not exceed 25% of the maximum allowed peak-to-peak vibration amplitude or 0.25 mils, whichever is greater. Though not mentioned in the standard, chrome plating of the shaft in the sensing area is unacceptable. Maximum vibration should not exceed 2.0 mils as given by the following relationship: Vibmax =

12000 12000 + .25 rpm rpm

(Vibration)

(3-9)

(runout)

At the trip speed of the driver (105% for a gas turbine), the vibration should not exceed this level by more than 0.5 mil. The impellers can be an open-face (stationary shroud) or closed-face (rotating shroud) design. As long as the tip velocities are below 1,000 ft/sec, closed-face impellers can be used. The standards allow the impellers to be welded, riveted, milled, or cast. Riveted impellers are unacceptable, especially if the impeller loading is high. Impellers are to be assembled on

| 104 |

| Compressor Definition and Standards | the shaft with a shrink fit with or without a key. Shrink fits should be carefully done because excessive shrink fits can cause a problem known as hysteresis whirl. In compressors where the impellers require their thrust to be balanced, a balance drum is acceptable and preferred. Seals of many types can be used and are detailed in the standard. Labyrinth seals may include carbon rings in addition to the labyrinth. Also, mechanical contact seals must be provided with labyrinths and slingers to minimize oil leakage. Buffer gas may be used to prevent contamination of the oil. The compressor must operate in a region away from any critical speed. The amplification factor used to indicate the severity of the critical speed is given by the relationship AF =

Critical Speed Peak Width of the “half – Power Point”

(3-10)

AF =

N C1 N 2 − N1

(3-11)

where (N2 – N1) are the rpm’s corresponding to the .707 peak critical amplitude. The amplification factor should be below eight and preferably below five. A rotor response plot is shown in Figure 3-10. The operational speed for units operating below their first critical speed should be at least 20% below the critical speed. For units operating above their first critical speed, the operational speed must be at least 15% above the critical speed and/or 20% below any higher critical speed. The preferred bearings for the various types of installation are tilting-shoe radial bearings and the selfequalizing tilting pad/thrust bearings. Radial and thrust bearings should be equipped with embedded temperature sensors to detect pad surface temperatures. Accessories such as couplings, gearboxes, lubrication systems, and controls are also dealt with in the codes. The codes also deal with the various tests and test requirements. The major tests are the hydrostatic test, the impeller over speed test, the mechanical test, and the performance test. The hydrostatic test consists of testing all pressure-containing parts to one-half times their maximum working pressure. The impeller over speed test subjects each impeller to 115% of maximum continuous speed for a minimum duration of one minute. The mechanical run test is conducted at 10% increments over the

| 105 |

| Centrifugal Compressors |

Figure 3-10: Rotor response plot. Figure 7 of Standard 617, Centrifugal Compressors for General Refinery Services, 4th Edition, 1979. (Courtesy of the American Petroleum Institute).

entire speed range. The compressor is run for at least 15 minutes at an overspeed of 110%, then the speed is reduced and the compressor run for four hours at maximum continuous speed. It is advisable for the user to witness these tests and tape the vibration signals. They may be used at a later date as base-line data. The performance test should be conducted per ASME Power Test Code 10. The standards call for a minimum of five points, including surge and overload. This minimum is not adequate, and the user should ask for at least three speed lines with a minimum of three points per line, including the surge point. These tests are very expensive ($100,000 plus), and, in many cases, the unit cannot be run at full-load conditions or with the actual gas. These performance test results will, therefore, have to be corrected to represent actual conditions. Field tests are more appropriate, but if a field test is scheduled, careful planning should go into it at the designing and planning stages. The codes also deal with the guarantee, warranty, and vendors’ data required. These codes are well written and are based on experience.

| 106 |

| Compressor Definition and Standards | DRIVES GAS TURBINES The API Standard 616, Gas Turbines for the Petroleum, Chemical, and Gas Industry Services is intended to cover the minimum specifications necessary to maintain a high degree of reliability in an open-cycle gas turbine used in refinery service for mechanical drive, generator drive, or hot-gas generation. The standard also covers the necessary auxiliary requirements directly or indirectly by referring to other listed standards. The standard defines terms used in the industry and describes the basic design of the unit. It deals with the casing, rotors and shafts, wheels and blades, combustors, seals, bearings, critical speeds, pipe connections and auxiliary piping, mounting plates, weather-proofing, and acoustical treatment. The specifications call for a two-bearing construction preferably. Twobearing construction is desirable in single-shaft units, because a threebearing configuration can cause considerable trouble, especially when the center bearing in the hot zone develops alignment problems. The preferable casing is a horizontally split unit with easy visual access to the compressor and turbine, permitting field balancing planes without removal of the major casing components. The stationary blades should be easily removable without removing the rotor. A requirement of the standards is that the fundamental natural frequency of the blade should be at least two times the maximum continuous speed and at least 10% away from the passing frequencies of any stationary parts. Experience has shown that the natural frequency should be at least four times the maximum continuous speed. Care should be exercised on units where there is a great change in the number of blades between stages. A controversial requirement of the specifications is that rotating blades or labyrinths for shrouded rotating blades be designed for slight rubbing. A slight rubbing of the labyrinths is usually acceptable, but excessive rubbing can lead to major problems. New gas turbines use “squealer blades.” Some manufacturers suggest using ceramic tips, but whatever is done, great care should be exercised or blade failure and housing damage may occur. Labyrinth seals should be used at all external points, and sealing pressures should be kept close to atmospheric. The bearings can be either

| 107 |

| Centrifugal Compressors | rolling element bearings usually used in aero-derivative gas turbines or hydrodynamic bearings used in the heavier frame type gas turbines. In the area of hydrodynamics bearings, tilting pad bearings are recommended, since they are less susceptible to oil whirl and can better handle misalignment problems. Critical speeds of a turbine operating below its first critical should be at least 20% above the operating speed range. The term commonly used for units operating below their first critical speed is that the unit has a “stiff shaft,” while units operating above their first critical speed are said to have a “flexible shaft.” There are many exciting frequencies that need to be considered in a turbine. Some of the sources that provide excitation in a turbine system are: 1. 2.

3. 4. 5. 6. 7.

Rotor Unbalance Whirling mechanisms such as: a) Oil Whirl b) Coulomb Whirl c) Aerodynamic Cross Coupling Whirl d) Hydrodynamic Whirl e) Hysteretic Whirl Blade and Vane Passing Frequencies Gear Mesh frequencies Misalignment Flow separation in boundary layer exciting blades Ball/race frequencies in antifriction bearings usually used in Aeroderivative gas turbines

Torsional criticals should be at least 10% away from the first or second harmonics of the rotating frequency. Torsional excitations can be excited by some of the following: •

Start up conditions such as speed detents

•

Gear problems such as unbalance and pitch line runout

•

Fuel pulsation especially in low NOx Combustors

The maximum unbalance is not to exceed 2.0 mils on rotors with speeds below 4,000 rpm, 1.5 mils for speeds between 4,000-8,000 rpm, 1.0 mil for speeds between 8,000-12,000 rpm, and 0.5 mils for speeds above 12,000 rpm. These requirements are to be met in any plane and also

| 108 |

| Compressor Definition and Standards | include shaft runout. The following relationship is specified by the API standard: Lv =

12000 N

(3-12)

where : Lv = Vibration Limit mils (thousandth of an inch), or mm (mils x 25.4) N = Operating speed (rpm) The maximum unbalance per plane (journal) shall be given by the following relationships: U max = 4W / N

(3-13)

where : Umax = Residual Unbalance ounce –inches (gram-millimeters) W= Journal static Weight Lbs (Kg) A computation of the force on the bearings should be calculated to determine whether or not the maximum unbalance is an excessive force. The concept of an amplification factor (AF) is introduced in the new API 616 standard. Amplification factor is defined as the ratio of the critical speed to the speed change at the root mean square of the critical amplitudes as seen in the compressor API code. AF =

N c1 (N 2 − N1 )

(3-14)

Balancing requirement in the specifications require that the rotor with blades assembled must be dynamically balanced without the coupling but with the half key, if any, in place. The specifications do not discuss whether this balancing is to be done at high-speeds or low-speeds. The balancing conducted in most shops is at low-speed. A high-speed balancing should be used on problem shafts and any units that operate above the second critical speed. Field balancing requirements should be specified. The lubrication system for the turbine is described. This system closely follows the outline in API Standard 614, which is discussed in detail in Chapter 11. Separate lubrication systems for various sections of

| 109 |

| Centrifugal Compressors | the turbine and driven equipment may be supplied. Many vendors and some manufacturers provide two separate lubrication systems: one for hot bearings and another for cool bearings. These and other lubrication systems should be detailed in the specifications. The inlet and exhaust systems in gas turbines are described. The inlet and exhaust systems consist of an inlet filter, silencers, ducting, and expansion joints. The design of these systems can be critical to the overall design of a gas turbine. Proper filtration is a must, otherwise problems of blade contamination and erosion ensue. The standards are minimal for specifications, calling for a coarse metal screen to prevent debris from entering, a rain or snow shield for protection from the elements, and a differential pressure alarm. Most manufacturers are now suggesting socalled high-efficiency filters that have two stages of filtration, an inertia stage to remove particles above five microns followed by one or more filter screens, self cleaning filters, pad type pre-filters, or a combination of them, to remove particles below five microns. Manufacturers provide differential pressure alarms, but the trend among users has been to ignore them. It is suggested to pay more attention to differential pressure than in the past to assure high-efficiency operation. Silencers are also minimally specified. Work in this area has progressed dramatically in the past few years with the NASA quiet engine program. There are some good silencers now available on the market, and inlets can be acoustically treated. Starting equipment will vary, depending on the location of the unit. Starting drives include electric motors, steam turbines, diesel engines, expansion turbines, and hydraulic motors. The sizing of a starting unit will depend on whether the unit is a single-shaft turbine or a multipleshaft turbine with a free-power turbine. The vendor is required to produce speed-torque curves of the turbine and driven equipment with the starting unit torque superimposed. In a free-power turbine design the starting unit has to overcome only the torque to start the gas generator system. In a single-shaft turbine the starting unit has to overcome the total torque. Turning gears are recommended in the specifications, especially on large units, to avoid shaft bowing. They should always be turned on after the unit has been “brought down” and should be kept operational until the rotor is cooled. The gears should meet API Standard 613. Gear units should be doublehelical gears provided with thrust bearings. Load gears should be provided with a shaft extension to permit torsional vibration measurements. On high-speed gears, proper use of the lubricant as a coolant should be

| 110 |

| Compressor Definition and Standards | provided. Spraying oil as a coolant on the teeth and face of the units is recommended to prevent distortion. Couplings should be designed to take the necessary casing and shaft expansion. Expansion is one reason for the wide acceptance of the dry flexible coupling. A flexible diaphragm coupling is more forgiving in angular alignment; however, a gear-type coupling is better for axial movement. Access for hot alignment checks must be provided. The couplings should be dynamically balanced independently of the rotor system. Controls, instrumentation, and electrical systems in a gas turbine are defined. The outline in the standard is the minimum a user needs for safe operation of a unit. More details of the instrumentation and controls are given in other chapters in this book. The starting system can be manual, semiautomatic, or automatic, but in all cases should provide controlled acceleration to minimum governor speed and then, although not called for in the standards, to full-speed. Units, which do not have controlled acceleration to full-speed, had burned out first- and second-stage nozzles when combustion occurred in those areas instead of in the combustor. Purging the system of the fuel after a failed start is mandatory, even in the manual operation mode. Sufficient time for the purging of the system should be provided so that the volume of the entire exhaust system has been displaced at least five times. Alarms should be provided on a gas turbine. The standards call for alarms to be provided to indicate malfunction of oil and fuel pressure, high exhaust temperature, high differential pressure across the air filter, excessive vibration levels, low oil reservoir levels, high differential pressure across oil filters, and high oil drain temperatures from the gearings. Shutdown occurs with low oil pressure, high exhaust temperature, and combustor flameout. It is recommended that shutdown also occur with high thrust bearing temperatures and with a temperature differential in the exhaust temperature. Vibration detectors suggested in the standards are non-contacting probes. Presently, most manufacturers provide velocity transducers mounted on the casing, but these are inadequate. A combination of non-contacting probes and accelerometers are needed to ensure the smooth operation and diagnostic capabilities of the unit. Fuel systems can cause many problems, and fuel nozzles are especially susceptible to trouble. A gaseous fuel system consists of fuel filters, regulators, and gauges. Fuel is injected at a pressure of about 60 psi (4 Bar) above the compressor discharge pressure for which a gas compression system is needed. Knockout drums or centrifuges are

| 111 |

| Centrifugal Compressors | recommended, and should be implemented to ensure no liquid carry-overs in the gaseous system. Liquid fuels require atomization and treatment to inhibit sodium and vanadium content. Liquid fuels can drastically reduce the life of a unit if not properly treated. Recommended materials are outlined in the standards. Some of the recommendations in the standard are carbon steel for base plates, heattreated forged steel for compressor wheels, heat-treated forged alloy steel for turbine wheels, and forged steel for couplings. The growth of materials technology has been so rapid, especially in the area of high temperature materials, that the standard does not deal with it. Details of some of the materials technology of the high temperature alloys and single crystal blades are dealt with in another chapter. The turbine undergoes three basic tests: hydrostatic, mechanical, and performance. Hydrostatic tests are to be conducted on pressurecontaining parts with water at least one-and-a-half times the maximum operating pressure. The mechanical run tests are to be conducted for at least a period of four hours at maximum continuous speed. This test is usually done at no-load conditions. It checks out the bearing performance and vibration levels as well as overall mechanical operability. It is suggested that the user have a representative at this test to tape record as much of the data as possible. The data are helpful in further evaluation of the unit or can be used as base-line data. Performance tests should be conducted at maximum power with normal fuel composition. The tests should be conducted in accordance with ASME PTC-22. Table 3-4 indicates the main points an engineer must consider in evaluating different gas turbine units. This table will enable the engineer to make a proper evaluation of each critical point and ensure that he is purchasing units of high reliability and efficiency. STEAM TURBINES The API 612 covers the minimum requirements for special-purpose steam turbines. These requirements include basic design, materials, related lubrication systems, controls, auxiliary equipment, and accessories. Special-purpose turbines are horizontal turbines used to drive equipment that is relatively large in size and is in critical service. This category of steam turbines is not limited by a temperature, pressure, or speed. The standards are very much like the gas turbines in the rotor dynamics of the standards.

| 112 |

| Compressor Definition and Standards | • • • • • • • • • • • • • • • • • • • • • • • • • • • • • •

Type of Turbine ° Aeroderivative ° Frame Type Type of Fuel Type of Compressor Number of stages and pressure ratio Types of blades, blade attachment, and wheel attachment Number of bearings Type of bearings Type of thrust bearings Critical speed Torsional criticals Campbell diagrams Balance planes Balance pistons Type of combustor Wet and dry combustors Types of Fuel nozzles Transition pieces Type of turbine Power Transmission curvic coupling Number of stages Free-power turbine Turbine inlet temperature Type of fuels Fuel additives Types of couplings Alignment data Exhaust diffuser Performance map of turbine and compressor Gearing Drawings

Accessories • • • • •

Lubrication Systems Intercoolers Inlet filtration system Control system Protection system

TABLE 3-4: Points to consider in a gas turbine.

The turbines have axial split casings, which would allow the removal of the rotor and other wear parts to be removed without removing the casing from the foundations. Axially split casings may be split radially between high and low pressure portions. The casings shall be designed to have sufficient strength and rigidity to limit any changes in shaft alignment at the coupling flanges

| 113 |

| Centrifugal Compressors | Lubrication systems shall follow the API 614 standards. Care must be taken to not allow steam to contaminate the oil. The important points to consider in evaluating a steam turbine are listed in Table 3-5. This table will enable the engineer to make a proper

•

•

• • • • • • • •

•

•

• • • • • • •

•

Number of Sections ° HP Section ° IP Section ° LP Section Steam Characteristics (per casing) ° Mass Flow ° Temperature ° Pressure Type of seals (inner seal) and oil seals Type of bearings (radial) Bearing stiffness coefficients Types of thrust bearings (Tapered land, non-equalizing tilting pad, and Kingsbury) Thrust float Temperature for journal and thrust bearings (operating temperature) Critical speed diagram (Speed vs. bearing stiffness curve) Types of Rotors ° Shrouded or unshrouded Blades ° Blading Type (Impulse or Reaction) ° Attachment of blades to hub and shroud Attachment of Rotor to shaft ° Shrink fit ° Key fit ° Curvic Couplings Campbell diagrams of Rotors and Nozzles. ° Number of blades (impellers) ° Number of blades (diffuser) ° Number of blades (guide vanes) Balance planes (location ° How it is balanced (detail) Weight of rotor (assembled) Type of casing ° Split casing Data on torsional vibration (Bending criticals) Alignment data Type of coupling between Units Performance curves (separate casings) ° Temperature ° Pressure ° Efficiency Condenser ° Back Pressure ° Pressure drop ° Effectiveness

TABLE 3-5: Points to consider in a steam turbine.

| 114 |

| Compressor Definition and Standards | evaluation of each critical point and ensure that he is purchasing units of high reliability and efficiency. GEARS This standard API Standard 613 covers special purpose gears. They are defined as gears that have either or both actual pinion speeds of more than 2,900 rpm and pitchline velocities of more than 5,000 ft/min (27 meters/sec). The standard applies to helical gears employed in speedreducer or speed-increaser units. The scope and terms used are well defined and includes a listing of standards and codes for reference. The purchaser is required to make decisions regarding gear-rated horsepower and rated input and output speeds. This standard includes basic design information and is related to AGMA Standard 421. Specifications for cooling water systems are given as well as information about shaft assembly designation and shaft rotation. Gear-rated power is the maximum power capability of the driver. Normally, the horsepower rating for gear units between a driver and a driven unit would be 110% of the maximum power required by the driven unit or 110% of the maximum power of the driver, whichever is greater. The tooth pitting index or K factor is defined as follows:

K=

Wt (R + 1) x Fxd R

(3-15)

where: Wt = transmitted tangential load, in pounds at the operating pitch diameter Wt =

12600 x Gear rated horsepower Pinion rpm x d

(3-16)

F = net face width, inches d = pinion pitch diameter, inches R = ratio (number teeth in gear divided by number teeth in pinion) The allowable K factor is given by

| 115 |

| Centrifugal Compressors | Allowable K = Material index number/Service Factor

(3-17)

Service factors and material index number tabulation are provided for various typical applications, allowing the determination of the K factor. Gear tooth size and geometry are selected so that bending stresses do not exceed certain limits. The bending stress number is given by St = bending Stress Number =

Wt xPnd 1.8 cos ϕ x(SF) x F J

(3-18)

where: Wt = as defined in Equation (3-11) Pnd = normal diametral pitch F = net face width, inches ϕ = helix angle J = geometry factor (from AGMA 226) SF = service factor Design parameters are provided for casings, joint supports, and bolting methods. Some service and size criteria are included. Critical speeds correspond to the natural frequencies of the gears and the rotor bearings support system. Knowing the natural frequency of the system and the forcing function helps to make a determination of the critical speed. Typical forcing functions are caused by rotor unbalance, oil filters, misalignment, and a synchronous whirl. Gear elements must be multiplane and dynamically balanced. Where keys are used in couplings, half keys must be in place. The maximum allowable unbalanced force at maximum continuous speed should not exceed 10% of static weight load on the journal. The maximum allowable residual unbalance in the plane of each journal is calculated using the following relationship: F = mrϖ2

(3-19)

Since the force must not exceed 10% of the static journal load, mr =

0.1W ϖ2

Taking the correction constants, the equation can be written

| 116 |

(3-20)

| Compressor Definition and Standards | Max. unbalanced force = 56,347 x Journal static weightload Max. unbalanced force = (rpm)2 Max. unbalanced force =

(3-21)

The double amplitude of unfiltered vibration in any plane measured on the shaft adjacent to each radial bearing is not to exceed 2.0 mils (0.05mm) or the value given by Amplitude =

12000 rpm

(3-22)

where rpm is the maximum continuous speed. It is more meaningful for gears to be instrumented using accelerometers. Design specifications for bearings, seals, and lubrication are also given. Accessories such as couplings, coupling guards, mounting plates, piping, instrumentation, and controls are described. Inspection and testing procedures are detailed. After notifying the vendor, the purchaser is allowed to inspect the equipment during manufacture. All welds in rotating parts must receive 100% inspection. To conduct a mechanical run test, the unit must be operated at maximum continuous speed until bearing and lube oil temperatures have stabilized. Then the speed is increased to 110% of maximum continuous speed and run for four hours. L U B R I C AT I O N S Y S T E M S This API Standard 614 standard covers the minimum requirements for lubrication systems, oil shaft sealing systems, and related control systems for special purpose applications. The terms are fully defined, references are well documented, and basic design is described. Lubrication systems should be designed to meet continuously all conditions for a nonstop operation of three years. Typical lubricants should be hydrocarbon oils with approximate viscosities of 150 SUS at 100°F (37.8ºC). Oil reservoirs should be sealed to prevent the entrance of dirt and water and sloped at the bottom to facilitate drainage. The reservoir working capacity should be sufficient for at least a five-minute flow. The oil system should include a main oil pump and a standby oil pump. Each pump must have its own driver sized according to API Standard 610. Pump capacities should be based on the systems’ maximum usage plus a minimum of 15%. For seal oil systems, the pump capacity should be maximum capacity plus 20% or 10 gpm, whichever is greater.

| 117 |

| Centrifugal Compressors | The standby oil pump should have an automatic startup control to maintain safe operation if the main pump fails. Twin oil coolers should be provided, and each should be sized to accommodate the total cooling load. Full-flow twin oil filters should be furnished downstream of the coolers. Filtration should be 10 microns nominal. The pressure drop for clean filters should not exceed 5 psi at 100oF operating temperature during normal flow. Overhead tanks, purifiers, and degassing drums are covered. All pipe welding is to be done according to Section IX of the ASME code, and all piping must be seamless carbon steel, minimum schedule 80 for sizes 11⁄2 inch (38.1 mm) and smaller, and a minimum of schedule 40 for pipe sizes 2 inches (50.8 mm) or greater. The lubrication control system should enable orderly startup, stable operation, warning of abnormal conditions, and shutdown of main equipment in the event of impending damage. A list of required alarm and shutdown devices is provided. The purchaser has the right to inspect the work and testing of subcomponents if he informs the vendor in advance. Each cooler, filter, accumulator, and other pressure vessels should be hydrostatically tested at 11⁄2 times design pressure. Cooling water jackets and other water-handling components should be tested at 11⁄2 times design pressure. The test pressure should not be less than 115 psig (7.9 Bar). Tests should be maintained for durations of at least 30 minutes. Operational tests should:

| 118 |

•

Detect and correct all leaks.

•

Determine relief pressures and check for proper operation of each relief valve.

•

Accomplish a filter cooler changeover without causing startup of the standby pump.

•

Demonstrate the control valves have suitable capacity, response, and stability.

•

Demonstrate the oil pressure control valve can control oil pressure.

| Compressor Definition and Standards | V I B R AT I O N M E A S U R E M E N T S The API Standard 670 covers the minimum requirements for noncontacting vibration in an axial-position monitoring system. The accuracy for the vibration channels should meet a linearity of ±5% of 200 millivolts per mil sensitivity over a minimum operating range of 80 mils. For the axial position, the channel linearity must be ±5% of 200 millivolts per mil sensitivity and a ±1.0 mil of a straight line over a minimum operating range of 80 mils. Temperature should not affect the linearity of the system by more than 5% over a temperature range of -30°F to 350°F (-34.4°C to 176.7°C) for the probe and extension cable. The oscillator demodulator is a signal conditioning device powered by -24 volts of direct current. It sends a radio frequency signal to the probe and demodulates the probe output. It should maintain linearity over the temperature range of -30°F to 150°F(-34.4°C to 65.6°C). The monitors and power supply should maintain their linearity over a temperature range of -20°F to 150°F(-28.9°C to 65.6°C). The probes, cables, oscillator demodulators, and power supplies installed on a single train should be physically and electrically interchangeable. The non-contacting vibration and axial position monitoring system, consists of probe, cables, connectors, oscillator demodulator, power supply, and monitors. The probe tip diameters should be .190-.195 inches (4.8-4.95mm) with body diameters of 1 ⁄4 (6.35mm) - 28 UNF – 2A threaded, or .3–.312 inches (7.62–7.92mm) with a body diameter of 3⁄8 (9.52mm) -24 UNF - 24A threaded. The probe length is about one inch long. Tests conducted on various manufacturers’ probes indicate that the .3-312-inch (7.62-7.92mm) probe has a better linearity in most cases. The integral probe cables have a cover of tetra-flouroethylene, a flexible stainless steel armoring which extends to within four inches of the connector. The overall physical length should be approximately 36 inches (914.4mm) measured from probe tip to the end of the connector. The electrical length of the probe and integral cable should be six feet. The extension cables should be coaxial with electrical and physical lengths of 108 inches (2743.2mm). The oscillator demodulator will operate with a standard supply voltage of -24 volts dc and will be calibrated for a standard electrical length of 15 feet (5 meters). This length corresponds to the probe integral cable and extension. Monitors should operate from a power supply of 117 volts ±5% with the linearity requirements specified. False shutdown from power interruption will be prevented regardless of mode or duration. Power supply failure should actuate an alarm.

| 119 |

| Centrifugal Compressors | The radial transducers should be placed within three inches of the bearing, and there should be two radial transducers at each bearing. Care should be taken not to place the probe at the nodal points. The two probes should be mounted 90° apart (±5°) at a 45° (±5°) angle from each side of the vertical center. Viewed from the drive end of the machine train, the x probe will be on the right side of the vertical, and the y probe will be on the left side of the vertical. The axial transducers should have one probe sensing the shaft itself within 12 inches (305mm) of the active surface of the thrust collar with the other probe sensing the machined surface of the thrust collar. The probes should be mounted facing in opposite directions. Temperature probes embedded in the bearings are often more useful in preventing thrustbearing failures than the proximity probe. This is because of expansion of the shaft casing and the probability that the probe is located far from the thrust collar. When designing a system for thrust bearing protection, it is necessary to monitor small changes in rotor axial movement equal to oil film thickness. Probe system accuracy and probe mounting must be carefully analyzed to minimize temperature drift. Drift from temperature changes can be unacceptably high. A functional alternative to the use of proximity probes for bearing protection is bearing temperature, bearing temperature rise (bearing temperature minus bearing oil temperature), and rate of change in bearing temperature. A matrix combining these functions can produce a positive indication of bearing distress. A phase angle transducer should also be supplied with each train. This transducer should record one event per revolution. Where intervening gearboxes are used, a mark and phase angle transducer should be provided for each different rotational speed. S P E C I F I C AT I O N S The previous API standards are guidelines to information regarding machine train applications. The more pertinent the information obtained during the evaluation of the proposal, the better the selection for the problem. These standards should be followed as much as possible as they are written with years of experience behind them.

| 120 |

| Compressor Definition and Standards | REFERENCES P E R F O R M A N C E S P E C I F I C AT I O N S ASME, Performance Test Code on Test Uncertainty: Instruments and Apparatus PTC 19.1, 1988 ASME, Performance Test Code on Gas Turbines, ASME PTC 22 1997, American Society of Mechanical Engineers 1997 ASME, Performance Test Code on Steam Turbines, ASME PTC 6 1996 ASME, Performance Test Code on Atmospheric Water Cooling Equipment PTC 23, 1997 ASME GAS TURBINE FUELS B 133.7M Published: 1985 (Reaffirmed year: 1992) ISO, Natural Gas—Calculation of Calorific Value, Density and Relative Density International Organization for Standardization ISO 6976-1983(E) Table of Physical Constants of Paraffin Hydrocarbons and other components of Natural Gas—Gas Producers Association Standard 2145-94 M E C H A N I C A L S P E C I F I C AT I O N S ASME BASIC GAS TURBINES B 133.2 Published: 1977 (Reaffirmed year: 1997) ASME GAS TURBINE CONTROL AND PROTECTION SYSTEMS B133.4 Published: 1978 (Reaffirmed year: 1997) ASME GAS TURBINE INSTALLATION SOUND EMISSIONS B133.8 Published: 1977 (Reaffirmed: 1989) ASME MEASUREMENT OF EXHAUST EMISSIONS FROM STATIONARY GAS TURBINE ENGINES B133.9 Published: 1994 ASME PROCUREMENT STANDARD FOR GAS TURBINE ELECTRICAL EQUIPMENT B133.5 Published: 1978 (Reaffirmed year: 1997) ASME PROCUREMENT STANDARD FOR GAS TURBINE AUXILIARY EQUIPMENT B133.3 Published: 1981 (Reaffirmed year: 1994) API Std 611, General Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services , 4th Edition, June 1997 API Std 613, Special Purpose Gear Units for Petroleum, Chemical and Gas Industry Services, 4th Edition, June 1995

| 121 |

| Centrifugal Compressors | API Std 614, Lubrication, Shaft-Sealing, and Control-Oil Systems and Auxiliaries for Petroleum, Chemical and Gas Industry Services, 4th Edition, April 1999 API Std 616, Gas Turbines for the Petroleum, Chemical and Gas Industry Services, 4th Edition, August 1998 API Std 617, Centrifugal Compressors for Petroleum, Chemical and Gas Industry Services, 6th Edition, February 1995 ANSI/API Std 670, Vibration, Axial-Position, and Bearing-Temperature Monitoring Systems, 3rd Edition, November 1993 API Std 671, Special Purpose Couplings for Petroleum Chemical and Gas Industry Services, 3rd. Edition, October 1998 API Std 672, Packaged, Integrally Geared Centrifugal Air Compressors for Petroleum, Chemical, and Gas Industry Services, 3rd. Edition, September 1996 API Std 677, General-Purpose Gear Units for Petroleum, Chemical and Gas Industry Services, 2nd Edition, July 1997, Reaffirmed March 2000 ISO 10436: 1993 Petroleum and Natural Gas Industries—General Purpose Steam Turbine for Refinery Service, 1st. Edition.

| 122 |

4

Design Characteristics

Centrifugal compressors are used extensively because of their smooth operation, large tolerance of process fluctuations, and their higher reliability compared to other types of compressors. Centrifugal compressors range in size from pressure ratios of 1:3 per stage to as high as 12:1 on experimental models. The selection of this type of compressor is prevalent in the petrochemical industry. The proper selection of a compressor is a complex and important decision. The successful operation of many plants depends on smooth and efficient compressor operations. To ensure the best selection and proper maintenance of a centrifugal compressor, the engineer must have knowledge of many engineering disciplines. Centrifugal compressors in general are used for higher-pressure ratios and lower-flow rates compared to lower-stage pressure ratios and higherflow rates in axial compressors. Figure 4-1 is a map for centrifugal compressors that shows the effect of specific speed (Ns) and specific diameter (Ds) on their efficiency. The most efficient region for centrifugal compressor operation is in a specific speed range between 60 1,500. Specific speeds of more than 3,000 usually require an axial-flow-type compressor. In a centrifugal compressor the angular momentum of the gas flowing through the impeller is increased partly because the impeller’s outlet diameter is significantly greater than its inlet diameter. The major difference between axial and centrifugal compressors is the variance in the diameters of the inlet and the outlet. The flow leaving the centrifugal compressor is usually perpendicular to the axis of rotation.

| 123 |

| Centrifugal Compressors |

Figure 4-1: Centrifugal compressor map (Balje, O.E., “A Study of Reynolds Number in Effects in Turbomachinery” Journal of Engineering for Power, ASME Trans., Vol. 86, Series A, p.227).

CENTRIFUGAL COMPRESSOR COMPONENTS The terminology used to define the components of a centrifugal compressor is shown in Figure 4-2. A centrifugal compressor is composed of inlet guide vanes, an inducer, an impeller, a diffuser, and a scroll for a single stage compressor or a transition piece for multiple stage centrifugal compressors as shown in Figure 4-3. The inlet guide vanes (IGVs) are used in only a high-pressure ratio transonic compressor. Vaned Diffusers, as shown in Figure 4-4, are used in compressors for high efficiency. They are preceded by a vaneless diffuser so as to ensure that the velocity reaching the vaned diffuser is subsonic. Vaned diffusers reduce the operating margin of a centrifugal compressor. Centrifugal compressor impellers are either open-faced, unshrouded, or closed-faced as seen in Figure 4-5 and Figure 4-6 respectively. The fluid comes into the compressor through an intake duct and is given prewhirl by the IGVs. It then flows into an inducer without any incidence angle, and its flow direction is changed from axial to radial. The fluid is given energy at this stage by the rotor as it goes through the

| 124 |

| Design Characteristics |

Figure 4-2: Single stage centrifugal compressor.

Figure 4-3: Multistage centrifugal compressor.

| 125 |

| Centrifugal Compressors |

Figure 4-4: Flow in a vaned diffuser.

Figure 4-5: Open-faced impeller (courtesy MAN Turbomaschinen AG-GHH BORSIG).

| 126 |

| Design Characteristics |

Figure 4-6: Closed-faced impellers (courtesy MAN Turbomaschinen AG Schweiz).

inducer

Figure 4-7: Pressure and velocity through a centrifugal compressor.

| 127 |

| Centrifugal Compressors | impeller while compressing. It is then discharged into a diffuser, where the\kinetic energy is converted into static pressure. The flow enters the scroll from which the compressor discharge is taken. Figure 4-7 shows the variations in pressure and velocity through a compressor. In a typical centrifugal compressor, the fluid is forced through the impeller by rapidly rotating impeller blades. The velocity of the fluid is converted to pressure, partially in the impeller and partially in the stationary diffusers. Most of the velocity leaving the impeller is converted into pressure energy in the diffuser, as shown in Figure 4-7. It is normal practice to design the compressor so that half the pressure rise takes place in the impeller and the other half in the diffuser. The diffuser consists essentially of vanes, which are tangential to the impeller. These vane passages diverge to convert the velocity head into pressure energy. The inner edge of the vanes is in line with the direction of the resultant airflow from the impeller as shown in Figure 4-4. IMPELLER An impeller in a centrifugal compressor imparts energy to a fluid. The impeller consists of two basic components: (1) an inducer like an axialflow rotor, and (2) the radial blades where energy is imparted by centrifugal force. Flow enters the impeller in the axial direction and leaves in the radial direction. The velocity variations from hub to shroud resulting from these changes in flow directions complicate the design procedure for centrifugal compressors. C. H. Wu has presented the threedimensional theory in an impeller, but it is difficult to solve for the flow in an impeller using the previous theory without certain simplified conditions. Others have dealt with it as a quasi-three-dimensional solution. It is composed of two solutions, one in the meridional surface (hub-toshroud), and the other in the stream surface of revolution (blade-toblade). These surfaces are illustrated in Figure 4-8. By the application of the previous method using a numerical solution to the complex flow equations, it is possible to achieve impeller efficiencies of more than 90%. The actual flow phenomenon in an impeller is more complicated than the one calculated. One example of this complicated flow is shown in Figure 4-9. The stream lines observed in Figure 4-9 do not cross but are actually in different planes observed near the shroud. Figure 4-10 shows the flow in the meridional plane with separation regions at the inducer section and at the exit.

| 128 |

| Design Characteristics |

Figure 4-8: The two major planes in a centrifugal impeller.

Figure 4-9: Flow map in a blade-to-blade plane (Boyce, M.P., “A Practical ThreeDimensional Flow Visualization Approach to the Complex Flow Characteristics in a Centrifugal Impeller” ASME paper No 66-GT-83).

| 129 |

| Centrifugal Compressors |

streamlines Figure 4-10: Flow in the meridional plane, also known as the hub-to-shroud plane (Boyce, M.P., “A Practical ThreeDimensional Flow Visualization Approach to the Complex Flow Characteristics in a Centrifugal Impeller,” ASME paper No 66-GT-83).

Figure 4-11: Velocity distribution in the meridional plane of a centrifugal impeller (Senoo, Y. and Nakaske, Y., “An analysis of flow through a mixed flow impeller,” ASME Paper No 71-GT-2)

| 130 |

| Design Characteristics | Experimental studies of the flow within impeller passages have shown that the distribution of velocities on the blade surfaces is different from the distributions predicted theoretically. It is likely that the discrepancies between theoretical and experimental results are due to secondary flows from pressure losses and boundary-layer separation in the blade passages. High-performance impellers should be designed, when possible, with the aid of theoretical methods for determining the velocity distributions on the blade surfaces. Examples of the theoretical velocity distributions in the impeller blades of a centrifugal compressor are shown in Figure 4-11 and Figure 4-12. The blades should be designed to eliminate large decelerations or accelerations of flow in the impeller that lead to high losses and separation of the flow. Potential flow solutions predict the flow in regions well away from the blades where boundary-layer effects are negligible. In a centrifugal impeller, the viscous shearing forces create a boundary layer with reduced kinetic energy. If the kinetic energy is reduced below a certain limit, the flow in this layer becomes stagnant, then it reverses.

Figure 4-12: Velocity distribution in the blade-to-blade plane of a centrifugal impeller (Senoo, Y. and Nakaske, Y., “An Analysis of Flow Through a Mixed Flow Impeller,” ASME Paper No 71-GT-2).

| 131 |

| Centrifugal Compressors | I N L E T G U I D E VA N E S There are two kinds of inlet systems. A single-entry and a doubleentry, as shown in Figure 4-13. A double-entry system halves the inlet flow so that a smaller inducer-tip diameter can be used, reducing the inducertip Mach number; however, the design is difficult to integrate into many configurations. Figure 1-25 shows such a compressor.

Figure 4-13: Types of entry—inducer systems.

The inlet guide vanes give circumferential velocity to the fluid at the inducer inlet and, in many cases, are also used for control as seen in Figure 4-14. This function is called prewhirl. Figure 4-15 shows inducer inlet velocity diagrams with and without IGVs. IGVs are installed directly in front of the inducer or, where an axial entry is not possible, located radially in an intake duct. A positive vane angle produces prewhirl in the direction of the impeller rotation, and a negative vane angle produces prewhirl in the opposite direction. The disadvantage of positive prewhirl is that a positive inlet whirl velocity reduces the energy transfer, since Vθ1 is positive according to the Euler Equation defined by: H = (U 1Vθ 1 − U 2Vθ 2 )

| 132 |

(4-1)

| Design Characteristics |

Figure 4-14: Adjustable inlet guide vane (courtesy Atlas Coppco Comptec, Inc.).

For non-prewhirl (without IGVs axial entry), Vθ1 is equal to zero. Then the Euler work is H = - U2 Vθ2. =

(4-2)

With positive prewhirl, the first term of the Euler Equation remains H = (U 1Vθ 1 − U 2Vθ 2 )

Therefore, Euler work is reduced by the use of positive prewhirl. On the other hand, negative prewhirl increases the energy transfer by the amount U1 Vθ1. This results in a larger pressure head being produced in the case of the negative prewhirl for the same impeller diameter and speed.

| 133 |

| Centrifugal Compressors |

Figure 4-15: Inlet velocity diagrams.

The positive pre-whirl decreases the relative Mach number at the inducer inlet. However, negative pre-whirl increases it. A relative Mach number is defined by

M rel =

W1 a1

where: Mrel = relative Mach number W1 = relative velocity at an inducer inlet a1 = sonic velocity at inducer inlet conditions

| 134 |

(4-3)

| Design Characteristics | The purpose of installing the IGVs is to decrease the relative Mach number at the inducer-tip (impeller eye) inlet because the highest relative velocity at the inducer inlet is at the tip section. When the relative velocity is close to or greater than the sonic velocity, a shock wave takes place in the inducer section. A shock wave produces shock loss and chokes the inducer. Figure 4-16 shows the effect of inlet prewhirl on compressor efficiency.

Figure 4-16: Estimated effect of inlet prewhirl (Rodgers, C., and Shapiro, L., “Design Considerations for High Pressure Ratio Centrifugal Compressors,” ASME Paper No 73-GT-31.

There are three kinds of prewhirl: 1

Free-vortex prewhirl. This type is represented by r1Vθ1 = constant with respect to the inducer inlet radius. This prewhirl distribution is shown in Figure 4-17. Vθ1 is at a minimum at the inducer inlet shroud radius. Therefore, it is not effective in decreasing the relative Mach number in this manner.

| 135 |

| Centrifugal Compressors | 2

Forced-vortex prewhirl. This type is shown as Vθ1/r1 = constant. This prewhirl distribution is also shown in Figure 4-17. Vθ1 is at a maximum at the inducer inlet shroud radius, contributing to a decrease in the inlet relative Mach number.

3

Control-vortex prewhirl. This type is represented by Vθ1 = Ar1 + B/r1, where A and B are constants. This equation shows the first type with A = 0, B ≠ 0 and the second type with B = 0, A ≠ 0.

Figure 4-17: Prewhirl distribution patterns.

Figure 4-18: Euler work distribution at an impeller exit, showing the effects of pre-rotation.

| 136 |

| Design Characteristics | The prewhirl distribution should be made not only from the relative Mach number at the inducer inlet shroud radius but also from Euler work distribution at the impeller exit. Uniform impeller exit flow conditions, considering the impeller losses, are important factors in obtaining good compressor performance. Euler work distributions at an impeller exit, due to pre-rotation of the flow at the inlet, with respect to the impeller width, are shown in Figure 4-18. INDUCER The function of an inducer is to increase the fluid’s angular momentum without increasing its radius of rotation. In an inducer section, the blades bend toward the direction of rotation as shown in Figure 4-19. There are three forms of inducer camber lines in the axial direction. They are circular arc, parabolic arc, and elliptical arc. Circular arc camber lines are used in compressors with low-pressure ratios, while the elliptical arc produces good performance at highpressure ratios where the flow has transonic mach numbers.

Figure 4-19: Inducer in a centrifugal compressor.

| 137 |

| Centrifugal Compressors |

Splitter Blades

Figure 4-20: Open-faced impeller with splitter blades (courtesy Man Turbomaschinen AG Schweiz).

Because of choking conditions in the inducer, many compressors incorporate a splitter-blade design, as seen in Figure 4-20. The flow pattern in such an inducer section is shown in Figure 4-21. This flow pattern indicates a separation on the suction side of the splitter blade. Other designs include tandem inducers. In tandem inducers, the inducer section is slightly rotated as shown in Figure 4-21. This modification gives additional kinetic energy to the boundary, which is otherwise likely to separate. CENTRIFUGAL SECTION OF IMPELLER The flow in this section of the impeller enters from the inducer section and leaves the impeller in the radial direction. There are three impeller vane types at the exit, as shown in Figure 4-22. These are defined according to the exit blade angles. Impellers with exit blade angle β2 = 90° are radial vanes. Impellers with β2 < 90° are backward-curved or

| 138 |

| Design Characteristics |

Figure 4-21: Splitter blade and tandem inducer (Boyce, M.P. and Nishida, A., “Investigation of Flow in Centrifugal Impeller with Tandem Inducer,” JSME Paper, Tokyo Japan, May 1977).

Figure 4-22: Velocity triangles for various types of impeller blading.

| 139 |

| Centrifugal Compressors | backward-swept vanes, and for β2 > 90o, the vanes are forward-curved or forward-swept. They have different characteristics of theoretical headflow relationship to each other, as shown in Figure 4-23. Although in Figure 4-23 the forward-curved head is the largest, in actual practice the head characteristics of all the impellers are similar to the backwardcurved impeller. Table 4-1 shows the advantages and disadvantages of various impellers. Types of Impellers Radial Vanes

Backward Curved Vanes

Forward Curved Vanes

Advantages Reasonable compromise between low energy transfer and high absolute outlet velocity. Low outlet kinetic energy = low diffuser inlet mach number Wide surge margin High energy transfer

Disadvantages

Low energy transfer Complex bending stress Hard manufacturing High outlet kinetic energy = high diffuser inlet mach number Complex bending stress Hard manufacturing

Table 4-1: Advantages and disadvantages of various types of blade types.

Figure 4-23: Head flow characteristics for various outlet blade angles.

| 140 |

| Design Characteristics | The Euler Equation, assuming simple one-dimensional flow theory, is the theoretical amount of work imparted to each pound of fluid as it passes through the impeller, and it is given by H = (U 1Vθ 1 − U 2Vθ 2 )

(4-4)

where: H U2 U1 Vθ2 Vθ1

= work per lb of fluid = Impeller peripheral velocity = inducer velocity at the mean radial station = absolute tangential fluid velocity at impeller exit = absolute tangential air velocity at inducer inlet

For the axial inlet, Vθ1 = 0 Then H = - 1/gc (U2 Vθ2).

(4-5)

Supposing constant rotational speeds, no slip, and an axial inlet, the velocity triangles are as shown in Figure 4-24. For the radial vane, the absolute tangential fluid velocity at the impeller exit is constant—even if the flow rate is increased or decreased.

Figure 4-24: Velocity triangles at the exit of the impeller.

| 141 |

| Centrifugal Compressors | Therefore, H ≈ U 2Vθ''2 ≈ U 2Vθ 2 ≈ U 2Vθ′2

(4-6)

For backward-curved vanes, the absolute tangential fluid velocity at the impeller exit increases with the reduction of flow rates and decreases with the increase in flow rate as shown in the following equation: H ≈ −U 2Vθ''2 > −U 2Vθ 2 < −U 2Vθ′2

(4-7)

For forward-curved vanes, the absolute tangential fluid velocity at the impeller exit decreases with the reduction of flow rates and increases with the decrease in flow rate as shown in the following equation: H ≈ −U 2Vθ''2 < −U 2Vθ 2 > −U 2Vθ′2

(4-8)

SLIP FACTOR The flow at the impeller exit is not completely guided by the blades, and, hence, the effective fluid outlet angle does not equal the blade outlet angle. To account for flow deviation (which is similar to the effect accounted for by the deviation angle in axial-flow machines), the slip factor is used:

µ=

Vθ 2 Vθ 2 ∞

(4-9)

where Vθ2 is the tangential component of the absolute exit velocity with a finite number of blades, and Vθ2∞ is the tangential component of the absolute exit velocity, if the impeller were to have an infinite number of blades (no slipping back of the relative velocity at outlet). With radial blades at the exit,

µ=

| 142 |

Vθ 2 U2

(4-10)

| Design Characteristics | Flow in a rotating impeller channel (blade passage) will be a vector sum of flow with the impeller stationary and the flow due to rotation of the impeller, as seen in Figure 4-25. In a stationary impeller, the flow is expected to follow the blade shape and exit tangentially to it. A high adverse pressure gradient along the blade passage and subsequent flow separation are not considered to be general possibilities. Inertia and centrifugal forces cause the fluid elements to move closer to and along the leading surface of the blade toward the exit. Once out of the blade passage, where there is no positive impelling action present, these fluid elements slow down.

Figure 4-25: Forces and flow characteristics in a centrifugal compressor.

CAUSES OF SLIP IN AN IMPELLER The definite cause of the slip phenomenon that occurs within an impeller is not known. However, some general reasons can be used to explain why the flow is changed.

Coriolis Circulation Because of the pressure gradient between the walls of the two adjacent blades, the Coriolis forces, the centrifugal forces, and the fluid follow the Helmholtz vorticity law. The combined gradient that results causes a fluid movement from one wall to the other and vice versa. This

| 143 |

| Centrifugal Compressors | movement sets up circulation within the passage as seen in Figure 4-26. Because of this circulation, a velocity gradient results at the impeller exit with a net change in the exit angle.

Figure 4-26: Coriolis circulation.

Boundary-layer development The boundary layer that develops within an impeller passage causes the flowing fluid to experience a smaller exit area as shown in Figure 4-27. This smaller exit is due to small flow (if any) within the boundary layer. For the fluid to exit this smaller area, its velocity must increase. This increase gives a higher relative exit velocity. Since the meridional velocity remains constant, the increase in relative velocity must be accompanied with a decrease in absolute velocity. Although it is not a new approach, boundary-layer control is being used more than ever before. It has been used with success on airfoil designs when it has delayed separation, thus giving a larger usable angle of attack. Control of the flow over an airfoil has been accomplished in two ways: by using slots through the airfoil and by injecting a stream of fastmoving air.

Figure 4-27: Boundary layer development

| 144 |

| Design Characteristics | Separation regions are also encountered in the centrifugal impeller, as shown previously. Applying the same concept (separation causes a loss in efficiency and power) reduces and delays their formation. Diverting the slow-moving fluid away lets the separation regions be occupied by a faster stream of fluid, which reduces boundary-layer buildup and thus decreases separation.

Figure 4-28: Laminar flow control in a centrifugal compressor (Boyce ASME)

| 145 |

| Centrifugal Compressors | To control the boundary layer in the centrifugal impeller, slots in the impeller blading at the point of separation are used. To realize the full capacity of this system, these slots should be directional and converging in a cross-sectional area from the pressure to the suction sides as seen in Figure 4-28. The fluid diverted by these slots increases in velocity and attaches itself to the suction sides of the blades. This results in moving the separation region closer to the tip of the impeller, thus reducing slip and losses encountered by the formation of large boundary-layer regions. The slots must be located at the point of flow separation from the blades. Experimental results indicate improvement in the pressure ratio, efficiency, and surge characteristics of the impeller as seen in Figure 4-28.

Leakage Fluid flow from one side of a blade to the other side is referred to as leakage. Leakage reduces the energy transfer from impeller to fluid and decreases the exit velocity angle.

Number of Vanes The greater the number of vanes, the lower the vane loading, and the closer the fluid follows the vanes. With higher vane loadings, the flow tends to group up on the pressure surfaces and introduces a velocity gradient at the exit.

Vane Thickness Because of manufacturing problems and physical necessity, impeller vanes are thick. When fluid exits the impeller, the vanes no longer contain the flow, and the velocity is immediately slowed. Because it is the meridional velocity that decreases, both the relative and absolute velocities decrease, changing the exit angle of the fluid. A backward-curved impeller blade combines all these effects. The exit velocity triangle for this impeller, with the different slip phenomenon changes, is shown in Figure 4-29. This triangle shows that actual operating conditions are far removed from the projected design condition. Several empirical equations have been derived for the slip factor as seen in Figure 4-30. These empirical equations are limited to the type of impeller and flow. Three of the more common slip factors are presented here.

| 146 |

| Design Characteristics |

Figure 4-29: Effect on exit velocity triangles by various parameters.

Figure 4-30: Slip flow as a function of the flow coefficient.

| 147 |

| Centrifugal Compressors | Stodola Slip Factor The second Helmholtz law states that the vorticity of a frictionless fluid does not change with time. Hence, if the flow at the inlet to an impeller is irrotational, the absolute flow must remain irrotational throughout the impeller. As the impeller has an angular velocity ω, the fluid must have an angular velocity – ω relative to the impeller. This fluid motion is called the relative eddy. If there were no flow through the impeller, the fluid in the impeller channels would rotate with an angular velocity equal and opposite to the impeller’s angular velocity. To approximate the flow, Stodola’s theory assumes that the slip is due to the relative eddy. The relative eddy is considered as a rotation of a cylinder of fluid at the end of the blade passage at an angular velocity of –ω about its own axis. The Stodola slip factor is given by  π  sin β 2 µ = 1 − 1 − Z  Vm 2 cot β 2  U2

    

(4-11)

where: β2 = the blade angle Z = the number of blades Vm2= the meriodional velocity U2 = blade tip speed. Calculations using this equation have been found to be lower than experimental values.

Stanitz Slip Factor Stanitz calculated blade-to-blade solutions for eight impellers and concluded that, for the range of conditions covered by the solutions, U is a function of the number of blades (Z) and the blade exit angle (β2) is approximately the same whether the flow is compressible or incompressible.

| 148 |

| Design Characteristics | (4-12)    0.63π  1 1 −  µ =1− Z  Wm 2 cot β 2    U2

Stanitz’s solutions were for p/4 < b2 < p/2. This equation compares well with experimental results for radial or near-radial blades.

Balje Slip Factor Tests were conducted on open-faced impellers with axial entry and radial blades at the exit. The following relationship was obtained: 1 6.2

µ= 1+

2

z(

D2 3 ) D1mean

(4-13)

COMPRESSOR IMPELLER DESIGN In this section, we will discuss the blade loading analysis programs, which are presently available. Over the last two decades, many papers have dealt with the problem of blade loading in radial flow compressors. The simple design philosophy for the flow in the rotor is to consider it to be inviscid and the process to be isentropic. The inclusion of viscous terms in the momentum equation for a rotating coordinate system leads to an equation, which does not lend itself easily to a solution. Thus, the method of superimposing a boundary layer on the solution of an inviscid flow calculation has been found to be very attractive. In the blade-to-blade plane, this can be accomplished by superimposing the slip flow on the flow in this plane. The importance of the impeller flow is (a) to determine the amount of work done on the fluid, and (b) to determine the distortion of the exit. The flow in the centrifugal impeller is complex, because of the fact that the flow passage is rotating, it has curvature both from the axial to the radial direction and in the angular direction. The blade loading for centrifugal compressors require analysis methods for flow that is mainly three-dimensional. Because of the

| 149 |

| Centrifugal Compressors | difficulty of obtaining a true three-dimensional solution, the traditional method is to superimpose several two-dimensional solutions. This is often called quasi-three-dimensional solution. Since there are several choices of two-dimensional surfaces, and even more ways of combining them, there are several approaches to obtaining a quasi-three-dimensional solution. In each plane a two dimensional solution is obtained, which is then combined analytically to obtain a three-dimensional flow solution. The method described here is a quasi-three-dimensional solution. It is composed of two planes, one in the meridional plane and the other in the blade-to-blade plane. A meridional plane (hub-to-shroud) solution is a flow solution on a stream surface between the blades. The shape of this surface, as shown in Figure 4-8, is often taken to be the same as the mean blade surface. The θ-plane surface of revolution is also called blade-toblade surface, as can be seen in Figure 4-8. There are several theoretical methods for finding the distribution of velocities on the blades of centrifugal impellers. Experimental studies of the flow within impeller passages have shown that the distributions of velocities on the blade surfaces are rather different from the distributions predicted theoretically. These discrepancies between the theoretical and experimental results are due to secondary flows from the pressure losses and boundary layer separation in the blade passages. Boyce and Senoo, two experimentalists, have attempted to visualize in a three-dimensional manner the flow in a centrifugal impeller. The results of these studies are shown in Figures 4-9, 4-10, 4-11, and 4-12. These results indicate that many of the theories forwarded do not accurately predict the flow in the blade-to-blade plane. The most basic and comprehensive work to date in the area of centrifugal impeller flow has been presented by Wu in his papers in the early 1950’s. He provides expressions for determining the complete flow picture. However, after studying his method in great detail, it was concluded that without knowledge of some of the flow characteristics to be encountered through experimental data, the equations could not be fully solved. Wu also suggested a finite-difference approach for the solution of the flow field on both surfaces, which would generate a set of non-linear equations, which could then be solved using either an iterative matrix technique or a relaxation procedure. A blade design method in which the extension from an infinite number of blades to a finite number of blades is accomplished by the use of a power series in the circumferential

| 150 |

| Design Characteristics | direction. The terms in the series are determined by a comparison of the equations for an infinite number of blades and a finite number of blades. An irregular finite difference net on the blade-to-blade surface, generates a set of non-linear equations, which can be solved using an iterative matrix approach. The importance of this irregular net is that the boundary grid points fall on the physical boundaries of the machine, even when the boundaries are curved, as they generally are in centrifugal impellers.

ONE DIMENSIONAL FLOW S T R E A M T U B E C O O R D I N AT E S The flow in a centrifugal impeller is described by the velocities as shown in Figure 4-31.

Figure 4-31: Flow characteristics in a centrifugal compressor.

The impeller is divided into two planes, the meridional plane and the blade-to-blade plane. In this section, we are studying first the flow in the meridional plane based on a one-dimensional flow. Figure 4-32 shows the

| 151 |

| Centrifugal Compressors | flow in the meridional plane describing the streamlines and potential flow lines. Note that the potential flow lines must be perpendicular to the streamlines. The flow in the meridional plane is divided into a number of stream tubes. This is done graphically by dividing the flow in streamlines and potential flow lines. Potential flow lines must be perpendicular to the stream flow lines.

Figure 4-32: Streamlines and potential flow lines describing the flow in the meridional plane. Note that the potential flow lines must be perpendicular to the streamlines.

The method for the hub-to-shroud design presented herein with the blade-to-blade design is a rapid method ideal for computer use. The idea here has been to improve on their solution while maintaining many of their excellent suggestions. The solution of the entire blade loading is obtained in three steps. First the passage is divided into a number of stream tubes, and their coordinates are obtained. The assumption here is that all the stream tubes carry the same mass flow rate, stream tubes never cross each other, and the flow in each stream tube remains the same from the entrance to the exit. The second step is the calculation of the velocity gradient from hub to shroud. The distance between stream tubes

| 152 |

| Design Characteristics | vary as the potential flow lines must always be perpendicular to the stream tubes. Further, the velocity gradient can also be obtained by use of the equations developed here. Once the velocity gradient is obtained, a check is made to satisfy the continuity equation requirements. This results in several iterations until the flow equations are satisfied. The third step is the blade-to-blade solution; this requires the use of equations calculating the velocity and pressure on the trailing and driving faces. The calculation of the stream tube width shown in Figure 4-32 is based on the assumption that the flow through each stream tube is equal, thus: n2 2π m = ρ W cos β r dθ dn n N ∫ 1 ∫0

(4-14)

which reduces to

m N ∆n = ρ W cos β ( 2πr − Zt )

(4-15)

where Zt is the blockage caused by the blades, and ∆n the perpendicular distance between two adjacent streamlines in the meridional plane. The coordinates are given by the following equations:

r j = r j −1 +

∆n Cosα j −1

(4-16)

∆n sin α

(4-17)

and

z j = z j −1 −

These equations have to be further modified at the inlet and outlet to fit the geometry. To design a centrifugal compressor impeller, the following conditions must be defined: •

compressor (impeller plus diffuser) total pressure ratio

•

inlet conditions (total temperature and total pressure)

•

adiabatic efficiency of the compressor

| 153 |

| Centrifugal Compressors | •

exit conditions (total temperature and total pressure)

•

properties of the gas being compressed (e.g., specific heats at constant pressure and volume and their variation with pressure and temperature)

•

compressors having high inlet pressures or temperatures should take into account the compressibility factor Z, (a function of working and critical temperatures and pressures of the gas) which modifies the equation of state as p = ZρRT

•

mass-flow rate through the compressor

The following are some of the assumptions that are commonly made in designing a centrifugal impeller:

| 154 |

•

The velocity of flow in the inlet duct leading to the eye of the impeller is taken to be along the axis of rotation of the impeller just before entering it. In other words, there is no prewhirl, (although under some circumstances, when, for a given rpm and flow rate, the relative velocity at the inlet tends to increase the Mach number into trans-sonic or super-sonic regions, it is desirable to introduce some prewhirl to limit W1). A relation between the prewhirl angle, blade angle, and relative Mach number at the inlet is shown in Figure 4-33.

•

The axial absolute velocity at the inlet is assumed constant from the hub to the eye of the impeller.

•

For an absolute Mach number between 0.30 and 0.35 at the inlet, keeping the relative Mach number subsonic to avoid shock and consequent losses, a relative Mach number of about 0.9 is assumed to start with at inlet. As the peripheral velocity is greatest at the eye of the impeller, this relative Mach number applies to the relative velocity at that point.

•

Blade angles at the inlet are usually between 50º-65° at the mean inlet diameter, increasing towards the eye, reducing closer to the hub, and causing a twist to the blade at the inlet. This variation about the mean inlet blade angle depends on the change in peripheral velocity.

| Design Characteristics | •

For axial entry and a given enthalpy rise through an impeller, higher pressure ratios demand higher peripheral tip speeds, developing higher stress levels at the tip of the impeller. In order to keep these stresses to a minimum for a given rpm, exit blade angles of 90º are adopted

•

Blade passages are assumed to flow full with no separation.

Figure 4-33: Optimum mach number and prewhirl angle.

Figure 4-34 is a sketch of the centrifugal compressor as it is to be designed. In the case shown, the blade angle at the inlet is axial, thus there is no pre-whirl because the inducer has splitter blades. This is due to the fact that, by use of splitter blades, the inlet area is increased and the relative Mach number at the inlet decreased. At the exit, the design has assumed radial blades. At the inlet we would like to maintain a relative mach number at the mean diameter not to exceed Mrel = 0.8

M r1 =

W1 a1

(4-18)

| 155 |

| Centrifugal Compressors | where W1 is the relative inlet velocity and the acoustic velocity a1 is given as:

a = γg c RT

(4-19)

To start with, Static Inlet Temperature (Ts1) is not known because the Absolute Inlet Velocity (V1) is not known. Therefore, as a first approximation let

a 01 = g c γRTT 1

(4-20)

where the Inlet Total Temperature (Tt1) is known. Hence, the inlet relative velocity,(W1) can be written as:

W1 = M r 1 x a1 = M r1 x g cγRTt1

Figure 4-34: Flow through a centrifugal impeller.

| 156 |

(4-21)

| Design Characteristics | and for a known blade angle (β1) V1 = W1 cos β1

(4-22)

using this value for the absolute Velocity (V1), we have 2

Ts1 = Tt1 −

V1 2 g c JC P

(4-23)

_______ Recalculating, a1 = √gcγRTs1 and so on, after two or three iterations, we arrive at V1 = W1 cos β1

(4-24)

U1 = W1 Sin β1

(4-25)

Density at inlet:

ρ1 =

Ps1 R Ts 1

(4-26)

Assuming that the process between the static and total conditions is adiabatic, then the density at the inlet can be re-written as follows: γ

 T  γ −1 Ps1  s1   To1  ρ1 = R Ts1

(4-27)

Required inlet flow-area can be computed as follows: From the cycle analysis, the mass flow rate is known and the continuity equation and can be written as: m = ρ1 A1Vm1

(4-28)

where Vm1 is the inlet meridional velocity, which at the inlet, if there is no pre-whirl, is equal to the absolute inlet Velocity (V1), thus the mass flow can be re-written as:

| 157 |

| Centrifugal Compressors | (4-29)

m = ρ 1 AinletV1

Thus the flow area at the inlet is: Ainlet = m ρ1V1

(4-30)

The minimum hub diameter would be designed from a strength point of view, i.e., the maximum shear stress developed should be less than the maximum allowable for that material. It is better to avoid key-fastenings in order to lessen balancing problems. Shrink-fitting or press-fitting the impeller onto the shaft, keeping in mind the operating temperatures, is a desirable option. P02 Knowing the Total Pressure Ratio -------as obtained from equation (2-51) P01 γ

P02 γ −1 =[ (U 1Vθ 1 − U 2Vθ 2 ) + 1] γ −1 P01 γRT01

From Euler’s turbine equation, the work input to the compressor which has an axial flow entry (Vθ1 = 0) and radial exit (Vθ2 = µU2)can be written as:

U2 =

U2 =

D2 =

γRT01  P02    µ (γ − 1)  P01 

γRT01  P02    µ (γ − 1)  P01  U2 πN

γ −1 γ

(4-31) γ −1 γ

(4-32)

(4-33)

The speed (N) is often dictated by the compressor drive. For example, in electric motor drives the speeds are usually limited to 3600 rpm and 1800 rpm for 60 Hz, and 3000 rpm and 1500 rpm for 50 Hz systems. Gas turbines and steam turbines usually have higher speeds, and, in many cases, the compressor are driven through a gear box.

| 158 |

| Design Characteristics | The Adiabatic Head as given by the Euler Turbine Equation (2-34) γ −1 γ −1       P02  γ   γRT01   P02  γ     − 1 − 1 = = c p (T02 − T01 ) = c p        P01    γ − 1   P01     

H ad

From Figure 4-1 for the computed Specific Speed: Ns =

N Q1 ( H ad ) 3 / 4

(4-34)

. m where Q1= the volume flow rate ft3/sec, is given by Q1 = -----ρ1 and N= rpm, Had =ft-lbf/lbm At the maximum efficiency, the Specific Diameter is read from Figure 4-1 1

Ds =

D2 H ad4 Q1

(4-35)

thus, the exit Diameter (D2)can be calculated:

D2 =

Ds Q1 1

H ad4

(4-36)

Slip is defined as the ratio of the impeller tip speed (U2) and the tangential component of the absolute velocity at the exit (Vθ2)

µ≡

Vθ 2 U2

(4-37)

Of the various expressions (empirical) for slip factors, the one suggested by Balje is most appropriate for Radial open-face impellers. From Figure 4-29, the value of slip factor based on flow coefficient and the Balje Equation is obtained. Balje’s Equation (4-13) relates the diameter ratios and the number of blades as shown below:

| 159 |

| Centrifugal Compressors | µ=

1+

1 6.2  D2   Z  D   1 mean 

=

Vθ 2 U2

2/ 3

by assuming the number of blades (Z) the optimum ratio of the D2/D1mean is obtained. The higher the number of blades, the more closely the flow follows the blade shape. At the same time the higher blade number causes more blade-blockage, a narrower passage, and more surface area. The higher flow velocities and surface area result in higher drag losses, and the higher velocities also might result in shock waves at the inlet to the diffuser. This value of D2 is counterchecked in a following step and changed if needed, which, in turn, would require Z to be re-estimated based on a new speed as seen in the design flow chart given in Figure 4-35.

Figure 4-35: Design flow chart for centrifugal compressor impeller and diffuser.

| 160 |

| Design Characteristics | N=

(4-38)

U1 π D1, mean

The value of D2 computed through equations 4-32 and 4-35 is then compared and can be iterated. The Area at the inlet can be defined as: π Ainlet = ------ D2eye - [hub blockage + blade blockage due to its finite 4 thickness + aerodynamic blockage due to boundary layer buildup on blade surfaces] Aerodynamic blockage at the inlet is assumed to be about 3-5% of the actual flow area available. Hence, it is by an iteration process that Deye can be calculated, neglecting the aerodynamic blockage, to start. For simplicity, it might be taken to be about 4-5% of the circular area corresponding to the eye-diameter. π Hub blockage = ------ Dhub2 (4-39) 4 Blade blockage = (blade height) (blade thickness) x (no. of blades)

=h xt x Z

Where blade height at the inlet =

(4-40)

 Deye − Dhub   =h   2  

Blade thickness (t) is determined from its fluid dynamic and mechanical strength viewpoints. At this stage, as the blade-shape and bladeloading are unknown, we assume some value which would later on be checked for strength.

Ainlet =

h=

(π xDmean x h − h x t x Z ) (1 − blockage)

Ainlet (1 − blockage) (π x Dmean − t x Z )

(4-41)

(4-42)

Dhub = Dmean - h

| 161 |

| Centrifugal Compressors | Deye = Dmean + h This now defines all the diameters at the inlet and exit. In the case of the radial exit impeller with axial entry, therefore, the following relationships:

Vm 2 = Vm1 = V1 Vθ 2 = µ U 2 = µ (πDN ) V2 = Vm 2 + Vθ22 W2 = Vm 2 + (U 2 − Vθ 2 ) 2

(4-43)

Air-angle at the exit (α2), as seen in Figure (4-31) is given by: W  ∝ 2 = tan −1  m 2   Vθ 2 

(4-44)

V22 Hence, Ts2 = T12 - --------------2gcJcp

Ps2 The density at the exit is ρ2 = -----------RTs2

i.e.,

T  Pt 2  s 2   Tt 2  ρ2 = RTs 2

γ γ −1

(4-45) The exit flow area required is then

A2 =

m ρ 2Wm 2

(4-46)

This area must equal the available flow area, i.e., A2 = πD2 b –{Blade Blockage + Aerodynamic Blockage} Blade blockage = (blade height x no. of blades x blade thickness) = b x z x t

| 162 |

| Design Characteristics |

A2 =

πD2 b − bxZxt (1 − blockage)

(4-47)

Thus, the exit blade height (b) is given by the following relationship

b=

A2 (1 − blockage) (πD2 − Zxt )

(4-48)

Aerodynamic blockage is assumed to be about 7–10% of the actual available flow area. Hence, it is by an iteration process that we can determine b from the above relation. The impeller axial distance as shown in Figure 4-33 is approximated by the following relationship D  H m ≈ (0.5 to 0.7 )  2   2 

(4-49)

Having computed Dhub, Deye, Dtip, and b and the impeller axial distance (Hm) as a first approximation, an iterative process of flowanalysis in the impeller comes up with the geometry of the blade in the meridional plane. To complete this design, the axial distance of the inducer and the blade angle distribution in the inducer will be assumed, and their value will be checked. The meridional shape now formed must be checked. This is done by calculating the area throughout the plane. The calculation of the area throughout the shape is done by graphically plotting the hub and shroud profiles. The meridional area is An = π xDn x h

(4-50)

The Area in the Inducer Section is An = π xDn x h x cos β

(4-51)

The Area curve in the impeller increases and then decreases after the inducer. These area changes should be smooth so that there is no sudden acceleration or deceleration in the impeller, as this could result in separation of the flow. Figure 4-36 shows the type of desired area changes

| 163 |

| Centrifugal Compressors | that are required in the impeller. Figure 4-37 shows the Velocity distribution that should be obtained in the impeller. The value of Ns calculated could be used to check the peak efficiency from a graph showing its variation with respect to Ns for radial flow compressors, as shown in Figure 2-8.

Adiabatic Efficiency Based on Loss Calculations The losses in the centrifugal compressor are based on the energy losses in the compressor. Figure 4-38 represents the energy balance. By calculating the losses, the compressor stage adiabatic efficiency (ηad) is computed and should be compared with what was originally selected from Figure 4-1. Continued iteration will allow these efficiencies to be the same. Chapter 6 details all the losses in the centrifugal impeller.

Figure 4-36: Meridional plane area distribution showing streamline and potential flow lines.

| 164 |

| Design Characteristics |

Figure 4-37: Velocities in a centrifugal impeller—inviscid flow.

Figure 4-38: Energy loss in an impeller.

| 165 |

| Centrifugal Compressors | THREE DIMENSIONAL BLADE LOADING ANALYSIS IN A CENTRIFUGAL COMPRESSOR Today the centrifugal impeller designs are based on computer fluid dynamics (CFD) technique. This requires development of a 3D computational mesh, as shown in Figure 4-39. Here we will discuss the equations that govern these techniques and the blade loading analysis programs, which are presently available. In this section, we will evaluate both the inviscid techniques and the viscous techniques for flow in the centrifugal impeller. The inclusion of viscous terms in the momentum equation for a rotating coordinate system leads to an equation, which does not lend itself easily to a solution. Thus, the method of superimposing a boundary layer on the solution of an inviscid flow calculation has been found to be very attractive.

Figure 4-39: Computational mesh for 3D viscous flow calculation impeller (courtesy Atlas Coppco Comptec, Inc.).

| 166 |

| Design Characteristics | The importance of the impeller flow is (a) to determine the amount of work done on the fluid, and (b) to determine the distortion of the flow at the exit. The flow in the centrifugal impeller is complex because of the fact that the flow passage is rotating, and it has curvature both from the axial to the radial direction and in the angular direction. The blade loading for centrifugal compressors requires analysis methods for flow that is mainly three-dimensional. Because of the difficulty of obtaining a true three-dimensional solution, the traditional method is to superimpose several two-dimensional solutions. This is often called quasithree-dimensional solution. Since there are several choices of two-dimensional surfaces, and even more ways of combining them, there are several approaches to obtaining a quasi-three-dimensional solution. In each plane, a two-dimensional solution is obtained, which is then combined analytically to obtain a three-dimensional flow solution. The method described here is a quasi-three-dimensional solution. It is composed of two planes, one in the meridional plane and the other in the blade–to–blade plane, similar to the planes described before. METHOD The method for the hub-shroud design presented herein with the blade-to-blade design is a rapid method ideal for computer use. The idea here has been to improve on their solution while maintaining many of their excellent suggestions. The solution of the entire blade loading is obtained in three steps. First, the passage is divided into a number of stream tubes, and their coordinates are obtained. The assumption here is that all the stream tubes carry the same mass flow rate. The second step is the calculation of the velocity gradient from hub to shroud. This is obtained by use of the equations developed here. Once the velocity gradient is obtained, a check is made to satisfy the continuity equation requirements. This results in several iterations until the flow equations are satisfied. The third step is the blade-to-blade solution; this requires the use of equations calculating the velocity and pressure on the trailing and driving faces.

| 167 |

| Centrifugal Compressors | VELOCITY GRADIENT IN THE MERIDIONAL PLANE The momentum equation in the rotational coordinate system: H

H H DW 1 + 2ω x W − ω 2 r = − ∇P Dt ρ

(4-52)

The above equation expanded in cylindrical coordinates can be written as follows:

W ∂Wr ∂Wr ∂Wr ∂Wr + Wr + θ + Wz − r ∂θ ∂z ∂t ∂r

W

θ2 1 − 2ωWθ − ω 2 r = − r ρ (4-53a)

1 ∂P ∂Wθ ∂Wθ W ∂Wθ ∂Wθ WW + Wr + θ + Wz + r θ + 2ωWr = − ∂t ∂r r ∂θ ∂z r or ∂θ

(4-53b)

∂Wz ∂Wz W ∂Wz ∂Wz 1 ∂P + Wr + θ + Wz =− ∂t ∂r r ∂θ θz ρ ∂z

(4-53c)

Let d ∂ ∂ W ∂ ∂ = + Wr + θ + Wz dt ∂z ∂r r ∂θ ∂z

Now, Equations (4-53-a, b, c) yield

dWr 1 1 ∂P − (Wθ + ωr ) 2 = − dt r ρ ∂r

(4-54a)

dWθ WW 1 ∂P + r θ + 2ωWr = − dt r ρr ∂θ

(4-54b)

dWz 1 ∂P =− dt ρ ∂z

(4-54c)

Let r and z be functions of θ. They can be represented in the form

| 168 |

| Design Characteristics | θ = θ (r,z) Assuming P* as the static pressure and denoting on the surface that extends from hub to shroud about midway between blades, we have P* = [r, θ (r,z), z]

(4-55)

Or P* = P (r,z) since θ on the surface is not an independent variable. The partial derivatives of the static pressure in the three dimensional field on the surface is given by ∂P * ∂P ∂P ∂θ = + ∂r ∂r ∂θ ∂r

(4-56a)

∂P * ∂P ∂P ∂θ = + ∂z ∂z ∂θ ∂z

(4-56b)

Substitution of equations (4-56a) and (4-56b) in equation (4-51) yields

dWr 1 1 ∂P * 1 ∂P ∂θ − (Wθ + 2ωr ) 2 = − + r dt r ρ ∂r ρr ∂θ ∂r

(4-57a)

dWθ WW 1 ∂P + r θ + 2ωWr = − dt r ρr ∂θ

(4-57b)

dWz 1 ∂P * 1 ∂P ∂θ =− + r dt ρ ∂z ρr ∂θ ∂z

(4-57c)

Substitution of equation (4-57b) in equation (4-57a) and (4-57c) and putting Vθ = Wθ + ωr yields ∂θ 1 d ( rVθ ) dWr Vθ 1 ∂P * − =− −r [ { }] ρ ∂r ∂r r dt r dt 2

(4-58a)

| 169 |

| Centrifugal Compressors | dWz ∂θ 1 d ( rVθ ) 1 ∂P * =− −r [ { }] ρ ∂z ∂z r dt dt

(4-58b)

The radial velocity component Wr and the axial velocity component Wz are related to meridional velocity Wm by the angle α. α is obtained from Fig. (4-31) -1 α = tan

dr

(4-59)

dz

and Wr = Wm sin α Wz = Wm cos α

(4-60a) (4-60b)

differentiation yields dWr dα dWm = Wm cos α + sin α dt dt dt dWz dα dWm = − Wm sin α + cos α dt dt dt

(4-61a)

(4-61b)

Let rc be the radius of curvature of meridional streamline and obtained from

1 / rc =

1 dα dα dα dm = / = dm dt dt Wm dt

(4-62)

Making use of equations (4-59), (4-60), and (4-62) and combining equation (4-58) and (4-61) yields Wm dWm V ∂θ 1 d ( rVθ ) 1 ∂P * −r cos α + sin α − θ = − [ { }] rc dt r dt ρ ∂r ∂r r 2

2

(4-63a) − Wm dWm ∂θ 1 d ( rVθ ) 1 ∂P * −r sin α + cos α = − [ { }] rc dt dt ρ ∂z ∂z r 2

(4-63b)

| 170 |

| Design Characteristics | Considering the directional derivative of P* in the direction of potential lines yields dP * ∂P * dr ∂P * dz = + dn ∂r dn ∂z dn

(4-64)

dz = cos (90 + α ) = − sin α dn

(4-65a)

dr = sin (90 + α ) = cos α dn

(4-65b)

But

Substituting equations (4-65a) and (4-65b) in equation (4-64) yields ∂P * ∂P * ∂P * = cos α − sin α ∂z ∂n ∂r

(4-66)

Multiply equation (4-63a) by cos α and (4-63b) by sin α to obtain Wm dWm V cos ∝ ∂∂ 1 d (rVθ ) 1 ∂P * =− cos 2 α + sin α cos α − θ cos α − r [ { } ] cos α ρ ∂r rc dt r ∂r r dt 2

2

(4-67a) Wm dWm 1 ∂P * sin 2 α + cos α sin α = − sin α ρ ∂z rc dt 2

−

−r

∂θ 1 d ( rVθ ) [ { } ] sin α ∂z r dt

(4-67b) Subtract equation (4-67b) by equation (4-67a) to obtain 2

2

Wm V 1 dP * B d ( rVθ ) − θ cos α = − + rc r ρ dn r dt

(4-68)

where B=r

∂θ ∂θ sin α − r cos α ∂z ∂r

(4-69)

| 171 |

| Centrifugal Compressors | Now P* can be replaced by P since there is no longer a need to differentiate between them. Therefore equation (4-68) yields. 2

2

Wm V B d ( rVθ ) 1 dP − θ cos α = − + rc r ρ dn r dt

(4-70)

or 2

2

W V B d ( rVθ ) 1 dP = − m + θ cos α + ρ dn rc r r dt

(4-71)

From the Figure (4-31) Wm = W cos β Wθ = W sin β

(4-72a) (4-72b)

Making use of equations (4-72a), (4-72b), and (4-69), Equation (4-71) can be written in the form: W 2 cos 2 β 1 dP (W sin β + ωr ) 2 =− + cos α rc r ρ dn + B W cos β (

dWθ W sin β sin α + + 2ω sin α dm r

(4-73)

Another relation between W and P can be obtained from the energy equation. The energy equation in the rotating coordinate system is

D' q D' W 2 (ωr ) 2 1 ∂' P = (h + − )− Dt Dt 2 2 ρ ∂t

(4-74)

For steady state along a streamline, it reduces to

h+

W2 ωr − ( )2 − q = 0 2 2

dh + d ( But

| 172 |

W2 ωr 2 ) − d( ) − dq = 0 2 2

(4-75)

(4-76)

| Design Characteristics | (4-77)

1 dP = dh − dq ρ Thus

1 dP = ω 2 rdr − WdW ρ

(4-78a)

P Since process in compressor is not isentropic, therefore ----= const ργ does not hold. Thus to take care of the non-isentropic process, polytropic exponent k is introduced

k=

ηγ 1 − γ (1 − η )

(4-78b)

Now integrating equation (4-78) along a streamline between the inlet and any station, the following relationship is obtained p k p 1 ω2 2 2 2 (r − ri ) ( − i ) = − (W 2 − Wi ) + k −1 ρ ρi 2 2

(4-79a)

noting that from the continuity equation (2-37)

dρ kP dP = dn ρ dn

(4-79b)

the equation (4-79) can be written as

1 / rc =

1 dα dα dα dm = / = dm dt dt Wm dt

(4-80)

Combining equations (4-80) and (4-73), the following is obtained W

dW B cos β sin α sin β cos 2 β sin 2 β cos α = W2 ( − − ) dn rc r r

− W ( 2 ω sin β cos α + B cos β (

dWθ + 2ω sin α ) dm

| 173 |

| Centrifugal Compressors | +(

d (Wi ) dr 1 dPi + Wi − riω 2 i ) dn dn ρ i dn

(4-81)

Assuming that between streamlines Pi, Wi on riω2 increase linearly, then one may assume that

1 dPi d 2 + (Wi ) − riω 2 cos α i ) = const. ρ dn dn

(

(4-82)

between streamline. Thus an equation of the form

1 dW 2 = W 2a −W b + c 2 dn

(4-83)

is obtained where a =(

B cos β sin β sin α cos2 β sin 2 β cos α − − ) rc r r

b = (2ω sin β cos α + B cos β ( dWθ + 2ω sin α )) dm 2

c= (

1 dPi d Wi + ( ) − riω 2 cos α i ) ρ i dn dn 2

The solution of equation (4-83) is

(

W j −1 − J Wj − J

) (

Wj − L W j −1 − L

L

) J = e a ∆n ( − J + L ) / J

(4-84)

where J=

L=

b−

b−

b 2 − 4ac 2a b 2 − 4ac 2a

Thus, knowing the characteristics of the previous streamline, the value of the adjacent streamline may be obtained.

| 174 |

| Design Characteristics | For the special case of the radial blades and a radial outlet, the value of the shape constant is b = 0. Thus the solution of equation (4-84) is reduced to the following: W j = { (W 2 j −1 + c / a ) e 2 a∆n −

c } a

1/ 2

(4-85)

The values of the relative velocity obtained in equations 4-84 and 4-85 are exact solutions of equation 4-83.

Derivation of Equations for Blade-to-Blade Solution Using the Energy equation in combination with the momentum equation in the blade-to-blade plane, relationships between the trailing and driving faces are obtained. The energy equation in the rotating coordinate system

D' q D' W2 (ωr ) 2 1 ∂' P = [h + − ]− Dt Dt 2 2 ρ ∂t

(4-86)

for steady state along a streamline reduces to

h+

W2 (ωr ) 2 − −q=0 2 2

(4-87)

differentiating dh + WdW − ω 2 rdr dq = 0

(4-88)

but

1 dP = dh − dq ρ WdW − ω 2 rdr +

1 dP = 0 ρ

(4-89)

Momentum equation in θ direction is

dWθ WW 1 ∂P + r θ + 2ωWr = dt r ρr ∂θ

(4-90)

| 175 |

| Centrifugal Compressors | From the solution in the meridional plane in the previous section, this reduces to

W cos β (

dWθ W 1 ∂P + sin β sin α + 2ω sin α ) = − dm r ρr ∂θ

(4-91a)

1 W dW sin β sinα + 2ω sinα ) rdθ = − dP + ρ r dm

(4-91b)

W cos β (

θ

Combining equations 4-91a and 4-91b W cos β (

dWθ W + sin β sin α + 2ω sin α ) rdθ = WdW − ω 2 rdr dm r

(4-92) Integrating and noting that dr remains constant

(WT − WD ) 2π − Ztθ dW W =( ) W cos β ( θ + sin β sin α + 2ω sin α ) 2 Z dm r 2

2

(4-93) Assuming (WD + WT ) =W 2

at constant radius. Equation (4-93) reduces to

(WT − WD ) = cos β (

2πr − Ztθ dWθ W )( + sin β sin α + 2ω sin α ) Z dm r

(4-94) The angle β at the exit must be the air angle and not the blade angle to take into account the effect of slip. The calculation of the pressure on the trailing and driving faces can now be computed easily. Compression in the rotor occurs at other than isentropic conditions, it occurs as some polytropic process. P / ρk = const η pγ k= 1 − γ (1 − η p )

| 176 |

| Design Characteristics | where ηp is the polytropic efficiency therefore (

Po T )=( o) Poi Toi

k k −1

k

 (T − T0i  k −1 Po = 1 + 0  Poi  T0i 

(4-95)

Euler’s Turbine equation 1 (To − Toi ) = (UVθ − U i Vθi ) cP k

Po  (UVθ − U iVθi  k −1 = 1 +  Poi  c p T0i  (U Vθ − U i Vθ i ) To  = 1 + Toi  (c P Toi )

(4-96)   

(4-97)

Assuming that the process from static to total conditions is isentropic, then T γ −1 V 2 = (1 − ) 2 ao 2 To

where a o = γRTo 2

γ

P  T  γ −1 =  Po T0  γ

γ

 T  γ −1  (UV − U iVθ i )  γ −1 P =   = 1 +  Poi T0  c pT0i  

(4-98)

This is an iterative solution to obtain the pressure and temperature in the blade-to-blade plane.

| 177 |

| Centrifugal Compressors | The viscous solution is much more complex than the inviscid solution. The results are more accurate. Figure 4-40 shows the velocity distribution in the centrifugal impeller using a viscous flow solution. Figures 4-41 and 4-42 compare the inviscid solution and the viscous solution for the velocity distribution in both the meridional and the blade–blade plane. The major difference occurs at the shroud where the impeller turns from an axial direction to the radial direction and the exit of the impeller. The viscous solution represents the build up of the boundary layer and the separation of the flow at the elbow and at the exit. In the Meridional plane, the velocity at the exit for the viscous solution indicates that hub velocity is higher than at the shroud. This agrees with experimental results but is opposite of that computed using an inviscid solution. In the blade-to-blade plane, there is much more of a change in the viscous and inviscid solutions. The suction side is affected more by the viscous solution than the pressure side.

Figure 4-40: Velocities in a centrifugal impeller—viscous flow.

| 178 |

| Design Characteristics |

Figure 4-41: Comparisons in velocities in the meridional plane in a centrifugal impeller using an inviscid flow solution as compared to a viscous flow solution.

Figure 4-42: Comparisons in velocities in the blade-to-blade plane in a centrifugal impeller using an inviscid flow solution as compared to a viscous flow solution.

| 179 |

| Centrifugal Compressors | REFERENCES CENTRIFUGAL COMPRESSORS Anderson, R.J., Ritter, W. K., and Dildine, D. M. “An Investigation of the Effect of Blade Curvature on Centrifugal Impeller Performance,” NACA TN-1313, 1947. Balje, O. E. “A Study of Reynolds Number Effects in Turbomachinery.” Journal of Engineering for Power, ASME Trans., Vol. 86, Series A, 1964, p. 227. Bammert, K. and Rautenberg, M, “On the Energy Transfer in Centrifugal Compressors.” ASME Paper No. 74-GT-121. Boyce, M. P. “Rerating of Centrifugal Compressors-Part II.” Diesel and Gas Turbine Worldwide. Jan.-Feb. 1989. 8-20. Boyce, M. P. “Rerating of Centrifugal Compressors-Part I.” Diesel and Gas Turbine Worldwide. Oct. 1988. 46-50. Boyce, M. P. “A Practical Three-Dimensional Flow Visualization Approach to the Complex Flow Characteristics in a Centrifugal Impeller.” ASME Paper No. 66-GT -83. Boyce, M. P. “Principles of Operation and Performance Estimation of Centrifugal Compressors.” Proceedings of the 22nd Turbomachinery Symposium. Dallas, TX. 14-16 Sept. 1993. 161-78. Boyce, M. P. and Nishida, A. “Investigation of Flow in Centrifugal Impeller with Tandem Inducer.” JSME Paper, Tokyo, Japan, May 1977. Boyce, M. P. “New Developments in Compressor Aerodynamics,” Proceedings of the 1st. Turbomachinery Symposium, Texas A&M, Oct. 1972. Boyce, M. P. and Bale, V. S. “A New Method for the Calculations of Blade Loadings in Radial-Flow Compressors,” ASME Paper No. 71-GT -60. Domercq, O. and Thomas, R. 1997. “Unsteady Flow Investigation in a Transonic Centrifugal Compressor Stage.” AIAA paper No. 97-2877. 1977. Klassen, H. A. “Effect of Inducer Inlet and Diffuser Throat Areas on Performance of a Low-Pressure Ratio Sweptback Centrifugal Compressor.” NASA TM X-3148, Lewis Research Center. Jan. 1975. Owczarek, J. A. Fundamentals of Gas Dynamics, International Textbook Company. Pennsylvania. 1968. pp. 165-197. Rodgers, C. “Effect of Blade Numbers on the Efficiency of a Centrifugal Impeller.” ASME 2000-GT-0455

| 180 |

| Design Characteristics | Rodgers, C. and Shapiro, L. “Design Considerations for High-PressureRatio Centrifugal Compressors.” ASME Paper No. 73-GT-31. Schlichting, H. Boundary Layer Theory, 4th edition, McGraw-Hill Book Co. 1962. pp. 547-550. Senoo, V. and Nakase, V. “An Analysis of Flow Through a Mixed Flow Impeller.” ASME Paper No. 71-GT-2. Shouman, A. R. and Anderson, J. R. “The Use of Compressor-lnlet Prewhirl for the Control of Small Gas Turbines.” Journal of Engineering for Power. Trans. ASME. Vol. 86. Series A. 1964. pp. 136-140. Wu, C. H. “A General Theory of Three-Dimensional Flow in Subsonic and Supersonic Turbomachines of Axial, Radial, and Mixed-Flow Type.” NACA TN-2604. 1952. Boyce, M. P. and Desai, A. R. “Clearance Loss in a Centrifugal Impeller.” Proc. of the 8th Intersociety Energy Conversion Engineering Conference. Aug. 1973. Paper No.7391 26. p. 638. Dallenback, F. “The Aerodynamic Design and Performance of Centrifugal and Mixed-Flow Compressors.” SAE International Congress. Jan. 1961. Boyce, M. P. “How to Achieve On-Line Availability of Centrifugal Compressors.” Chemical Weekly, June 1978. p. 115-127.

| 181 |

5

Diffuser Design Characteristics

Diffusers are designed to convert the high velocity exiting from the impeller to static pressure. The velocity in a diffuser is reduced, and the static pressure is increased while the total pressure and temperature remains constant. In a subsonic diffuser, the flow area increases. The diffuser throat is a region where the Mach number reaches sonic flow and controls the flow in the diffuser. Choke conditions are reached when the Mach number reaches 1.0 and thus no more flow can be sent through the diffuser. Diffusing passages have always played a vital role in obtaining good performance from turbomachines. Their role is to recover the maximum possible kinetic energy leaving the impeller with a minimum loss in total pressure. The velocity of the fluid is converted to pressure, partially in the impeller and partially in the stationary diffusers. Most of the velocity leaving the impeller is converted into pressure energy in the diffuser. It is normal practice to design the compressor so that half the pressure rise takes place in the impeller and the other half in the diffuser. The efficiency of centrifugal compressor components has been steadily improved by advancing their performance. However, significant further improvement in efficiency will be gained only by improving the pressure recovery characteristics of the diffusing elements of these machines, since these elements have the lowest efficiency. The performance characteristics of a diffuser are complicated functions of diffuser geometry, inlet flow conditions, and exit flow conditions. Figure 5-1 shows typical vaneless diffusers classified by their geometry. The selection of an optimum channel diffuser for a particular task is

| 183 |

| Centrifugal Compressors | difficult, since it must be chosen from an almost infinite number of crosssectional shapes and wall configurations. In radial and mixed-flow compressors, the requirement of high performance and compactness leads to the use of vaned diffusers, however, vaned diffusers reduce the operating margin of the compressor. Vaned diffusers are classified into four major classifications a) the most common is the airfoil style diffuser as seen in Figure 5-2, this type of diffuser relies on conventional standard cascade technology; here the swirl velocity component is reduced by the turning vanes; b) the vaned island or channel diffuser is shown in Figure 5-3, these types of diffuser control the increasing passage area by increasing the vane thickness with radius; at the exit of the vanes there is a substantial increase in area due to the large vane thickness; c) a low solidity vaned diffuser is a diffuser where the vanes are far apart to form any type of choking as there is no throat area; a schematic of such a diffuser is shown in Figure 5-4; d) the rib diffuser is a low solidity set of diffuser vanes attached to the shroud surface. The vanes extend only part of the way in the diffuser passage (between 25% to 60%), as shown in the schematic in Figure 5-5.

Figure 5-1: Vaneless diffuser shapes

| 184 |

| Diffuser Design Characteristics |

Figure 5-2: Adjustable air-foil style diffuser with variable guide vanes (courtesy Atlas Copco).

Figure 5-3: Island diffuser.

| 185 |

| Centrifugal Compressors |

Figure 5-4: Low solidity vaned diffuser.

Figure 5-5: Schematic of a rib diffuser.

In gas turbine applications where single-impeller very high-pressure ratios are achieved, the multistage vaned cascade diffuser is used. These cascade diffusers are multi-staged (Figure 5-6). Matching the flow between the impeller and the diffuser is complex because the flow path changes from a rotating system into a stationary one. This complex, unsteady flow is strongly affected by the jet-wake of the flow leaving the impeller, as seen in Figure 5-7. The three-dimensional boundary layers, the secondary flows in the vaneless region, and the flow separation at the blades also affect the overall flow in the diffuser.

| 186 |

| Diffuser Design Characteristics |

Figure 5-6: Cascade diffuser—showing three stages of diffusion (courtesy DresserRand Corporation).

| 187 |

| Centrifugal Compressors |

Figure 5-7: Flow wake leaving the impeller.

The flow in the diffuser is usually assumed to be of a steady nature to obtain the overall geometric configuration of the diffuser. In a channeltype diffuser, the viscous shearing forces create a boundary layer with reduced kinetic energy. If the kinetic energy is reduced below a certain limit, the flow in this layer becomes stagnant and then reverses. This flow reversal causes separation in a diffuser passage, which results in eddy losses, missing losses, and changed-flow angles. Separation should be avoided or delayed to improve compressor performance. The high-pressure-ratio centrifugal compressor has a narrow, yet stable, operating range. This operating range is due to the close proximity of the surge and choke flow limits. The word “surge” is widely used to express unstable operation of a compressor. Surge is the flow breakdown period during unstable operation. The unsteady flow phenomena during the onset of surge in a high-pressure-ratio centrifugal compressor causes

| 188 |

| Diffuser Design Characteristics | the mass flow throughout the compressor to oscillate during supposedly “stable” operations. The throat pressure in the diffuser increases during the precursor period up to collector pressure Pcol at the beginning of surge. All pressure traces (except plenum pressure) suddenly drop at the surge point. The sudden change of pressure can be explained by measuring the occurrence of backflow from the collector through the impeller during the period between the two sudden changes. R A D I A L VA N E L E S S D I F F U S E R In many high-performance centrifugal compressors, the diffuser system has a radial vaneless diffuser followed by a vaned diffuser. The main purpose of using the vaneless diffuser in this diffuser combination is to reduce the Mach number of fluid leaving the impeller to a value which can be efficiently accepted by the vaned diffuser. A vaneless diffuser can be used to reduce the velocity of supersonic flow to subsonic conditions without shock losses, provided the radial-velocity component is subsonic. The performance characteristics of a diffuser are complicated functions of diffuser geometry, inlet flow conditions, and exit flow conditions. The selection of an optimum channel diffuser for a particular task is difficult, since it must be chosen from an almost infinite number of crosssectional shapes and wall configurations. The flow in vaneless diffusers has been analyzed using one dimensional method. The Momentum and the Continuity Equation govern the flow in the radial vaneless diffuser. Introducing the skin friction factor, which is the ratio of the shear stress and the velocity head and, in this case, is given by: Cf =

τ 1 ρ V 2 g0

2

(5-1)

where τ is the wall shear stress. Modifying the momentum equation by substituting the skin friction, the following relationship is obtained for the meridional plane: 2

g 0 dP V V dV + C f 2 cos β = θ − V m m dr ρ dr b r

(5-2a)

| 189 |

| Centrifugal Compressors | the relationship in the tangential plane:

−C f

V2 dV V V sin β = Vm θ + θ m r b dr

(5-2b)

where β is the blade exit angle and b is the axial width of the diffuser. The effect of the blade exit angle is predominant when a very small friction coefficient was included in the calculation. Continuity ρrVm = const

(5-3)

The problems of wake mixing at the entrance to vaneless diffusers have been studied carefully showing that mixing takes place in a very small radial distance and, in most cases, the wakes are mixed at a radius 5% greater than the impeller radius. The instability in vaneless diffusers is caused by stall patches at high values of diffuser inlet angle, and this contributes to compressor surge. The limiting value of fluid inlet angle for instability considerations was found to be largely a function of diameter ratio of the vaneless diffuser. Diffuser width has a secondary effect on the stability. The parallel wall divergence should not exceed more than 81⁄2º otherwise separation of the flow will occur in the diffuser. The length of the channel beyond l/b > 4 does not improve the performance of the diffuser. What is gained by a further velocity recovery is lost in additional diffuser losses. The losses in the vaneless diffuser are functions of the skin friction losses. Similar to pipe flow, the losses are a function of both the divergence and the velocity of the flow stream. The resistance coefficient Cf is a function of a Reynolds Number, as shown in the Moody Diagram in Figure 5-8. The Reynolds number is given by the following relationship: Re =

VxD ν

D = hydraulic Diameter = 4 x Area/Wetted Perimeter

where v = kinematic viscosity

| 190 |

(5-4)

| Diffuser Design Characteristics |

Reynolds Number Re

Figure 5-8: Moody diagram showing the relationship of the frictional coefficient of diffusers as a function of the Reynolds Number (courtesy ASME).

VA N E D D I F F U S E R S It is generally acknowledged that a well-designed vaned diffuser arrangement inherently has a higher efficiency but has a smaller flow range of efficient operation than a vaneless diffuser; in addition, a vaned diffuser system occupies less space than an equivalent vaneless diffuser. Adding vanes in the vaneless diffuser causes significant change in the unsteady flow field inside the compressor. Figure 5-9 shows the flow regions in a vaned diffuser. These changes spread over the entire flow field and result in a different overall performance of the same compressor. In general, the degree of the unsteadiness in the compressor is reduced. The reduction of the unsteadiness naturally leads to less mixing loss, and turbulent energy dissipation in the system. This mechanism is one of the important points that leads to higher compressor efficiency. The peak efficiency of compressors with vaned diffusers is higher, to the extent of 2 points, than that of compressors with a plain volute and is better to the extent of 4 points, than that of units with vaneless diffuser and volute. Mechanical complications and cost are the main reasons why vaned diffusers are not used; reduced operating range is another important consideration. Another important disadvantage to the use of vaned diffusers in the petrochemical industry is the changes in molecular weight in the process over the life of the compressor, which often results in high incidence angle to the entry of the vaned diffuser.

| 191 |

| Centrifugal Compressors | The total inlet throat area of the diffuser is made equal to that of a volute casing for the same conditions. Although the restrictive action for through flow in the vaned diffuser is greater than that of the volute casing, the flow through the diffuser and in the collecting chamber is more orderly and the same capacities are realized in both designs for the same total throat area. Unsteady flow analysis indicates the flow fluctuation patterns are changed in addition to the mean flow field. Unlike the widely spreading flow fluctuation in the vaneless case, the flow in the vaned case shows a more controlled behavior. The temporal variation of the field quantities concentrates on the region near impeller exit and diffuser vane channels. The right design reduces the unsteadiness in these regions and improves the compressor characteristics of performance, vibration, and acoustics. The evolution of the stationary waves near diffuser vanes reveals that the matching the numbers of impeller blades and diffuser vanes impacts

Figure 5-9: Flow regions of the vaned diffuser.

| 192 |

| Diffuser Design Characteristics | the dynamic feature of the flow field. The conventional design practice is to avoid common divisors. Matching the flow between the impeller and the diffuser is complex because the flow path changes from a rotating system into a stationary one. The flow in the diffuser is usually assumed to be of a steady nature to obtain the overall geometric configuration of the diffuser. In a channel-type diffuser, the viscous shearing forces create a boundary layer with reduced kinetic energy. If the kinetic energy is reduced below a certain limit, the flow in this layer becomes stagnant and then reverses. This flow reversal causes separation in a diffuser passage, which results in eddy losses, missing losses, and changed-flow angles. Separation should be avoided or delayed to improve compressor performance. The inlet region of the vaned diffuser is very important, and care should be taken to ensure that velocity peaks are avoided on the blade surfaces. Many vaned diffusers have a slightly curved entrance region, to reduce velocity peaks, followed by straight blades. Experimental results have shown that diffuser vanes should be set at minus 4 degrees incidence in order to accommodate the variation in fluid angle along the leading edge of the diffuser vanes. Vaned diffusers for mixed flow compressors are designed using similar methods to those used in the design of impeller blades. The base circle D3 and the vane entrance angle for the vaned diffuser are established to ensure that the flow enters the vaned diffuser in a subsonic manner. Experiments have been conducted with notched leading edges for supersonic operation, as shown in Figure 5-10. This ensures that there

Figure 5-10: Schematic of a supersonic diffuser.

| 193 |

| Centrifugal Compressors |

Velocity Ratio (V3/Vu)

are no standing shock waves. Tests on such a vaned diffuser at an inlet Mach number of approximately 1.3 obtained efficiencies of 79 per cent efficiency. The diffuser width is made equal to or slightly greater than (1.05 to 1.10) the impeller width, depending on the number of vanes, vane angle, and the throat area (i.e., specific speed). The number of vanes should be a minimum necessary to realize the required throat area and maintain the shape of the channel between two vanes as discussed below. The outside diameter of the vaned part of the diffuser is not a criterion by itself, but depends on the number of vanes and channel proportions and the gap between the impeller and diffuser vanes. In a well-designed diffuser, the outside diameters vary from 1.35D2 to 1.6D2. For straight walled, two-dimensional diffusers, the ideal velocity ratio that can be attained without boundary layer separation depends on the geometry of the diffuser, as shown in Figure 5-11.

Figure 5-11: Separation limit for two-dimensional straight walled diffusers.

The optimum shape of the channel between two vanes of a diffuser is due to cascade testing and the following are some of the conditions which govern the design of the vaned diffuser: •

| 194 |

For a given area, the hydraulic radius of the channel should be a minimum. The best practical approach to this is a square section at or near the entrance to the diffuser. Round sections have

| Diffuser Design Characteristics | proven more efficient, but are difficult to adopt for practical use, except in special designs. •

The diffuser channel, confined between the two adjacent vanes, should be straight-walled conical.

•

The number of vanes should be a minimum required to form a good channel, the optimum length of the confined channel being fixed in a rather narrow limit. This is one of the major differences in design between airfoil cascade vaned diffusers and the low chord solidity (d = -----------------) diffusers. spacing

•

The angle of divergence of the diffuser channel should be equal to or smaller than those established for straight channels with a uniform velocity of approach. For a straight circular conical diffuser, an included angle of divergence of 81 ⁄2° was already quoted in connection with volute discharge nozzles. For a square section. The optimum divergence angle is about 6°. For a rectangular diffuser section, formed between two parallel walls, the divergence angle is about 11°. These values have been established experimentally and are not inconsistent if compared on the basis of the rate of area expansion.

•

The correct diffuser entrance angle is secondary to the optimum diffuser channel proportions. Although, as a starting point, this angle should be taken from the input discharge velocity triangle.

•

The diffuser depth (D4 - D3)/2 or the diffuser ratio D4 /D4 is not a controlling factor, however, tests have indicated that the confined diffuser passage is four times the opening between the two adjacent inlet vane tips, or l=4xa, as shown in Figure 5-12. Extending the length of the cascade diffuser channel beyond l/a > 4 does not improve the performance of the diffuser. What is gained by a further velocity recovery is lost in additional diffuser loss and eddy losses accompanying the joining of two streams from the two adjacent diffuser channels.

•

The preferred area expansion ratio is b/a = 1.6. This corresponds to about 81⁄2° of divergence for a diffuser with parallel sidewalls. The number of vanes and the exit diameter D4 are adjusted to

| 195 |

| Centrifugal Compressors | obtain the desired shape of the diffuser channel, however, very high solidity increases the losses in the vaned diffuser. Thus, there is a compromise between the high solidity which reduces the separation of flow but increases the drag losses.

Figure 5-12: Airfoil diffuser geometry.

Long and curved diffuser channels cannot perform conversion of velocity into pressure effectively as the flow becomes crowded toward the outer wall of the channel where local high velocities are restored. This has been the reason why the low solidity diffuser and the rib diffuser have become popular. Using flat, straight vanes for the cascade diffuser has its advantages. It is easier to manufacture and has lower stresses. Its disadvantages are flow separation and, in many cases, too many vanes, which results in high frictional losses. The number of vanes is fixed when the diffusion angle of the channel is selected. This is apparent when the vanes are radial, but the relationship remains when all blades are tilted through the same angle: z = 360 / δ

(5-5)

where z is the number of vanes and δ is the diffusion angle, as shown in Figure 5-13.

| 196 |

| Diffuser Design Characteristics |

Figure 5-13: Straight vaned flat plate diffuser showing the diffusion angle δ and the air angle entering the diffuser α .

| 197 |

| Centrifugal Compressors | Selection of the optimum number of diffuser vanes is difficult, but the number of diffuser vanes should be less than the number of impeller blades in order to avoid the possibility that large wakes from the impeller blades would sweep around the vaneless diffuser and block off individual vaned diffuser passages. In most cases this produces early stalling. The experimental evidence on this aspect tends to confirm this view although there are some anomalous results. The low solidity diffusers as seen in Figure 5-4 have a solidity less than one: σ =

chord < 1.0 spacing

(5-6)

These diffusers should have an incidence angle, which is determined by subtracting the vane setting angle from the diffuser flow angle calculated at the design flow rate. The incidence angle should be less than 4.0º. The vane configuration is a simple, flat plate with leading edge taper. Some of the advantages of the low-solidity vaned diffusers is that it causes the static pressure distribution at the impeller exit to be more uniform than in the vaneless diffuser case, this also reduces the strains within the impeller by as much as 60% and does not cause any unacceptable stresses to arise in the impeller. Thus, the LSDs reduce the load on an impeller due to static pressure non-uniformity caused by a discharge volute. CONTINUOUS DIFFUSER RETURN CHANNEL Diffuser return channels receive the flow from the diffuser at a reduced velocity, turn it 180° toward the inlet of the next impeller, as seen in Figure 5-14, and take out what is left of the tangential component of the velocity. If a tangential component is designed, it should be addressed separately. To perform this function efficiently, the flow must be gradually accelerated. To arrive at a proper curvature of the return channel vane, it is better to assume the development of a one-piece vane combining the diffuser and return channel vanes as seen in Figure 5-15. The diffuser vanes can be further modified by taking out some of the curvature of the channel and incorporating the favorable proportions discussed. The number of return channel vanes is established from the same considerations as that of the impeller and is lower than the latter by two to three vanes. Return channels are very important in the overall efficiency of the compressor. They can affect the overall efficiency by as much as 8 points.

| 198 |

| Diffuser Design Characteristics |

Figure 5-14: A five stage horizontally split centrifugal compressor, showing the return channels.

| 199 |

| Centrifugal Compressors |

Figure 5-15: Continuous diffuser return channel.

Utilization of the kinetic energy of the preceding stage in the following stage is the important guiding principle. The recovery of pressure should proceed no more than necessary to a slight acceleration in the U turn and next-stage impeller approach. A quite common fault of the return channels of the older design was that the fluid was “dumped” from the diffuser into the return channel without any attempt to guide the flow or control the velocities. There is an excessive hydraulic loss in the sharp 180° turn between the diffuser and return channel vanes. This loss is of the nature of eddy losses due to the abrupt change in velocity and direction of flow. Due to the short passage length, the friction loss is insignificant in this part of the flow passages. The straight flow length, before the U turn, is a major asset to ensure that the flow entering the U turn has an even flow distribution. Dividing the channel from stage to stage into a combination of two-vaned

| 200 |

| Diffuser Design Characteristics | parts (diffuser and return channel) and a vaneless U turn is done for manufacturing reasons. This could result in a change of 2 to 3 points in efficiency. Efficiencies of 80 percent are well established in the centrifugal compressor units for high-pressure service. SCROLL AND VOLUTES The purpose of the volute is to collect the fluid leaving the impeller or diffuser and deliver it to the compressor outlet pipe. Figure 5-16 shows a two-stage compressor where the last stage exits into a volute. The volute has an important effect on the overall efficiency of the compressor. Volute design embraces two schools of thought. First, the angular momentum of the flow in the volute is constant, neglecting any friction effects. The tangential velocity V5θ is the velocity at any radius in the volute. The following equation shows the relationship if the angular momentum is held constant: V5θ r = constant = K

(5-7)

Figure 5-16: Two stage closed wheel centrifugal compressor showing the return passage and the exit volute (courtesy MAN Turbomaschinen AG Schweiz).

| 201 |

| Centrifugal Compressors | Assuming no leakage past the tongue and a constant pressure around the impeller periphery, the relationship of flow at any section Qθ to the overall flow in the impeller Q is given by: Qθ =

θ Q 2π

(5-8)

Thus, the area distribution at any section θ can be given by the following relationship: Aθ = Qr x

θ L x 2π K

(5-9)

where: r = radius to the center of gravity L = volume width Second, design the volute by assuming that the pressure and velocity are independent of θ. The area distribution in the volute is given by: Aθ = K

Q θ x V5 θ 2 π

(5-10)

To define the volute section at a given θ, the shape and area of the section must be decided. Flow patterns in various types of volute are shown in Figure 5-17. The flow in the asymmetrical volute has a singlevortex instead of the double-vortex in the symmetrical volute. Where the impeller is discharging directly into the volute, it is better to have the volute width larger than the impeller width. This enlargement results in the flow from the impeller being bounded by the vortex generated from the gap between the impeller and the casing. At flows different from design conditions, there exists a circumferential pressure gradient at the impeller tip and in the volute at a given radius. At low flows, the pressure rises with the peripheral distance from the volute tongue. At high flows, the pressure falls with distance from the tongue. This condition results because, near the tongue, the flow is guided by the outer wall of the passage. The circumferential pressure gradients reduce efficiency away from the design point. Non-uniform pressure at the impeller discharge results in unsteady flows in the impeller passage, causing flow reversal and separation in the impeller.

| 202 |

| Diffuser Design Characteristics |

Figure 5-17: Flow patterns in a volute.

| 203 |

| Centrifugal Compressors | REFERENCES Boyce, M. P. and Bale, Y. S. “Diffusion Loss in a Mixed-Flow Compressor.” Intersociety Energy Conversion Engineering Conference. San Diego. Paper No. 729061. Sept. 1972. Dawes, W. 1995 “A Simulation of the Unsteady Interaction of a Centrifugal Impeller with its Vaned Diffuser: Flows Analysis,” ASME Journal of Turbomachinery. Vol. 117. pp. 213-222. Deniz, S., Greitzer, E. and Cumpsty, N. 1998. “Effects of Inlet Flow Field Conditions on the Performance of Centrifugal Compressor Diffusers Part 2: Straight-Channel Diffuser,” ASME paper, No. 98-GT-474. Filipenco, V., Deniz, S., Johnston, J., Greitzer, E., and Cumpsty, N. 1998 “Effects of Inlet Flow Field Conditions on the Performance of Centrifugal Compressor Diffusers Part 1: Discrete Passage Diffuser.” ASME paper, No. 98-GT-473. Johnston, R. and Dean, R. 1966. “Losses in Vaneless Diffusers of Centrifugal Compressors and Pumps.” ASME Journal of Basic Engineering. Vol. 88. pp. 49-60. Phillips, M. 1997. “Role of Flow Alignment and Inlet Blockage on Vaned Diffuser Performance.” Report No. 229. Gas Turbine Laboratory. Massachusetts Institute of Technology. Rodgers, C. “The Performance of Centrifugal Compressor Channel Diffusers.” ASME paper, No. 82-GT-10. Rodgers, C. “Influence of Impeller and Diffuser Characteristics and Matching on Radial Compressor Performance.” SAE Preprint 268B. Jan. 1961.

| 204 |

6

Off-Design Performance Characteristics

COMPRESSOR PERFORMANCE The performance of a centrifugal compressor is dependent on many factors, the most important being the composition and the physical and chemical characteristics of the gas it is compressing. Thus, the characteristics of the compressor must be corrected for the characteristics of the gas, to be evaluated at off-design considerations. P E R F O R M A N C E C A L C U L AT I O N S The performance calculations are based on the basic equations, as outlined in Chapter 2. To properly evaluate the trend of values, absolute values must be corrected to some conditions such as design conditions and, in the case of off-design operational conditions, some values must be transposed to design conditions. The following are some of the governing equations for correcting major performance parameters:

Corrected Compressor Mass Flow M act M corr =

P1 Pref

T1 Tref

1 MWact MW ref

(6-1)

| 205 |

| Centrifugal Compressors | where Mcorr and Mact is the corrected and actual mass flow, and Pref , Tref , MWref are the Pressure, Temperature, and Molecular Weight at reference conditions, such as design inlet conditions.

Corrected Compressor Volume Flow Qact Qcorr =

MWact MWref T1 T ref

(6-2)

where Qcorr and Qact is the compressor’s corrected and actual Volume flows.

Corrected (Aerodynamic) Speed MWact MWref

N act N corr =

T1 T ref

(6-3)

where Ncorr and Nact are the compressor’s corrected and actual speed

Corrected Power HPact HPcorr =

P1 Pref

MWact MWref T1 T

(6-4)

where HPcorr and HPact is the compressor’s corrected and actual required power

| 206 |

| Off-Design Performance Characteristics | Corrected Head (Adiabatic or Polytropic)

H corr = H act

MWact MW ref T1 T ref

(6-5)

where Hcorr and Hact is the compressor’s corrected and actual Head produced.

Corrected Pressure Pcorr =

Pact P1 Pref

(6-6)

where Pcorr and Pact are the compressor’s corrected and actual Total Pressure.

Corrected Temperature Tcorr =

Tact T1 Tref

MW MWref

(6-7)

where Tcorr and Tact are the compressor’s corrected and actual Total Temperature. The Compressor Performance map shown in Figure 6-1 is a typical performance map of a centrifugal compressor. The map shows a variation of the total pressure ratio or Head as a function of the corrected compressor mass flow at various aerodynamic speeds. The mass flow is corrected for pressure and temperature, and the mechanical speed is corrected for temperature. In this manner, the map can be used to check the performance of a compressor at variable inlet conditions for a given gas. These maps sometimes can be further modified to also account for changes in molecular weight.

| 207 |

| Centrifugal Compressors |

Figure 6-1: Compressor performance map.

CENTRIFUGAL COMPRESSOR LOSS Calculating the performance of a centrifugal compressor in both design and off-design conditions requires knowledge of various losses encountered in a centrifugal compressor. The losses in the centrifugal compressor are based on the energy losses in the compressor. Figure 6-2 represents the energy balance. By calculating the losses, the compressor stage adiabatic efficiency (ηad) is computed and should be compared with what was originally selected from Figure 4-1. The accurate calculation and proper evaluation of losses within a centrifugal compressor is as important as the calculation of the bladeloading parameters. If the proper parameters are not controlled, efficiency decreases. The evaluation of various losses is a combination of experimental results and theory. The losses are divided into two groups: (1) losses encountered in the rotor, and (2) losses encountered in the stator.

| 208 |

| Off-Design Performance Characteristics |

Figure 6-2: Energy loss in an impeller.

A loss is usually expressed as a loss of heat or enthalpy. A convenient way to express them is in a non-dimensional manner with reference to the exit blade speed. The theoretical total non–dimensional head available (qtot) is equal to the head available from the energy equation qth =

1 U2

2

[U 2Vθ 2 − U1Vθ1 ]

(6-8)

The total Head is the Impeller Head plus the Head which is lost externally to the impeller due to the air which by passes the impeller (∆qcl)and the disc friction loss (∆qdf) of the air between the stationary components and the impeller. q tot = q th + ∆ qdf + ∆ qcl

(6-9)

| 209 |

| Centrifugal Compressors | The adiabatic head that is actually available at the rotor discharge is equal to the theoretical head minus the heat from the shock in the rotor (∆qsh), the inducer loss (∆qin), the diffusion blading loss (∆qdbl), and the viscous losses encountered in the flow passage (∆qsf). q i a = qth − qin − ∆ q sh − ∆ qdbl − ∆ qsf

(6-10)

Therefore, the adiabatic efficiency in the impeller is ηimp =

qia qtot

(6-11)

The calculation of the overall stage efficiency must also include losses encountered in the diffuser. Thus, the overall actual adiabatic head attained will be the actual adiabatic head of the impeller minus the head losses encountered in the diffuser from wake and the recirculation flow caused by the impeller blade (∆qrw), the loss of part of the kinetic head at the exit of the diffuser (∆qex), and the loss of head from frictional forces

Figure 6-3: Characteristic dimensions in a centrifugal compressor.

| 210 |

| Off-Design Performance Characteristics | encountered in the vaneless diffuser space (∆qvs) and the vaned diffuser space (∆qvd) q oa = qia − ∆ q rw − ∆ q vs − ∆ qvd − ∆ qex

(6-12)

The overall adiabatic efficiency in an impeller is given by the following relationship: ηov =

qoa q tot

(6-13)

The individual losses can now be computed. Figure 6-3 shows the compressor dimensions needed for these computations. These losses are broken up into three categories: (1) external losses, (2) losses in the rotor, and (3) losses in the diffuser. EXTERNAL LOSSES External losses are divided into the following categories.

Clearance Loss When a fluid particle has a translatory motion relative to a noninertial rotating coordinate system, it experiences the Coriolis force. A pressure difference exists between the driving and trailing faces of an impeller blade caused by Coriolis acceleration. The shortest and least resistant path for the fluid to flow and neutralize this pressure differential is provided by the clearance between the rotating impeller and the stationary casing. With shrouded impellers, such a leakage from the pressure side to the suction side of an impeller blade is not possible. Instead, the existence of a pressure gradient in the clearance between the casing and the impeller shrouds, predominant along the direction shown in Figure 6-4, accounts for the clearance loss. Tip seals at the impeller eye can reduce this loss considerably. This loss may be quiet substantial. The leaking flow undergoes a large expansion and contraction caused by temperature variation across the clearance gap that affects both the leaking flow and the stream into which it discharges. δ

∆ qCL = 0.17 qth

(6-14)

| 211 |

| Centrifugal Compressors | where δ = S /b

where S = gap between the blades and the stationary shroud, and b = blade height.

Figure 6-4: Leakage affecting clearance loss.

| 212 |

| Off-Design Performance Characteristics | Disc Friction Loss This loss results from frictional torque on the back surface of the rotor, as seen in Figure 6-5. This loss is the same for a given size disc, whether it is used for a radial-inflow compressor or a radial-inflow turbine. Losses in the seals, bearings, and gear box are also lumped in with this loss, and the entire loss can be called an external loss. Unless the gap is of the magnitude of the boundary layer, the effect of the gap size is negligible. The disc friction in a housing is less than that in a free disc due to the existence of a “core” which rotates at half the angular velocity.

q df

 P  C f 1 + 2  P1   = 2 W   D    D   2  Z1  qth  2  2 1 −  h    U 2   De    D2  

(6-15)

where Cf = 0.0622 Re-0.2 (turbulent flow Re > 3 x 105) Cf = 2.67 Re-0.5 (laminar flow Re < 3 x 105)

Figure 6-5: Secondary flow at the back of the impeller to determine disk friction loss.

| 213 |

| Centrifugal Compressors | ROTOR LOSSES Rotor losses are divided into the following categories.

Shock in Rotor Losses This loss is due to shock occurring at the rotor inlet. The inlet of the rotor blades should be wedge-like to sustain a weak oblique shock, and then gradually expanded to the blade thickness to avoid another shock. Figure 6-6 shows the shock attached at the eye of the compressor for sharp edge blades. If the blades are blunt, a bow shock will result, as shown in Figure 6-6, causing the flow to detach from the blade wall and the loss to be much higher. W ∆ qsh = 1 −  sh  W1

γ −1   2     Psh  γ 2   −     − 1  2    − γ M P 1 ( )   rel   1   

(6-16)

where Wsh and Psh is the Relative Velocity and Static Pressure after the shock.

Figure 6-6: Shock wave at the impeller eye. The oblique shock shown is the desirable shock as opposed to the standing bow shock.

| 214 |

| Off-Design Performance Characteristics | Incidence Loss At off-design conditions, flow enters the inducer at an incidence angle that is either positive or negative, as shown in Figure 6-7. A positive incidence angle causes a reduction in flow. Fluid approaching a blade at an incidence angle suffers an instantaneous change of velocity at the blade inlet to comply with the blade inlet angle. Continuity of flow requires that the flow velocity remain unchanged. Hence, the relative velocity undergoes a change equal to W 2in. Separation of the blade can create a loss associated with this phenomenon, as shown in the following relationship: 2

[

]

W  2 ∆ qin =  1  1+ ε 2 (φ 0 − φ ) + m( M riφ − .111) U 2 

(6-17)

where Φ0 = maximum flow coefficient

Figure 6-7: Inlet velocity triangles at non-zero incidence to determine incidence loss.

| 215 |

| Centrifugal Compressors | φ=

Vin U2

ε=

2D2 De + D h

and m = 0 if MriΦ < 0.11 m = 1 if MriΦ > 0.11

Diffusion-Blading Loss This loss develops because of negative velocity gradients in the boundary layer and the recirculation of flow in the impeller, as shown in Figure 6-8. Deceleration of the flow increases the boundary layer and gives rise to separation of the flow. The adverse pressure gradient that a compressor normally works against increases the chances of separation and causes significant loss. The momentum thickness θ describes the buildup of the boundary layer, as shown:

2

θ=

1

R en −1

   A′     +2 4   n 

n n +1

n

  Wi  4 nn+ 2  n −1    1−     W2    Wi      1 −   W2   

(6-18)

where Wi and W2 are the relative velocities at the inlet and exit respectively.

Re =

(

)

1 2 2 Wi + W2 Dhy 2 ν

(6-19)

where the Hydraulic Diameter Dhy Dhy =

4 x Area Wetted Perimeter

π cos β i De2 − Dh2 π D2 b cos β e 2 Dhy = + π cos β i (De + Dh )+ 2 Z (De − Dh ) π D2 cos β e + Zb

(

| 216 |

)

(6-20)

| Off-Design Performance Characteristics | Where D2 = Outlet Diameter De = Eye Diameter at the Inlet Dh = Hub Diameter at the Inlet β= Blade angle at the inlet or the exit depending on the subscript and the constants A' and n are given as A' = 0.46 and n = 1, for laminar flow A' = 0.076 and n = 6, for turbulent flow Solidity factor is given by σ =

Z  D2 − De + 4b   π  2(D2 + De ) 

(6-21)

Diffusion blading loss can now be written as  2θσ   θσH 2 ∆ qdb =   1 +  sin β i   2 sin β i

  

(6-22)

The form factor H has a value ranging between 1.18 and 1.30.

Figure 6-8: Schematic of secondary flow circulation in a centrifugal impeller.

| 217 |

| Centrifugal Compressors | Skin Friction Loss Skin friction loss is the loss from the shear forces on the impeller wall caused by turbulent friction. This loss is determined by considering the flow as an equivalent circular cross section with a hydraulic diameter. The loss is then computed based on well-known pipe flow pressure loss equations, which, for the case of the centrifugal compressor, is given by the following relationship: 2

 L  1  Wave  π  D + Dh  ⋅ ⋅  D − e ∆ qSF = 4 K SF C F   ⋅ − b + 2 H 1  (6-23)  Dhy  2  U 2  8  2 2     

Where Dhy = Hydraulic Diameter Equation (6-20 KSF = 1.4 an empirical constant Cf = 0.0622 Re-0.2 (Re > 3 x 10.5) Cf = 2.67 Re-0.5 (Re < 2 x 10.5) Re = Wavg

W avg ⋅ Dhy ν 1

1 2 =  W1 2 + W2 2  2 

(

)

Dhy = Hydraulic Diameter Equation (6-20) KSF = 1.4 an empirical constant L = meridional path length Wavg = mean square relative velocity ν = Kinematic viscosity

STATOR LOSSES Recirculation and Wake Mixing Loss This loss is from the impeller blades, which causes a wake in the vaneless space behind the rotor, and recirculation of the flow. It is minimized in a diffuser, which is symmetric around the axis of rotation. This loss is further increased because of the backflow into the impeller

| 218 |

| Off-Design Performance Characteristics | exit of a compressor and is a direct function of the air exit angle. As the flow through the compressor decreases, there is an increase in the absolute flow angle at the exit of the impeller as seen in Figure 6-9. Part of the fluid is recirculated from the diffuser to the impeller, and its energy is returned to the impeller.

Figure 6-9: Diffuser recirculating loss.

∆ qrw = 0.02F 2 (tan α e )

0. 5

(6-24)

where the Diffusion Factor F is given by F = 1−

0.75 qth W2 + We W e  Z  D e  D  +2 e    1+ W2  π  D2  D2 

(6-25)

and αe = Exit air angle

| 219 |

| Centrifugal Compressors | Vaneless Diffuser Loss The flow in a vaneless diffuser follows a logarithmic spiral, as shown in Figure 6-10. The loss experienced in the vaneless diffuser results from friction and the absolute flow angle. ∆ qvs =

4 Cf 12

x

D2 1 x bi cos α avg

  D  1 .5  1 −  2     D3  

where  γ −1 2  C f = 0.592 Re −0.2 1 + M  2   V 3  D2   = V 2  D3 

Re =

−0. 45

1.25

V avg x length of passage v

M = Mach number

Figure 6-10: Flow trajectory in a vaneless diffuser.

| 220 |

(6-26)

| Off-Design Performance Characteristics | Vaned Diffuser Loss Vaned diffuser losses are based on conical diffuser test results. The flow in a vaned impeller is directed and proceeds from the inner diameter to the outside diameter in a much shorter distance as seen in Figure 6-11. They are a function of the impeller blade loading and the vaneless space radius ratio. They also take into account the blade incidence angle and skin friction from the vanes. ∆ qvd =

K vd 2

3  V   V   0.07 + 0.0076  s − 1   s    V4    U 2 

(6-27)

where Kvd accounts for the nonuniformity of the flow entering the diffuser. It is a function of impeller blade loading and the vaneless space radius ratio as found experimentally. 3

[

 V  1/ 3 Kvd = 1 + 0.44  1 − 3  1− (dd 43 − 1) V  4 

]

(6-28)

Figure 6-11: Flow in a vaned diffuser.

| 221 |

| Centrifugal Compressors | Exit loss. The exit loss assumes that one-half of the kinetic energy leaving the vaned diffuser is lost. ∆ qrw = 0.02F 2 (tan α e )

0. 5

(6-29)

Losses are complex phenomena and, as discussed here, are a function of many factors, including inlet conditions, pressure ratios, blade angles, and flow. Figure 6-12 shows the losses distributed in a typical centrifugal stage of pressure ratio below 2:1 with backward-curved blades at a constant speed. This figure is only a guideline. Note that the efficiency is highest near the surge point. The clearance loss is the highest near the maximum flow or choke point since the loss is a direct function of the flow while the disk friction loss is constant. The incidence loss is minimal near design as would be expected since the incidence angle is minimal at that point. The diffusion blading loss is largest near the surge point due to the fact that the pressure is highest at that point and there is a larger flow circulation and flow separation at that point. The skin friction loss is largest near choke since the loss is a direct function of the flow, which causes the relative velocity to be higher, and the skin friction loss is a direct function of the velocity squared. Recirculation loss is highest at the surge point since there is a flatter flow angle leaving the impeller. The vaneless diffuser, the vaned diffuser, and the exit losses are a direct function of the flow and are thus the highest at the choke point.

| 222 |

| Off-Design Performance Characteristics |

Figure 6-12: Losses in a centrifugal compressor.

| 223 |

7

Surge Control

INTRODUCTION Compressor surge is a phenomenon of considerable interest; yet it is not fully understood. It is a form of unstable operation and should be avoided. It is a phenomenon that, unfortunately, occurs frequently in the process industry, sometimes with damaging results. Surge has been traditionally defined as the lower limit of stable operation in a compressor, and it involves the reversal of flow. This reversal of flow occurs because of some kind of aerodynamic instability within the system. Usually, a part of the compressor is the cause of the aerodynamic instability, although it is possible for the system arrangement to be capable of augmenting this instability. Compressors are usually operated at a working line, separated by some safety margin from the surge line. Surge is often indicated by excessive vibration and an audible sound; however, there have been cases in which surge problems that were not audible have caused failures. Extensive investigations have been conducted on surge. Poor quantitative universality or aerodynamic loading capacities of different diffusers and impellers, and an inexact knowledge of boundary-layer behavior make the exact prediction of flow in turbomachines at the design stage difficult. However, it is quite evident that the underlying cause of surge is aerodynamic stall. The stall may occur in either the impeller or the diffuser. When the impeller seems to be the cause of surge, the inducer section is where the flow separation begins. A decrease in the mass flow

| 225 |

| Centrifugal Compressors | rate, an increase in the rotational speed of the impeller, or both can cause the compressor to surge. Whether surge is caused by a decrease in flow velocity or an increase in rotational speeds, either the inducer or the diffuser can stall. Which one stalls first is difficult to determine, but considerable testing has shown that for a low-pressure-ratio compressor, the surge usually initiates in the diffuser section. For units with single-stage pressure ratios above 3:1, surge is probably initiated in the inducer. Figure 7-1 shows the regions of surge initiation in a centrifugal compressor. Centrifugal compressor systems at numerous installations have suffered serious mechanical damage to compressor internals or to other parts of the piping systems as a result of operation in the surge condition, due to a reversal of thrust during surge, as shown in Figure 7-1. In other cases, compressors, usually smaller machines, have operated for long periods with intermittent or even continuous light surge without mechanical harm, although with significant impairment of aerodynamic performance. These conditions generally are the product of one or more of the following deficiencies:

Figure 7-1: Initiation of surge in centrifugal compressor.

| 226 |

| Surge Control | •

Poor matching of the compressor to the system’s requirements

•

Inappropriate compressor design

•

Inadequate anti-surge control system

•

Unfavorable arrangement of piping and process components of the system, which often can magnify surge

Because of the shape of their performance curves, the application of centrifugal compressors frequently is more complex than that of reciprocating machines. A centrifugal compressor driven by an electric motor offers an imposing challenge to the controls engineer when this constantspeed machine must accommodate a wide range of variables in its operating conditions. Centrifugal compressor applications continue to become more complicated with increases in the numbers of stages per casing, casings in tandem with a single driver, side-stream nozzles, higher pressures and speeds, more operating conditions for a given machine involving wider ranges of flows, molecular weights, and pressures. This ever-increasing complexity requires a better understanding of the causes of surge and its detrimental effects in order that adequate control systems may be applied. A typical performance map for a centrifugal compressor is shown in Figure 7-2. The operating line is the surge line modified by a safety margin to ensure trouble-free operation. Note that the total pressure ratio changes with flow, speed, molecular weight, suction pressure, and temperature. Also, note that operating at higher efficiency implies operation closer to surge. It should be noted here that total pressure increases occur only in the impeller. To make the curve general, the concept of aerodynamic speeds and corrected mass flow rates has been used. The surge line slope on multistage compressors can range from a simple single parabolical relationship to a complex curve containing several break-points or even “notches.” The complexity of the surge line shape depends on whether or not the flow limiting stage changes with operating speed from one compression stage to another; in particular, very closely matched stage combinations frequently exhibit complex surge lines. In the case of compressors with variable inlet guide vanes, the surge line tends to bend more at higher flows than with units that are speed controlled. Usually, surge is linked with excessive vibration and an audible sound; yet, there have been cases where surge not accompanied by audible sound has caused failures. Usually, operation in surge and, often, near surge is

| 227 |

| Centrifugal Compressors |

Figure 7-2: A typical compressor performance map.

accompanied by several indications, including general and pulsating noise level increases, axial shaft position changes, discharge temperature excursions, compressor differential pressure fluctuations, and lateral vibration amplitude increases. Frequently, with high-pressure compressors, operation in the incipient surge range is accompanied by the emergence of a low frequency, asynchronous vibration signal, which can reach predominant amplitudes, as well as excitation of various harmonics of blade passing frequencies. Besides the well-known effects of extended operation in surge (thrust and journal bearing failure, impeller rub), impeller hub and/or shroud failures resulting from severe stimulation at one of their natural response frequencies occasionally can be found. The antisurge methods discussed in this chapter, as well as most other proven surge prevention schemes, utilize conventional instrumentation to measure and control certain parameters (e.g., flow, pressure rise, etc.). Compressor and process applications vary so much that it could be difficult, if not impossible, to devise a universal standard antisurge control scheme. Each application must be evaluated in order to determine the

| 228 |

| Surge Control | required control functions. This requires not only a knowledge of instrumentation, but especially a good understanding of the compressor and load characteristics. The advent of the microprocessor has enabled consolidation of many control functions into one unit, but it has not decreased the amount of basic engineering that must be performed prior to programming. A reliable control which is not influenced by surge point prediction errors and other conditions, such as compressor fouling and gas variations (mole weight, temperature, pressure and RPM), certainly is desirable and should reduce the application engineering required for each machine. Efforts to develop such a control package continue, based on the theory that a measurable parameter exists in the region near surge. Additional research is needed to determine if the threat of surge can be detected early enough to facilitate control action in time to actually prevent it. The surge control system engineer has to take into account system valving, response rate, valve selection and location, bypass line sizing, and a host of other factors in order to develop a successful design. Dynamic simulation of compressor systems is a tool used to evaluate process design and system behavior at the design stage. It involves computer based mathematical modeling. The simulation procedure can provide a better understanding of how the system will behave during process operation, emergency shutdown, and process startups and shutdowns. Dynamic simulation also permits an evaluation of control hardware. It is a design tool that, if properly used, can avoid costly problems with compressor operation. DEFINITION OF SURGE The phenomenon of surge, as it pertains to a centrifugal compressor and its connected system, is an unstable condition resulting in flow reversals and pressure fluctuations in the system. This condition occurs when there is sufficient aerodynamic instability within the compressor that the compressor is unable to produce adequate pressure to deliver continuous flow to the downstream system. The system and compressor then interact, causing the surge conditions with large and sometimes violent flow oscillations in the system. Surge, then, is an overall system phenomenon and is not confined to the compressor only. Surge is the result of an excessive increase in the resistance of the system while the compressor is operating at a certain speed. The added resistance reduces the flow to an unstable level and causes the flow to reverse as shown in Figure 7-3. The flow initially reverses near the shroud,

| 229 |

| Centrifugal Compressors | and then, as the surge grows in strength, the entire flow is reversed. Alternatively, if the resistance is unchanged, but the speed is reduced appreciably, most systems will surge. Thus, surge occurrence depends on the type of system and the shape of the resistance curve.

Figure 7-3: Flow reversal in a centrifugal compressor as the compressor approaches surge.

R O TAT I N G S TA L L The stall of an element of a compressor stage may be compared with that of an airfoil of an airplane. The lift of an airfoil is related to the velocity of the airflow and the angle of attack (incidence), as shown in Figure 7-4. If the angle becomes excessive, the lift collapses, and a stalled condition results. The stall occurs because the air stream separates from the surface of the airfoil as seen in Figure 7-5.

| 230 |

| Surge Control |

Figure 7-4: Lift as a function of the angle of incidence.

Smooth flow

As angle of attack increases, break away starts

Past critical α break down of flow, lift force reduced

Figure 7-5: Schematic of the flow around an airfoil with increasing angle of attack of the airfoils.

| 231 |

| Centrifugal Compressors | Rotating stall (propagating stall) consists of large stall zones covering several blade passages and propagates in the direction of the rotor and at some fraction of rotor speed. The number of stall zones and the propagating rates vary considerably. Rotating stall can and does occur in centrifugal compressors. The propagation mechanism can be described by considering the blade row to be a cascade of blades (say, an inducer), as shown in Figure 7-6. A flow perturbation causes Blade 2 to reach a stalled condition before the other blades. This installed blade does not produce a sufficient pressure rise to maintain the flow around it, and an effective flow blockage or a zone of reduced flow develops. This retarded flow diverts the flow around it so that the angle of attack increases on Blade 3 and decreases on Blade 1. The stall propagates downward relative to the blade row at a rate about half the block speed; the diverted flow stalls the blades below the retarded-flow zone and unstalls the blades above it. The retarded flow or stall zone moves from the pressure side to the suction side of each blade in the opposite direction of rotor rotation, and it may cover several blade passages. The relative speed of propagation has been observed from compressor tests to be less than the rotor speed (40-75% of rotor speed).

Figure 7-6: Rotating stall.

| 232 |

| Surge Control | Observed from an absolute frame of reference, the stall zones appear to be moving in the direction of rotor rotation. This phenomenon can lead to inefficient performance and excitation of the resonant frequency of the inducer, thus leading to failure of that section. Rotating stall is accompanied sometimes by a pulsating sound and pressure pulsations that can be noted at both the inlet and exit sections of the impellers. The aerodynamic instability is brought about by flow reduction, which causes stalling of one or more of the elements of a stage or stages of the compressor. The stalling can occur at the inducer of the impeller, in the radial portion of the impeller, in the diffuser, or in the volute. A stall in one of these elements may not have sufficient effect to cause the stage to be unstable. In fact, several elements of a stage can stall without the entire stage stalling. However, if the stalling is of sufficient strength, the stage will become unstable, and this can lead to surge of the compressor.

CENTRIFUGAL COMPRESSOR OPERATION A typical centrifugal multi-stage compressor performance map used in the process industry is shown in Figure 7-7. The family of curves depicts the performance at various speeds, where N represents rpm. The ordinate may be polytropic head “H,” pressure ratio, discharge pressure, or sometimes differential pressure. The abscissa, usually the flow, is often shown as actual inlet volume per unit of time, such as acfm or icfm, where “a” is for “actual” or “i” for inlet.” It is important to understand that the inlet flow volume or capacity is based on a gas with a particular molecular weight, specific heat ratio, and compressibility factor at a pressure and temperature corresponding to the gas condition in the suction line to the compressor. If any of these parameters is changed, the performance map is no longer exactly valid. If the deviations are relatively small, the map still may be used with a fair degree of accuracy by making small adjustments. It is often impractical to make a separate set of curves for every small variation in these parameters. Since the impeller recognizes only actual volume, weight flow, such as pounds per minute or standard volume such as scfm, can only be used on the abscissa when there are absolutely no significant variations in gas conditions at suction. The line on the left represents the surge limit or “pumping limit” as it is often called. Operation to the left of this line is unstable, resulting in unsatisfactory performance, and is often harmful mechanically. Notice that

| 233 |

| Centrifugal Compressors | the surge flow increases as the speed increases. The surge line in many cases has an approximately constant value of Q/N. In many multi-stage units, and especially in units which are controlled by variable inlet guide vanes, the surge line curves bend dramatically at higher flows and speeds. On the other side of the map, the capacity limit or overload line is shown. Operation to the right of this line causes the head-producing capability of the machine to fall off very rapidly, and the performance is difficult to predict. The area to the right of this line is commonly known as “stonewall” or “choke.” Operating the machine in this region is usually harmless mechanically, although a few impeller failures have been ascribed to prolonged operation in stonewall. A compressor in stonewall is like an orifice at critical flow. It means that a gas velocity somewhere in the compressor, usually at the impeller inlet (inducer), has reached sonic velocity. The capacity limit line also is a line of roughly constant Q/N values. Plant operators sometimes get the terms “surge” and “stonewall” confused, because, presumably, machine performance is seriously impaired in either case. However, it must be remembered that the two phenomena are completely different and create completely different operational problems.

Figure 7-7: Typical centrifugal compressor performance map.

| 234 |

| Surge Control | Figure 7-8 is the same as Figure 7-7, except that some points have been labeled A, B, C, and D, and three typical system-operating curves have been plotted. Terms frequently used to define performance are “stable range” and “percent stability.” The rated stable range is generally taken as QA-QB, where QA is the design or rated point, and QB is the surge point along the 100% speed line. The percent stability, expressed as a percentage, is : Stability (%) =

QA − QB QA

(7-1)

The use of these terms is somewhat misleading since operation between the surge line and the capacity limit line is stable and predictable. Notice that the span between the capacity limit and the surge limit decreases as speed is reduced. It is also important to observe that, at lower speeds, the compressor does not surge until it reaches lower inlet flows.

SURGE CURVE
– CONSTANT PRESSURE SURGE CURVE  – PARTIALLY FRICTION – PARTIALLY CONSTANT SURGE CURVE  – PREDOMINANTLY FRICTION

Figure 7-8: Performance map with system curves and surge.

| 235 |

| Centrifugal Compressors | Rise-to-surge (RTS) is another term often used to assess the capability of a compressor to recover from a disturbance in the system. A high RTS means that the machine can accommodate a modest increase in discharge pressure with little change in inlet flow. With a low RTS, the curve is relatively flat, and there could be a large reduction in inlet flow corresponding to a very small increase in discharge pressure. Pressure RTS is expressed as a percentage: RTS =

PB − PA PA

(7-2)

RTS also may be expressed in terms of head. When using the RTS, it is essential to state whether it is based on pressure or head. Pressure RTS is not necessarily numerically equal to head RTS, however, it signifies a similar change. The three representative system curves in Figure 7-8 need little explanation. The shape of these curves is governed by the amount of friction, fixed pressure drop, or pressure control in the particular system external to the compressor through which the flow is being pumped. However, it should be noted that, to follow any one of these system curves, the speed must be changed, which in turn, changes the flow. In the System 2 and 3 curves, the head is also changed. These statements are valid only if no changes within the system itself occur. Changing the setting of a control valve, adding another piping loop, changing the catalyst level in a reactor, or other process change will modify the system curve. These obvious relationships are mentioned only to emphasize the operational difference between constant speed and variable speed machines. At constant speed, something in the system itself must be changed to relocate an operating point if the gas composition and compressor hardware stay the same. A typical surge cycle is represented in Figure 7-8 by the circuit between Points B, C, D, and back to B. Suppose that events gradually or suddenly take place to establish operation at Point B. Here, the pressure in the system is equal to the output pressure of the compressor. If any transient variable caused operation to shift slightly to the left, reverse flow would begin because the compressor discharge pressure would be less than the pressure already in the system. In order for reverse flow to occur, the flow delivered by the compressor must be reduced to zero at Point C, which corresponds to a certain value of head or pressure, called “shut-off head.” When the system pressure had blown itself down to the

| 236 |

| Surge Control | compressor’s shut-off head capability at Point C, the machine would begin to pump anew, since the flow volume requirement of the compressor would have been satisfied by the backflow gas. Now that the compressor had some gas to pump, operation would immediately shift to the right in an approximately horizontal path to Point D on the speed line at a discharge pressure about equal to the shut-off head. With the compressor now delivering flow in the forward direction, pressure would build up in the system, and operation would follow the characteristic speed curve back to Points B and C. The cycle would repeat itself again and again unless the original cause of the surge was corrected or other favorable action, such as increasing the speed or opening the flow valve, was taken. To a large extent, the frequency of the surge cycle varies inversely with the volume of the system. For example, if the piping contains a check valve located near the compressor discharge nozzle, the frequency will be correspondingly much higher than that of a system with a large volume in the discharge upstream of a check valve. The frequency can be as low as a few cycles per minute up to 20 or more cycles per second. Generally speaking, if the frequency is higher, the intensity of surge is lower. The intensity or violence of surge tends to increase with increased gas density, which is directly related to higher molecular weights and pressures and lower temperatures. Higher differential pressure generally increases the intensity. The location of Points C and D in Figure 7-8 were randomly selected for illustration purposes, so the values of head and flow may not be realistic. Often flow rates oscillate continuously at large amplitudes, even when a compressor is operating in its “stable” region. The time average operating point must lie at a significantly larger mass flow than the instantaneous stability limit.

EFFECTS OF INTERNAL LOSSES ON CHARACTERISTIC CURVE Details of the internal factors that influence the shape of the characteristic performance curve and the location of the surge point are given in detail in chapter 4. From the foregoing description of surge, it can be seen that the shape of the head-capacity characteristic curve is fundamentally responsible for the location of the surge point at a certain speed. On the right side of the performance map, the slope of the curve is negative. As inlet flow is reduced, the slope becomes less negative until it reaches

| 237 |

| Centrifugal Compressors | essentially zero at the surge point. As flow is reduced further to the left of the surge point, the slope becomes increasingly positive. It should be pointed out that neither zero slope nor a positive slope are absolute criteria for surge; however, in most cases, it is recommended that surge controls be set at a point where the slope becomes zero. As might be expected, there are very definite interdependent effects resulting from the individual geometries of the principal components of the compressor as described in chapter 4. Such pertinent variables as the impeller configuration and blade angle, inlet guide vane angle, diffuser size and shape, etc. can be adjusted by the machine designer for optimum performance under a specified set of operating conditions. The vectorial summation of all the gas velocity vectors associated with the several components of the machine govern the theoretical head-capacity relationship. Ideal or theoretical head is a straight line when plotted against inlet capacity, and the overall geometry dictates the slope of the line as seen in Figure 7-9, depending on the exit blade angle. As usual with any machine, inevitable losses, which cannot be recovered, intervene to produce a significant difference between the theoretical and actual output. Mechanical losses such as those incurred in a journal or thrust

Figure 7-9: Head flow rate characteristics for various outlets blade angles.

| 238 |

| Surge Control | bearing affect the power input to the machine but do not influence the shape of the head-capacity curve. Aerodynamic losses that do influence the shape of the curve consist mainly of wall friction, fluid shear, seal losses, recirculation in flow passages, and diffusion blading losses (separation losses). Diffusion blading losses are the result of expansion, contraction, and change of direction and recirculation associated with flow separation, eddies, and turbulence. Figure 7-10 indicates graphically how the combined losses reduce the linear theoretical head to the actual curve. The geometry and dimensions of the elements within the machine can be originally designed to move the rated operating point to a more desirable location on the curve from the viewpoint of a certain variable, but this relocation will usually create one or more disadvantages from another viewpoint. For instance, if the design point is located as shown, a move slightly to the left would gain some efficiency and head output and might result in eliminating one impeller from a multistage application; however, the cost could be an objectionable loss in the stable operating range. A move to the right would produce opposite results, which could be desirable in some situations. Nevertheless, the reduced capacity overload limit or proximity to stonewall could

Figure 7-10: Typical compressor head losses for backward curve blades.

| 239 |

| Centrifugal Compressors | have significant disadvantages. To summarize, the combined losses force the actual curve to go from a negative to a positive slope as flow decreases; that is, the actual curve always has a “hump” in it. Changes in machine geometry can relocate the crest of the hump, change the breadth of the hump, and alter the severity of the curve to some extent. However, the hump cannot be eliminated. The surge point is located at or very near the point of zero slope on the compressor’s characteristic curve. Friction losses can be reduced somewhat by improving surface finishes and enlarging passages. Blade diffusion losses may sometimes be mitigated by further streamlining of flow passages and assuring that there are no abrupt changes in the flow area. These techniques will improve efficiency and tend to reduce the surge point; however, they also cost money, and there is a point of diminishing marginal returns. The phenomenon of slip, which is due to the build-up of boundary layer on the vanes of the impeller, occurs in all centrifugal machinery. Boundary layer controls have been used to reduce slip and increase the flow stability of the impeller, as detailed in chapter 4. DESIGN FACTORS AFFECTING SURGE Some elaboration on the principal machine design factors that affect surge will now be covered. A greater number of impellers in a given casing will tend to reduce the stable range. This effect is presented in Figure 7-11. Manufacturers have found it physically and economically impractical to design each stage of a multi-stage machine to be exactly the optimum size. As the pressure builds up across the stages, the volumetric flow is reduced. Ideally, then, each succeeding wheel should be proportionately smaller with respect to flow passage area. For example, the width of the flow exit at the tip of the impeller should actually get smaller and smaller. It would be virtually impossible to achieve perfect proportionality. This lack of optimum proportionality aggravates the stability problem. Some larger machines are custom built, and, therefore, better proportionality can be attained. Mechanical limitations, such as the axial length of the shaft between bearings and its relation to critical speed, may dictate the axial width of the impellers through the hub and the impeller spacing. The shaft diameter may be constrained by torque and/or critical speed considerations, and the inside diameter limits of the impeller hub may be restricted accordingly. Practical economics and standardization occasionally will limit the number of types and sizes of wheels in a given machine. For example, in a six-wheel machine, there might be only two

| 240 |

| Surge Control |

Figure 7-11: Effect of number of impellers on stability and capacity limit.

diffident wheel designs. Therefore, where the first impeller might be operating at optimum conditions, subsequent impellers of the same type and size usually could be operating at less favorable flow conditions, owing to the volume reduction across the stages. There are a great many other considerations which affect the overall stability. The relationship in Figure 7-11 is typical only, and the design can be changed to improve stability to some extent. However, more often than not, it will be at the sacrifice of some other important performance factor. The diminishing effect of the number of wheels on the location of the capacity limit is also shown in Figure 7-11. Just as the greater number of wheels in a casing reduces the stability, so does the number of sections of compression, or the number of casings in series. This effect is illustrated by the sample calculation shown in Figure 7-12. It can be seen that individual percent stability of the HP casing is better than that of the LP casing. However, as LP casing flow is reduced, the HP casing surges before the LP casing because of increase in intercasing pressure, which decreases acfm to that of the HP casing. Overall stability is less than the individual stability of either casing.

| 241 |

| Centrifugal Compressors |

Figure 7-12: Typical overall stability vs. individual stability for two casings operating in series.

The information and curves presented thus far have been, for the most part, based on impeller with backward leaning impeller blades. Figure 713 depicts the effects of impeller blade angle on the stable range, and shows the variance in steepness of the slope of the head-flow curve. The three curves are based on the same speed and show actual head. The relationship of ideal or theoretical head to inlet flow for different blade angles would be represented by straight lines. For backward leaning blades, the slope of the line would be negative. The line for radial blades would be horizontal. Forward leaning blades would have a positively sloped line. For the average petrochemical process plant application, the compressor industry in the USA commonly uses a backward-leaning blade with an angle (β2) of between 55° and 75° (or backward leaning angle of 15°- 35°), because it provides a wider stable range and a steeper slope in the operating range. This impeller design has proven to be about the best compromise between pressure delivered, efficiency, and stability. Forward leaning blades are not commonly used in compressor design, since the high exit velocities lead to large diffuser losses. A plant air compressor

| 242 |

| Surge Control |

Figure 7-13: Effect of blade angle on stability.

operating at steady conditions from day to day would not require a wide stable range, but a machine in a processing plant can be the victim of many variables and upsets. So more stability is highly desirable. Actually, the lower curve in Figure 7-13 appears to have a more gentle slope than either the middle or upper curve. This comparison is true in the overall sense, but it must be remembered that the normal operating range lies between 100% Q and Q at surge, plus a safety margin of, usually, about 10%. The right-hand tail ends of all three curves are not in the operating range. The machine must operate with a suitable margin to the left of where these curves begin their steep decent or tail-off, and, in the resultant operating range, the curve for backward leaning blades is steeper. This steeper curve is desirable for control purposes. Such a curve produces a meaningful Delta-P for a small change in Q. The blade angle by itself does not tell the overall performance story. The geometry of other components of a stage will contribute significant effects also. Most centrifugal compressors in service in petroleum or petrochemical processing plants use vaneless diffusers. A vaneless diffuser is

| 243 |

| Centrifugal Compressors |

Figure 7-14: Design condition velocity triangles.

generally a simple flow channel with parallel walls and does not have any elements inside to guide the flow. Figure 7-14 shows velocity diagrams at the eye and exit of an impeller and illustrates the trajectory a particle in the gas flow would take through a vaneless diffuser at the design condition (compressor rated point). When the inlet flow to the impeller is reduced while the speed is held constant, there is a decrease in Vr2 and α2. As α2 decreases, the length of the flow path spiral increases. The effect is shown in Figure 7-15. If the flow path is extended enough, the flow momentum at the diffuser walls is excessively dissipated by friction and stall. With this greater loss, the diffuser becomes less efficient and converts a proportionately smaller part of the velocity head to pressure. As this condition progresses, the stage will eventually stall. This could lead to surge. Vaned diffusers are used to force the flow to take a shorter, more efficient path through the diffuser. There are many styles of vaned diffusers, with major differences in the types of vanes, vane angles and contouring, vane spacing, etc. Commonly used vaned diffusers employ wedge-shaped

| 244 |

| Surge Control |

Figure 7-15: Flow trajectory in vaneless diffuser.

vanes (vane islands) or thin-curved aerodynamic vanes. The latter type of vane is illustrated in Figure 7-16. In high head stages there can be two to four stages of diffusion. The first stage is usually the vaneless space, which acts as a vaneless diffuser to decelerate the flow, followed by one to three levels of vaned blades in order to prevent build-up of boundary layer, thus causing separation and surging of the unit. Figure 7-16 indicates the flow pattern in a vaned diffuser. The vaned diffuser can increase the efficiency of a stage by two to four percentage points, but the price for the efficiency gain is generally a narrower operating span on the headflow curve with respect to both surge and stonewall. Figure 7-16 also shows the effect of off-design flows. Excessive positive incidence at the leading edge of the diffuser vane occurs when the flow is too small at reduced flow, and this condition brings on stall. Conversely, as flow increases beyond the rated point, excessive negative incidence can cause stonewall. Despite its narrowing effect on the usable operating range on the characteristic curve, the vaned diffuser has its application in situations where efficiency is of utmost importance. Although seldom

| 245 |

| Centrifugal Compressors |

Figure 7-16: Flow in a vaned diffuser.

Figure 7-17: Effect of guide vane setting (stationary or variable).

| 246 |

| Surge Control | used, movable diffuser vanes or vane islands can be used to alleviate the shock losses at off-design conditions. However, as the adjusting mechanisms required are quite complicated, they generally are applied only to single-stage machines. It should be noted that the illustrations of the flow paths in Figures 7-14 through 7-16 are somewhat simplistic. Each flow path is indicated by a single streamline. The actual flow field is far more complex, with flow separation and recalculation present. Nevertheless, these figures should help with a practical understanding of the effects of changes in velocity triangles. Chapter 4 presents a detailed discussion of the flow in the impeller. Stationary guide vanes direct the flow to the eye of the impeller in an orderly fashion. Depending upon the head requirements of an individual stage, these vanes may direct the flow in the same direction as the rotation or tip speed of the wheel, an action known as positive pre-swirl. This is usually done to reduce the relative mach number entering the inducer in order to prevent shock losses. This, however, reduces the head delivered but improves the operating margin. The opposite action is known as counter-rotation or negative pre-swirl. This increases the head delivered but also increases the inlet relative mach number. Negative pre-swirl is rarely used, since it also decreases the operating range. Sometimes the guide vanes are set at zero degrees of swirl; these vanes are called radial guide vanes. The theoretical effect of these settings is shown in Figure 7-17. Movable inlet guide vanes are occasionally employed on single-stage machines, or on the first stage of multi-stage compressors driven by electric motors at constant speed. The guide vane angle can be manually or automatically adjusted while the unit is on stream to accommodate offdesign operating requirements. Because of the mechanical complexity of the adjusting mechanism and physical dimensional limitations, the variable feature can only be applied to the first wheel in almost all machine designs. Hence, the effect of changing vane angle is diluted in the stages downstream of the first stage. Although the flow to the entire machine is successfully adjusted by moving the first stage vanes, the remaining stages must pump the adjusted flow at a fixed guide vane angle. A typical performance map of a constant-speed compressor with variable inlet guide vanes is represented in Figure 7-18. Incidentally, a butterfly throttle valve in the suction line to the machine will produce nearly the same effects as moving the first stage guide vanes. However, throttling is not as efficient as moving the guide

| 247 |

| Centrifugal Compressors | vanes, so that in many cases the added cost of the movable vane mechanism can be justified by horsepower savings.

Figure 7-18: Constant speed machine with variable inlet guide vanes.

EFFECTS OF GAS COMPOSITION Figure 7-19 shows the performance of an individual stage at a given speed for three levels of gas molecular weight. The heavy gas class includes gases such as propane, propylene, and standardized refrigerant mixtures. Air, natural gases, and nitrogen are typical of the medium class. Hydrogen-rich gases found in hydrocarbon processing plants are representative of the light class. The following observations can be made with respect to the curve for heavy gas:

| 248 |

| Surge Control | •

The flow at surge is higher.

•

The stage produces slightly more head than that corresponding to medium gas.

•

The right-hand side of the curve turns downward (approaches stonewall) more rapidly.

•

The curve is flatter in the operating stage.

It is the last point that often presents a problem to the designer of the antisurge control system. It should be noted that the flatness gets worse as stages are added in series. Since the RTS is small, there is a large change in flow corresponding to a small change in head. The control system, therefore, must be more responsive. It should be obvious that curves for lighter gases have a more desirable shape.

Figure 7-19: Effect of gas composition on the characteristics of a centrifugal impeller.

| 249 |

| Centrifugal Compressors | External Causes and Effects of Surge The following are some of the usual causes of surge that are not related to machine design: •

Restriction in suction or discharge of system

•

Process changes in pressure, temperatures, or gas composition

•

Internal plugging of flow passages of compressor (fouling)

•

Inadvertent loss of speed

•

Instrument or control valve malfunction

•

Malfunction of hardware such as variable inlet guide vanes

•

Operator error

•

Maldistribution of load in parallel operation of two or more compressors

•

Improper assembly of compressor, such as a mispositioned rotor

•

Excess extraction of gases between sections

The effects of surge can range from a simple lack of performance to serious damage to the machine or to the connected system. Internal damage to labyrinths, diaphragms, thrust bearing, and rotor can be experienced. There have been many cases of covered impeller shroud rub and destruction of shroud and diaphragm caused by violent surge. In the case of an open-faced impeller, results in excitation of the inducer causes failure of the blades near the eye of the impeller. Surge often excites lateral shaft vibration and could produce torsional damage to such items as couplings and gears. Externally, devastating piping vibration can occur, causing structural damage, shaft misalignment, and failure of fittings and instruments. The effects of the size and configuration of the connected system, as well as different operating conditions, on the intensity of surge can be astonishing. For example, a compressor system in a test set-up at the factory may exhibit only a mild reaction to surge. At the installation, however, the same compressor with a different connected system may react in a tumultuous manner. Surge can often be recognized by check

| 250 |

| Surge Control | valve hammering, piping vibration, noise, wriggling of pressure gauges or an ammeter on the driver, or lateral and/or axial vibration of the compressor shaft. Mild cases of surge sometimes are difficult to discern.

PRACTICAL ASPECTS OF SURGE CONTROL SYSTEMS SURGE DETECTION AND CONTROL A surge control system consists of a set of detection devices (located around the compressor) that anticipate the surge, and control devices, which act to prevent surge from occurring. Surge detection devices may be broken into two groups: static and dynamic. Static surge detection devices are those devices that attempt to avoid stall and surge by the measurement of certain compressor parameters and thereby ensure that a predefined value is not exceeded. When a parameter meets or exceeds the limit, some control action is taken. To this date, static surge detection devices are the most widely used. It is important to note that surge or stall itself is not measured by present day instrumentation. This is the reason for the introduction of dynamic surge control concepts, which measure the actual onset of surge. It is probably the dynamic detection device that will meet the requirements and hopes of many engineers for a control device that can detect incipient surge and, hence, prevent its further development. However, much research on this type of device still is needed. Before energy became such a critical factor, the allowance margin between actual physical surge and control operation was adjusted as required for all of the practical problems encountered. Very little motivation existed to narrow the margin. Even with the energy waste associated with wide allowance margins, very few operators would be willing to sacrifice reliability to narrow the margin. The challenge to the modern designer is to save energy without sacrificing control reliability. Control systems presently are modern, microprocessor-based, digital systems. However, there are many analog systems in service that will continue to be used. Static anti-surge control systems must be used to prevent surge by maintaining the minimum flow at a value safely away from the capacity at

| 251 |

| Centrifugal Compressors | which surge occurs. This is accomplished by allowing some gas to recirculate through an anti-surge valve and recycle line from the compressor discharge back to its inlet. For air, and occasionally for other contaminant-free gases such as O2 and N2, the anti-surge valve vents gas to the atmosphere for surge prevention. Figure 7-20 illustrates surge, and control points B and C, and process operating points A and D at 100% discharge pressure. The anti-surge control maintains 80% flow through the compressor, even though the process flow requirement is less than 80%. For example, if the process requires only 60% flow (Point A), the antisurge control maintains 20% flow through the recycle line. Flow through the compressor is equal to the process flow (60%) plus the recycle flow (20%), or 80%, as illustrated by Figure 7-20. Recycle flow would, of course, be zero whenever the process was using 80% flow or more. The highenergy cost for compression would justify minimizing recycling, when operating at reduced throughput, by using a variable speed drive which, in turn, complicates the anti-surge control system.

Figure 7-20: Process operating points, flow 60% (A) and flow 100% (D), surge point B, and control points C.

| 252 |

| Surge Control | System surge lines are affected by changes in molecular weight, compressor fouling, changes in inlet conditions, and speed of the unit. Therefore, in many cases, the preset value need not be exceeded before the compressor will surge. Optimum design entails considerably more than selecting primary flow elements, control valves, transmitters, and panel instruments and integrating them into a control scheme based on a well-defined standard for designing antisurge control systems. Unfortunately, there is no so-called “universal best method” of anti-surge control. Before a control scheme is developed and components are selected, the operating requirements of the compressor and process must be known and evaluated. There are many variations in compressor and process characteristics; therefore, a control scheme that is ideal for one application might be far from adequate for another. M E T H O D S O F C A PA C I T Y C O N T R O L OF COMPRESSORS The two types of turbo compressors are centrifugal and axial flow. Methods of capacity control vary, depending on the type of compressor and its driver. This discussion will be limited to the methods outlined in Table 7-1, which are the most commonly used surge detection devices. Type Driver Constant Speed Variable Speed Constant Speed

Type Compressor Centrifugal Centrifugal Axial Flow

Method of Capacity Control Suction Throttle Variable Speed Variable Angle Stator Vanes

Table 7-1: Surge control system.

Typical performance curves and surge lines for the above are shown in Figure 7-21(a). It must be emphasized that the word “typical,” used to describe these curves, means just that. The actual surge line of many compressors is significantly different from the typical surge line illustrated. For example, instead of the typical shape (Fig. 7-21(b)), the actual surge lines of many centrifugal compressors are shaped more like the surge line of an axial flow machine as shown in Figure 7-21(c) or 7-21(d). Here the surge line bends to higher flow, which is a common signature of inlet vane control.

| 253 |

| Centrifugal Compressors |

Figure 7-21 (a): A centrifugal compressor-inlet-throttling.

Figure 7-21 (b): Centrifugal compressor variable speed.

| 254 |

| Surge Control |

Figure 7-21 (c): Axial flow compressor variable angle stator vanes.

Figure 7-21 (d): Axial flow compressor variable speed.

| 255 |

| Centrifugal Compressors | C O N S TA N T S P E E D C E N T R I F U G A L A N T I S U R G E CONTROL For this type machine, Figure 7-21(a) shows that surge occurs at a constant inlet volume flow rate, regardless of the inlet pressure. Therefore, inlet volume flow rate would be an appropriate measurement for antisurge control. Instruments can be used as shown schematically in Figure 7-22 to solve Equation 7-1. Q =C

∆h Pi

(7-3)

where Q = Volume Flow Rate at the Inlet ∆h= Differential head in flow meter. Pi = Inlet Pressure C = Proportionality Constant Temperature compensation is not used in the control scheme, therefore, if inlet temperature increases, the controlled flow will be lower than the set point. Actual flow can be calculated by Equation 7-4 for changes in Temperature and Molecular Weight shown below: Qactual = Qsetpoint

Tdesign MWactual Tactual MWdesign

(7-4)

A decrease in mole weight will also result in a lower controlled flow, in accordance with Equation 7-4. These flow measurement errors are actually desirable, because gas temperature or mole weight variations also cause the surge point of most compressors to shift in the same direction. The surge point usually shifts more than the control point; therefore, temperature or mole weight variations normally change the safety margin (distance between surge and controlled flow points). Small to moderate gas variations normally can be accommodated by a set point, which provides an adequate safety margin at the lowest gas temperature or highest mole weight condition. To handle extreme variations, it may be necessary to change the controller setpoint. The flow measurement in Figure 7-22 is pressure compensated. Therefore, the system will control a constant volume flow, regardless of the inlet pressure. This feature is needed because surge occurs at the same volume flow rate, regardless of the Inlet pressure.

| 256 |

| Surge Control |

Figure 7-22: Anti-surge control scheme, based on inlet volume flow.

The inlet flow control system shown in Figure 7-22 includes some undesirable features. A relatively long inlet line is required to provide sufficient straight run of pipe between the throttle valve and flow element. Also, the flow element (FE) is relatively expensive, and it produces a pressure loss that results in slightly greater operating costs. These disadvantages can be eliminated by an antisurge control scheme based on measurements of power (J) and discharge pressure (Pd), as shown in Figure 7-23. Anti-surge control by this method is facilitated by the following characteristics of a constant speed, inlet throttled compressor: 1.

At the surge point, the following parameters are unaffected by suction pressure variations: •

Volume flow (acfm, m3/hr.)

•

Head, polytropic (Hp)

•

Ratio of compression (Rc); Rc = Pd /Pi

2. Power is proportional to weight flow (W), which in turn is proportional to inlet pressure (Pi) 3.

Pd is proportional to Pi; i.e., Pd = Rc x Pi

| 257 |

| Centrifugal Compressors |

Figure 7-23: Centrifugal compressor antisurge control J vs.Pd.

Figure 7-24: Wet weight flow vs. brake horsepower and discharge pressure.

| 258 |

| Surge Control | Based on the above, surge is a straight-line relationship of Pd vs. W, or Pd vs. J, as illustrated in Figure 7-24. As with the inlet flow system previously described, this J vs. Pd system also is subject to some error if inlet temperature or mole weight changes. For example, a temperature decrease at constant inlet pressure will cause surge to occur at higher volume flow, higher discharge pressure, and higher power. VA R I A B L E S P E E D C E N T R I F U G A L A N T I S U R G E CONTROL The “Flow/Delta-P” antisurge control system is widely used with variable speed centrifugal compressors. The name is derived from the fact that control action is based on measurements of compressor inlet flow and differential pressure (Pd – Pi). Two Flow/Delta-P type control schemes are illustrated by Figures 7-25 and 7-26. The controller setpoint in the typical system shown in Figure 7-25 is manipulated by the Pressure Differential “Pd – Pi” signal, and the flow signal is the process input to the controller. In the alternate system, shown in Figure 7-26, the process signal to the controller is a value, which represents the compressor operating point (Pd – Pi vs. flow signal), and a constant controller set point is used.

Figure 7-25: Anti-surge control based on typical flow/differential pressure system.

| 259 |

| Centrifugal Compressors | The theory of the Flow/Delta-P system is based on the following flow element and compressor surge line characteristics. •

Differential (h) produced by the flow element is proportional to flow squared.

•

Pressure rise (Pd – Pi) produced by the compressor is proportional to rotating speed squared.

•

Flow rate at which surge occurs is proportional to rotating speed.

For example, 100% pressure rise (Pd – Pi) and 100% flow are used for the surge point at 100% speed. Thus, at 90% speed; % Pd – Pi = 100 (90/100)2 = 81%; % flow signal = 100 (90/100) = 90 %. Therefore, at 90% speed the setpoint will be 81%, and flow will be controlled at 90%. The Flow/Delta-P type control scheme is widely used and, in many cases, is the most suitable method of antisurge control for centrifugals. Often, however, the system has to be modified to suit particular needs.

DISCHARGE VOLUME FLOW (P AND T C O M P E N S AT E D ) / A N T I S U R G E C O N T R O L , VA R I A B L E S P E E D C O M P R E S S O R A pressure and temperature compensated discharge volume flow antisurge control system is shown in Figure 7-27. This particular centrifugal compressor is operated in series with an axial flow compressor. Because of very large interstage coolers, separators, piping, and space limitations, it was deemed impractical to use an inlet flow element, so alternate methods were investigated. Calculated values of various parameters (i.e., weight flow, inlet and discharge volume flow, pressure ratio, power, etc.) at the surge point were examined for various operating speeds. Calculated discharge volume flows for all surge points, from 75% to 95% speed, varied less than ± 1%. Thus, a careful examination of various calculated parameters indicated discharge volume flow to be not only the most practical but also a very effective variable to use for surge prevention. Field surge tests of the actual machine verified the calculated predictions. Surge and control lines are shown in Figure 7-28 in the compressor performance curve. The flow measurement is temperature compensated; therefore, the control line does not change with temperature variations. In this case, temperature compensation decreased

| 260 |

| Surge Control |

Figure 7-26: Alternate flow differential pressure type anti-surge control system.

Figure 7-27: Discharge volume flow (P&T compensated), antisurge control variable speed centrifugal.

Figure 7-28: Surge and control lines discharge volume flow (P&T compensated) antisurge control.

| 261 |

| Centrifugal Compressors |

Figure 7-29: Discharge flow antisurge control variable speed centrifugal compressor.

Figure 7-30: Surge control lines, discharge flow control, variable speed centrifugal.

| 262 |

| Surge Control | the control error at speeds above 95%, but slightly increased the error at speeds below 95%. Because of the benefit at normal speeds, temperature compensation was justified. With this system, the inlet flow element was eliminated, the discharge flow element was smaller and less expensive, the pressure loss due to the flow element was less significant at the discharge, and a conventional flow computing system was used.

DISCHARGE VOLUME FLOW (NONC O M P E N S AT E D ) A N T I S U R G E C O N T R O L , VA R I A B L E S P E E D C O M P R E S S O R A similar system, but without pressure or temperature compensation, is shown in Figure 7-29. The surge and control line curves for the compressor and control system are shown in Figure 7-30, the control line is about 15-17% away from the surge line. The system is very simple, yet it provides very effective control. Several things count for the surge and control lines having a similar slope over the normal 85 to 105% speed range: 1.

Flow units at normal conditions are used. Q =C

∆h x P T

(7-5)

In this case, the system measures only the differential, head (h) across the orifice. The controller maintains “∆h” constant at its setpoint value. Therefore, the controlled flow varies in proportion to P/T. Discharge pressure and temperature decrease with speed. On a percentage basis, discharge pressure decreases much faster than temperature, so the slope of the control line is primarily a function of Pd. 2. The capacity of the compressor relative to its compression ratio influences the slope of the surge line. In this case, the relationship is such that the surge line slope is very similar to the flow vs. Pd slope of the control line (in the normal speed range). By comparison, the compression ratio of this machine is approximately four times that of the compressor used for the preceding example, but the flow capacity is lower by a factor of eight. The antisurge control scheme shown in Figure 7-29 is not always best suited for all variable speed machines,

| 263 |

| Centrifugal Compressors | but, because of its simplicity and effectiveness, it should always be considered.

F L O W / D E LTA - P ANTISURGE CONTROL The results of a simple Flow/Delta-P type surge control system are illustrated for normal and startup conditions in Figures 7-31. The startup gas is more than twice as heavy as the normal gas. This illustration shows the effect that a heavier gas has on the compressor plus the inherent selfcompensating effect of the Flow/Delta-P control scheme. At any given speed, the machine produces a much higher discharge pressure on the heavier gas and, in the example shown, at 70% speed, it produces more pressure than on normal gas at 100% rpm. The safety margin increases due to a heavier gas or lower temperature. The Flow/Delta-P control is also very practical and effective for a constant speed, inlet throttled compressor. Depending on how the system is calibrated, errors due to inlet pressure variations will be small or nonexistent. If the system calibration is such that the control line originates at zero flow and Delta-P (Pd -Pi), the error will be zero. This is explained by the following characteristics of the flow element and a constant speed compressor: 1. Inlet pressure variations do not affect the volume flow or the compression ratio at which surge occurs. The control, therefore, must maintain a constant volume flow regardless of inlet pressure. 2. At a constant volume flow, the differential head produced by the orifice varies directly with pressure.

SYSTEM PIPING AND C O N F I G U R AT I O N A S P E C T S O N S U R G E Piping and valving are important considerations in surge control system design. A very fundamental and important aspect is the sizing of the bypass line. When startup, shut-down, and normal operation are considered, there are a number of variations. Field experience has shown this straightforward task to be the cause of many field problems because the line was too small. Another consideration in line sizing is the dynamic or momentary requirements in the limiting type of control, such as the

| 264 |

| Surge Control |

Figure 7-31: Surge and control lines, flow delta-P, anti-surge control, and normal gas.

surge control system. Because steady state criteria would indicate a line sized to pass only the surge plus margin flow, there is a temptation to utilize the apparent economics of a smaller line size. In actual practice, the line should handle full compressor flow with a minimum pressure drop. Correct surge control valve selection and sizing is also of critical importance for achievement of good control sensitivity and a rapid response in the vicinity of surge. While other components (i.e., actuator, controller) also affect the overall control system response, valve sensitivity and response have the major impact. The control valve is a key link in the control chain. In most cases, the valve should have linear flow and fast response. There is very little benefit in having a highly sophisticated controller set point computer coupled to a valve either too slow or too small to handle the problem. When the surge system is used for startup, a second valve may be required, as the “turn down” burden of a single valve may become too large. Frequently, surge control valves are selected too large, resulting not only in unnecessary cost but also in significant loss of control sensitivity, a loss that cannot be overcome through reduced signal sampling intervals, actuator changes, etc. Improper valve type selection can result in insufficient sensitivity and cause excessive noise, which usually limits valve durability.

| 265 |

| Centrifugal Compressors | When air is the compressed medium, the surge control valve normally can relieve to atmosphere, usually through an adequate silencer. If, however, the compressed medium is a process gas or a refrigerant, the valve will have to bypass the section being protected and return the flow to the proper suction nozzle. In such cases, some form of cooling is required to remove the heat of compression. A shell and tube exchanger or an air exchanger, as shown in Figure 7-32, may be required. The system shown in Figure 7-32(a), which is a single recycle loop system, did not work well. This was especially true for high molecular weight operation where the low-pressure (LP) casing was in surge while the high-pressure (HP) casing was in stonewall. The situation was impossible to correct by opening the recirculation valve. It was impossible to start the system without mismatching the LP and HP casings during speed changes. The LP and HP signals to the recirculation valve also interacted and created a problem. The configuration shown in Figure 7-32(b) also was unworkable, because of reverse train rotation after trip. This was due to the large amount of energy stored in the discharge cooler. The emergency shutdown valve (ESD) was not fast enough to cope with this problem. Also, when the HP spill-back came open first, the LP casing went into surge. In addition, there was difficulty in maintaining system balance during start-up and loading. The preferred arrangement is shown in Figure 7-32(c). If the gas is sent through a condenser (as in the case of a refrigerant), direct contact cooling may be possible. On smaller units, a liquid level can be retained in the bottom of the suction drum with normal flow into the side above liquid level. The surge control by-pass flow passes through the bottom, contacting the liquid and thereby de-superheating. As the systems get larger, this method is not feasible, due to the shaking forces from the boiling liquid. For such service, a sparger or a modified steam desuperheater can be used, in which the liquid is sparged (or sprayed) into the bypass line ahead of the suction drum. A ring of temperature sensors monitors the spray and is set for the desired amount of residual superheat. Cooling a system on bypass can make startup on process gas much easier. On multi-nozzle compressors or compressors with intercoolers, bypassing should be considered for individual sections. This will bypass the necessary flow to keep the affected stage out of surge, rather than waste energy by bypassing the entire compressor or train (if multibodied). The bypass scheme must be evaluated for both energy and stability, as there may be a tendency for some multiple loops to interact if proper care has not been exercised.

| 266 |

| Surge Control |

Figure 7-32: Unacceptable and acceptable equipment arrangement.

Figure 7-33 shows the location of a check valve in the system. Figure 7-33(b) is the required arrangement if parallel operation is anticipated. This system is also required if the compressor must start up against stagnation pressure, as the pressure on the downstream side would keep the check valve from opening. Field experience and experimental work have shown some advantage to the control shown in Figure 7-33(a), if the compressor is not in parallel operation and the downstream pressure would not prevent the check valve from opening. This configuration, however, is rather infrequently used in the petrochemical industry. The inherent stability of such a scheme is much greater and allows for more gain and responsiveness in the system, without the effort required for the scheme shown in Figure 7-33(b). Microprocessor based surge control systems can be made very sophisticated. Most of the variables external to the compressor are readily measured and can be digitized without problem, or can be treated and combined in analog form. The signals may be processed either in analog form or digitally, with the latter having an advantage. For multiple

| 267 |

| Centrifugal Compressors | component gas, measuring the molecular weights “K” and “Z” is still somewhat of a problem. Sometimes a density meter can be used if the variations are not too great. This compensates for apparent molecular weight, but not “K” or “Z.” There is no easy answer for this problem.

Figure 7-33: Location of the check valve.

RESPONSE TIME OF SURGE CONTROL SYSTEMS Often, surge control systems that are operating satisfactorily under normal conditions are inadequate when fast response time is required. A fast rise in head would be an example. Such a fast rise can occur if one compressor in a series combination shuts down because of malfunction or if the flow is decreased rapidly because of process malfunction. As can be seen from Figure 7-34, on shutdown of one compressor in a series combination, the remaining compressor would experience an increase in Delta P, due to a backpressure effect. Without provisions in the surge control system, it would start opening the surge valve when the operating

| 268 |

| Surge Control | point approached the surge line. Because of the response time of the system (actuators, valve travel, etc.), the compressor most likely would experience surge. This can be avoided by adding anticipatory action, either by adding one-way derivative action to the surge controller on a fast rise in head or by fast decrease in flow. Surge valve opening would occur immediately after detection of such fast changes, reducing the response time and thus preventing surge.

Figure 7-34: Series operation of compressors.

STABILITY OF SURGE CONTROL SYSTEMS AND PARALLEL COMPRESSORS CLOSE TO SURGE Parallel operation of compressors close to the surge line can lead to unstable operation of the surge systems if suitable control precautions are not taken. Only in rare cases are the surge lines of parallel compressors the same. Manufacturing tolerances and piping differences result in different surge lines, as shown in Figure 7-35. Two types of control design are possible. In the first type of design, the speed signal or other control

| 269 |

| Centrifugal Compressors | signal is biased so that both operating points have an equal distance to the surge line. In the second design, the speed or control signal to both units is the same, which would result in a different location of the operating point in the compressor envelope. The surge control system is equipped with a feature to detect the distance from the operating point to the surge line. Any speed decrease of the unit closest to surge will be interrupted at a preset distance from the surge line. This preset distance is selected just to the right of the point where surge valve opening would be required. Any additional reduction of throughput will be achieved with the other unit until it also reaches this preset distance. This feature allows maximum operation without opening the recycle valves, which considerably decreases the plant efficiency. The first design is more complex than the second; the preferred design depends on the location of the efficiency lines. A comparison of combined efficiency of the parallel compressors will most likely determine the type of design preferred.

Figure 7-35: Parallel operation of compressors.

| 270 |

| Surge Control | Surge system stability of parallel or series compressors is important, since oscillation of the surge control systems, because of their time constants, can lead to excursions into deep surge, which are as damaging as the surge that is to be prevented. To prevent oscillation of compressors, their time constants should be as far apart as possible. One of the simplest ways of achieving this is by affecting the travel time of the recycle (surge) valve. Since, obviously, the opening time of the valve should be as short as possible, we are left with the adjustment of the closing time. This adjustment has proven effective in preventing oscillation in most applications. Due consideration also should be given to integral and proportional action of the surge control systems. However, it is much easier to find a stable condition empirically with valve closing adjustment than to use other methods, which often require dynamic modeling. D Y N A M I C S I M U L AT I O N OF COMPRESSOR SYSTEMS Dynamic simulation is an engineering tool used to evaluate a process design. It involves the use of a mathematical model describing the behavior of the system in an equation form that is then solved using a computer. The equations consist of algebraic and differential relationships, whose origins come from laws of physics and from empirical data. The computer representation then can be used to predict how the actual system will behave in “what if” situations. The results describe both the steady state operation and the transient behavior during upsets. Because this analysis can be performed during the design phase, the knowledge gained can be used to reduce overall costs of design, construction, startup and operation of the system of interest. The development of a model for a specific application involves many people, e.g., process engineers, design engineers, clients, simulation analysts, and vendors. The exchange of information between these groups also serves a quality assurance function for a project. With higher and higher trends in the capacity, pressure, speed, and horsepower of centrifugal compressors, the design process becomes more involved. The compressor designers and the application engineers, who are responsible for the design of the entire process, are faced with the problem of designing machines which are often larger and more complex than any ever built. They must somehow provide reasonable assurances that the machine will operate as intended, once installed. This process of satisfying one’s self (and the plant’s owner) that the compressor systems

| 271 |

| Centrifugal Compressors | design will perform satisfactorily can be very subjective. Traditional methods of analysis cannot provide accurate predictions of how a compressor and a process will interact. Because of this, deigns are often based upon the designer’s experiences and opinions. A tool that has been extremely valuable in evaluating the performance of a compressor system prior to installation is dynamic simulation. The simulation studies result in design recommendations concerning the number and location of recycles required, sizing of recycle control valves, set point, gain, and reset settings for control system instrumentation. The studies also may lead to a better understanding of how the compressors will interact with processes. Startup procedures are tested and amended as required The question of utilizing single vs. multiple recycle loops for multicasing applications is difficult to answer conclusively based on intuition, as past experience may not be a sufficient guide. The dynamic behavior of such systems is influenced by process parameters, control system responses, and the turbine-compressor performance curves. The impracticality of manual calculations and intuitive design approaches creates a need for computer-aided design. C O M P R E S S O R S Y S T E M S I M U L AT I O N Figure 7-36 illustrates the procedural stops of a typical simulation study. The steps can be grouped into five phases. In Phase I, the boundaries of interest are established, operating modes and limit identified, criteria of acceptable performance documented, and the types of upsets of interest identified. The mathematical representation is prepared and assembled, and pertinent physical data are gathered. In Phase 2, the mathematical model is converted into computer language, and the program is checked out. A simulated steady state operating condition is achieved and compared with known/expected operating conditions. In Phase 3, the simulation model is evaluated at transient conditions. The model behavior is analyzed and validated as representing the system performance. This validation involves comparison of past knowledge from similar systems, review of behavior by specialists, comparison with expected behavior of physical laws, and a review by personnel with many years of dynamic simulation experience. In Phase 4, the simulation model is used to analyze “what if” scenarios. A series of test runs is performed to document how the actual system is expected to behave, and the results are analyzed. In Phase 5, conclusions and

| 272 |

| Surge Control | recommendations regarding system performance are documented and submitted in a final report. The data from the test runs are assembled, edited and presented. The final report also documents important aspects of the simulation activity.

Figure 7-36: Simulation study diagram.

Figure 7-37 illustrates how the mathematical model for a typical compressor study is organized. The model for a compressor system can be considered to consist of three separate modules, among which information is passed: 1. Process Module. The process module consists of algebraic and differential equations describing the accumulation of mass and energy at various places in the system such as vessels or piping junctions (nodes), the transfer of mass and energy between the nodes, and thermodynamic characteristics such as enthalpies or phase equilibria. The process model can be relatively simple, such as the model for an injection system in which variations in temperature or molecular weight are not significant and, therefore, only mass balances need be

| 273 |

| Centrifugal Compressors | considered; or the process model may be highly complex, such as refrigeration system models in which heat transfer, phase equilibria, and temperature variations have to be accounted for, in addition to mass balances. 2. Compressor Module. The compressor module consists of the empirical head/flow characteristics of each section as provided by the manufacturer, thermodynamic compression equations, and a torque balance between the compressor and its driver. The compressor module receives suction and discharge pressure and temperature from the process module. It then calculates compressor flow and discharge temperature as outputs to the process module. In the calculation process, compressor head, speed, and volumetric flow are determined. The compressor model may be a simple representation of a constant speed machine, or a very complex one in which variations in speed, efficiency, and polytropic coefficients are considered. The degree of sophistication required is a function of both the type of system under study and the objectives of the study. For example, a model to study start-up would have to be more complex than a model used to study constant speed operation. 3. Control System Module. The control system module represents the behavior of the process control instrumentation and final control elements. Controllers, transmitters, computing relays, and control valves usually are modeled mathematically. A tie-in to actual control hardware, e.g., distributed control systems, may be required to evaluate the response of software algorithms. Control valve dynamics (essentially the valve stroking period) are modeled, as are significant measurement lags.

Example of a Simulated Compression system Figure 7-38 is a simulation flow diagram for one train of a parallel train configuration that was used to handle gas from an oil field separation process. The incoming gas/water/oil flow from the field at low-pressure conditions was separated, and the gas was routed to the suction header of the compression system. The gas was compressed for distribution to a central compression plant, which further increased the pressure for injection back into the oil reservoir. The major objectives of the simulation study were to evaluate alternate process configurations during both trip and normal process upsets,

| 274 |

| Surge Control |

Figure 7-37: Compression model system.

Figure 7-38: Compressor flow diagram.

| 275 |

| Centrifugal Compressors | to evaluate the effects of valve failures, and to determine the sensitivity of valve parameters to the prevention of surge. The methodology described in the previous section was followed. The model was subjected to a series of test runs. Of particular importance were the “Emergency Trip” and the blocked compressor discharge upset; such as the sudden failure and the instantaneous closure of block valve 5 in Figure 7-38. The simulation model results were instrumental in the decision to add a second check valve between the compressor discharge and the discharge cooler, and the addition of a hot bypass loop around the compressor. Also, control system parameters, such as valve parameters and controller parameters, were needed for the prevention of surge to be determined. In a follow-up study, the performance of parallel trains during normal operation, upset conditions, and start-up was analyzed. The first simulation investigated the behavior of the gas turbine compressor train during a trip of the gas turbine. Figures 7-39 and 7-40 illustrate typical results for the revised design under “Emergency Trip” conditions. Figure 7-39 shows the time history of important compressor variables, such as the effect on the compressor suction pressure, compressor discharge pressure, compressor load, and the compressor speed. Figure 7-40 shows that the compressor did not surge at high energy levels. In Figure 7-40, the surge parameter variable reflects the flow that would indicate the boundary of surge based on the surge line. The actual flow is also depicted and reflects the actual flow through the machine. If the actual flow is greater than the surge flow, no surging occurs. Without the process changes identified above, this condition could not be met. Once a machine enters surge, the simulation model does not describe the resulting phenomena. The characteristic changes in the centrifugal compressor due to an abrupt closure of the block valve (due to valve failure) are shown in Figures 7-41, 7-42, and 7-43. As the block valve closes suddenly, the compressor goes into the recycle mode and is operated on the set control line set by the Flow/Delta-P controller.

Control Hardware Evaluation Once the parameters for the antisurge recycle loop, e.g., valve capacity, characteristics, response time, and controller algorithm and settings, have been determined, the next question centers on the type of control hardware to be used. A decision has to be made whether the control algorithm is implemented with a continuous system, e.g., electronic analog, or with a sampled-data system, e.g., distributed digital control.

| 276 |

| Surge Control |

Figure 7-39: Compressor characteristics as the gas turbine trips at compressor full load.

Figure 7-40: Compressor characteristics as the gas turbine trips at compressor full load.

| 277 |

| Centrifugal Compressors | The simulation model run in real time provides “field like” transmitter outputs for the flow and delta-pressure measurements and acceptable. “Field like” transducer inputs from the actual control hardware. The interconnections between the input/output (I/O) of the computer and the control hardware were the same as field signals (4-20 ma, 1-5 volts). When the train discharge block valve (Valve 5 in Figure 7-38) closed, the discharge pressure increased, and the compressor moved toward the surge region. The response of the antisurge recycle loop put the compressor in recycle flow at a flow set by the control line.

Figure 7-41: Block valve failure at compressor steady state flow at full capacity.

| 278 |

| Surge Control |

Figure 7-42: Block valve failure at steady state compressor flow at full capacity.

Figure 7-43: Block valve failure at steady state flow at full capacity.

| 279 |

| Centrifugal Compressors | REFERENCES Boyce, M.P., Bohannan, W.R., Brown, R.N., Gaston, J.R., Meher-Homji, C.B., Meier, R.H., and Pobanz, N.E. “Practical Aspects of Centrifugal Compressor Surge and Surge Control Systems.” Proc. of the 12th TurboMachinery Symposium. Houston, TX . 15-17 Nov. 1983. 147-73. Boyce, M. P. “A Practical Three-Dimensional Flow Visualization Approach to the Complex Flow Characteristics in a Centrifugal Impeller,” ASME Paper No. 66-GT- 83. Brown, R. N. “An Experimental Investigation of a Pneumatic Close Loop Anti-surge Control for Centrifugal and Axial Flow Compressors.” Masters Thesis–University of Wisconsin, Madison, Wisconsin (1966). Bullock, R. O., Wilcox, W. W., and Moses, J. J. “Experimental and Theoretical Studies on Surging in Continuous Flow Compressors.” NACA Report No. 861 (1946). Chiu, K. C. and Pobanz, N. E. “Applicability of Distributed Control to a Fast Response Loop.” Proceedings of the ISA ’83 International Conference (1983). Dean, R. C. and Young, L. R. “The Time Domain of Centrifugal Compressor and Pump Stability and Surge.” 1976 Joint Gas Turbine & Fluids Engineering Divisions Conference, ASME (March 22-25, 1976). Eby, R. S. and Pobanz, N. E. “Dynamic Simulation and Verification of a Compression-Liquefaction System for Material Withdrawn From a Uranium Enrichment Plant.” Proceedings for the 1983 Summer Computer Simulation Conference (1983). Emmons, H. W., Pearson, C. E., and Grant, II. R. “Compressor Surge and Stall Propagation. “Trans. ASME (May 1955). Fulleman, J. “Centrifugal Compressors.” Chapter 8 in Advances in Petroleum Chemistry and Refining. Interscience Publishers (division of John Wiley & Sons), New York-London (1962). Gaston, J. R. “Antisurge Control Schemes for Turbocompressors.” Chemical Engineering (April 19, 1982). Hallock, D. C. “Centrifugal Compressors…the Cause of the Curve.” Air and Gas Engineering. 1, 1 (1968). Japikse, D. “Stall, Stage Stall, and Surge.” Proceedings of l0th Annual Turbomachinery Symposium, Texas. A&M University (December 1981) Stanley, R.A. and Bohannan, W. R. “Dynamic Simulation of Centrifugal Compressor Systems.” Proceedings of the 6th Turbomachinery Symposium, Texas A&M University (1977). White, M. H. “Surge Control for Centrifugal Compressors.” Chemical Engineering (December 25, 1972).

| 280 |

8

Gas Turbines

The gas turbine is one of the main prime movers of a compressor train. For remote locations, gas turbines are mostly used, due to their low maintenance and the ability to prepackage the units. Their light weight makes them a must for offshore platforms. Gas turbines also have an advantage because of their multi-fuel capabilities. In many cases, this capability assures that the unit must be a gas turbine. It has been used widely as the drive for both on-shore and offshore compressor trains. The gas turbine is a power plant, which produces a great amount of energy for its size and weight. The gas turbine has found increasing service in the past 15 years in the power and petrochemical industry throughout the world. Its compactness, low weight, and multiple fuel application make it a natural power plant in all applications from power plants to offshore platforms. Today there are gas turbines, which run on natural gas, diesel fuel, naphtha, methane, crude, low-Btu gases, vaporized fuel oils, and even waste. The last twenty years has seen a large growth in gas turbine technology. The growth is spearheaded by the growth of materials technology, new coatings, and new cooling schemes. This, with the conjunction of increase in compressor pressure ratio, has increased the gas turbine thermal efficiency from about 15% to more than 45%. The advanced gas turbines are operating at very high-pressure ratios, and very high firing temperatures, ensuring high performance of power and efficiency. These turbines are pushing the envelope of technology in

| 281 |

| Centrifugal Compressors | the areas of material science and aerodynamics to their limit. The new gas turbines have high efficiencies in the low 40s as compared to the older turbine with efficiencies in the high 20s. Since fuel costs are nearly 75% of the life cycle cost of a plant, these new advanced gas turbines are here to stay and will be in large demand. The aerospace engines have been the leaders in most of the technology in the gas turbine and have been used on offshore applications to drive centrifugal compressors and pumps. The design criteria for these engines was high reliability, high performance, with many starts and flexible operation throughout the flight envelope. The engine life of about 3500 hours between major overhauls was considered good. The aerospace engine performance has always been rated primarily on its thrust / weight ratio. Increase in engine thrust / weight ratio is achieved by the development of high aspect ratio blades in the compressor as well as optimizing the pressure ratio and firing temperature of the turbine for maximum work output per unit flow. This gives the aeroderivative engine a small footprint making it attractive for offshore applications The industrial gas turbine has always emphasized long life, and this conservative approach has resulted in the industrial gas turbine in many aspects giving up high performance for rugged operation. The industrial gas turbine has been conservative in the pressure ratio and the firing temperatures. This has all changed in the last ten years, spurred on by the introduction of the “Aeroderivative Gas Turbine.” The industrial gas turbine has dramatically improved its performance in all operational aspects. This has resulted in dramatically reducing the performance gap between these two types of gas turbines. Figures 8-1 and 8-2 show the growth of the pressure ratio and firing temperature. The growth of both the pressure ratio and firing temperature parallel each other, as both growths are necessary to achieving the optimum thermal efficiency. Gas turbines in the power industry can be classified into five broad groups:

| 282 |

•

Industrial heavy-duty gas turbines

•

Aircraft-derivative gas turbines

•

Medium-range gas turbines

•

Small gas turbines

•

Microturbines

| Gas Turbines |

Figure 8-1: Development of engine pressure ratio over the years.

TREND IN IMPROVEMENT IN FIRING TEMPERATURE

Figure 8-2: Trend in improvement in firing temperature.

| 283 |

| Centrifugal Compressors | In the past, the gas turbine was perceived as a relatively inefficient power source when compared to other power sources. Its efficiencies were as low as 15% in the early 1950’s, today’s its efficiencies are in the 45%50% range. The limiting factor for most gas turbines has been the turbine inlet temperature. With new schemes of air-cooling and breakthroughs in blade metallurgy, higher turbine temperatures have been achieved. The new gas turbines have fired inlet temperatures as high as 2300ºF (1260ºC), and pressure ratios of 40:1 with efficiencies of 45% and above. Some factors one must consider in deciding what type of power plant is best suited for the needs at hand are capital cost, time from planning to completion, maintenance costs, and fuel costs. The gas turbine has the lowest maintenance and capital cost. It also has the fastest completion time to full operation of any other plant. Its disadvantage was its high heat rate, but this has been addressed. The new turbines are among the most efficient types of prime movers. The design of any gas turbine must meet essential criteria based on operational considerations. Chief among these criteria are: •

High Efficiency

•

High Reliability and, thus, High Availability

•

Ease of Service

•

Ease of Installation and Commission

•

Conformance with Environmental Standards

•

Incorporation of Auxiliary and Control Systems which have a high degree of reliability

•

Flexibility to meet various service and fuel needs

A look at each of these criteria will enable the user to get a better understanding of the requirements. The two factors that most affect high turbine efficiencies are pressure ratios and temperature. The axial flow compressor, which produces the high-pressure gas in the turbine, has seen dramatic change as the gas turbine pressure ratio has increased from 7:1 to 40:1. The increase in pressure ratio increases the gas turbine thermal efficiency when accompanied with the increase in turbine firing temperature. Figure 8-3 shows the effect on the overall cycle efficiency of the increasing Pressure Ratio and the Firing Temperature. The increase in the pressure ratio

| 284 |

| Gas Turbines | increases the overall efficiency at a given temperature, however, increasing the pressure ratio beyond a certain value at any given firing temperature can actually result in lowering the overall cycle efficiency. It should also be noted that the very high-pressure ratios tend to reduce the operating range of the turbine compressor. This causes the turbine compressor to be much more intolerant to dirt build up in the inlet air filter and on the compressor blades and creates large drops in cycle efficiency and performance. In some cases it can lead to compressor surge, which in turn can lead to a flameout, or even serious damage and failure of the compressor blades and the radial and thrust bearings of the gas turbine. Overall Cycle Efficiency Tamb = 69°F (15°C), Eff. Comp. = 87%, Eff. Turb. = 92%

Figure 8-3: Overall cycle efficiency.

The effect of temperature is very predominant—for every 100°F (55.5ºC) increase in temperature, the work output increases approximately 10% and gives about a 11⁄2% increase in efficiency. Higher-pressure ratios and turbine inlet temperatures improve efficiencies on the simple cycle gas turbine. Figure 8-4 shows a simple cycle gas turbine performance map as a function of pressure ratio and turbine inlet temperature.

| 285 |

| Centrifugal Compressors |

Efficiency, %

The Performance Map of a Simple Cycle Gas Turbine

Figure 8-4: Performance map of a simple cycle gas turbine.

The Performance Map of a Regenerative Gas Turbine

Figure 8-5: Performance map of a regenerative gas turbine.

| 286 |

| Gas Turbines | Another way to achieve higher efficiencies is with regenerators. Figure 8-5 shows the effects of pressure ratio and temperatures on efficiencies and work for a regenerative cycle. The effect of pressure ratio for this cycle is opposite to that experienced in the simple cycle. Regenerators can increase efficiency as much as 15-20% at today’s operating temperatures. The optimum pressure ratios are about 20:1 for a regenerative system compared to 40:1 for the simple cycle at today’s higher turbine inlet temperatures that approach 3000ºF (1649ºC). The regenerative turbine is used in many pipeline compressor applications. High availability and reliability are the most important parameters in the design of a gas turbine. The availability of a power plant is the percent of time the plant is available to generate power in any given period. The reliability of the plant is the percentage of time between planed overhauls. Serviceability is an important part of any design, since fast turnarounds result in high availability to a turbine and reduces maintenance and operations costs. Service can be accomplished by providing proper checks such as exhaust temperature monitoring, shaft vibration monitoring, and surge monitoring. Also, the designer should incorporate borescope ports for fast visual checks of hot parts in the system. Split casings for fast disassembly, field balancing ports for easy access to the balance planes, and combustor cans, which can be easily disassembled without removing the entire hot section, are some of the many ways that afford ease of service. Ease of installation and commissioning is another reason for gas turbine use. A gas turbine unit can be tested and packaged at the factory. Use of a unit should be carefully planned so as to cause as few start cycles as possible. Frequent start-ups and shutdowns at commissioning greatly reduce the life of a unit. Environmental considerations are critical in the design of any system. The system’s impact on the environment must be within legal limits and, thus, must be addressed by the designer carefully. Combustors are the most critical component, and great care must be taken to design them to provide low smoke and low NOx output. The high temperatures result in increasing the NOx emissions from the gas turbines. This resulted in initially attacking the NOx problem by injecting water or steam in the combustor. The next stage was the development of Dry Low NOx Combustors. The development of new Dry Low NOx Combustors has been a very critical component in reducing the NOx output as the gas turbine firing temperature is increased. The new low NOx combustors increase the number of fuel nozzles and the complexity of the control algorithms.

| 287 |

| Centrifugal Compressors | Lowering the inlet velocities and providing proper inlet silencers can reduce air noise. Considerable work by NASA on compressor casings has greatly reduced noise. Auxiliary systems and control systems must be designed carefully, since they are often responsible for the downtime in many units. Lubrication systems, one of the critical auxiliary systems, must be designed with a backup system and should be as close to failure-proof as possible. The advanced gas turbines are all digitally controlled and incorporate on-line condition monitoring to some extent. The addition of new on-line monitoring requires new instrumentation. Control systems provide acceleration-time and temperature-time controls for startups as well as control various anti-surge valves. At operating speeds, they must regulate fuel supply and monitor vibrations, temperatures, and pressures throughout the entire range. Flexibility of service and fuels are criteria that enhance a turbine system, but they are not necessary for every application. The energy shortage makes closer to its operating point and thus operate at higher efficiencies. This flexibility may entail a two-shaft design incorporating a power turbine, which is separate and not connected to the gasifier unit. Multiple fuel applications are now in greater demand, especially where various fuels may be in shortage at different times of the year. The previous criteria are some of the many that designers must meet to design successful units.

INDUSTRIAL HEAVY-DUTY GAS TURBINES These gas turbines were designed shortly after World War II and introduced to the market in the early 1950s. The early heavy-duty gas turbine design was largely an extension of steam turbine design. Restrictions of weight and space were not important factors for these ground-based units, and so the design characteristics included heavy-wall casings split on horizontal centerlines, sleeve bearings, large-diameter combustors, thick airfoil sections for blades and stators, and large frontal areas. The overall pressure ratio of these units varied from 5:1 for the earlier units to 35:1 for the units in present-day service. Turbine inlet temperatures have been increased and run as high as 2500°F (1371ºC) on some of these units. This makes the gas turbine one of the most efficient prime movers on the market today, reaching efficiencies of 50%. Projected

| 288 |

| Gas Turbines | temperatures approach 3000ºF (1649ºC) and, if achieved, would make the gas turbine an even more efficient unit. The Advanced Gas Turbine Programs sponsored by the U.S. Department of Energy has these high temperatures as one of its goals. To achieve these high temperatures, steam cooling is being used in the latest designs to achieve the goals of maintaining blade metal temperatures below 1300ºF (704ºC) and to prevent hot corrosion problems. The industrial heavy-duty gas turbines employ axial-flow compressors and turbines. The industrial turbine consists of a 15-17 stage axial flow compressor with multiple can-annular combustors, each connected to the other by crossover tubes. The crossover tubes help propagate the flames from one combustor can to all the other chambers and also assures an equalization of the pressure between each combustor chamber. The earlier industrial European designs have single stage side combustors. The new European designs do not use the side combustor in most of their newer designs. The newer European designs have can-annular or annular combustors since side (silo type) combustors had a tendency to distort the casing. Figure 8-6 is a cross-sectional representation of the GE Industrial Type Gas Turbine, with can-annular combustors, and Figure 8-7 is a cross-sectional representation of the Siemens Silo Type Combustor Gas Turbine. The turbine expander consists of a 2-4 stage axial flow turbine, which drives both the axial flow compressor and the generator.

Figure 8-6: A frame type gas turbine with can-annular combustors (courtesy GE Turbines).

| 289 |

| Centrifugal Compressors |

Figure 8-7: Frame type gas turbine with silo type combustors (courtesy Siemens).

The large frontal areas of these units reduce the inlet velocities, thus reducing air noise. The pressure rise in each compressor stage is reduced, creating a large, stable operating zone. The auxiliary modules used on most of these units have gone through considerable hours of testing and are heavy-duty pumps and motors. The advantages of the heavy-duty gas turbines are their long life, high availability, and slightly higher overall efficiencies. The noise level from this type of turbine is considerably less than an aircraft-type turbine. The heavy-duty gas turbine’s largest customers are the electrical utilities and independent power producers. Since the 1990s, the industrial turbines have been the bulwarks of most combined cycle power plants.

AERODERIVATIVE GAS TURBINES The Aeroderivative gas turbines consist of two basic components: an aircraft-derivative gas generator and a free-power turbine. The gas generator serves as a producer of gas energy or gas horsepower. The gas generator is derived from an aircraft engine modified to burn industrial fuels. Design innovations are usually incorporated to ensure the required long-life characteristics in the ground-based environment. In case of fan jet designs, the fan is removed, and a couple of stages of compression are added in front of the existing low-pressure compressor. The axial flow compressor in many cases is divided into two sections: a low-pressure compressor followed by a high-pressure compressor. In those cases, there

| 290 |

| Gas Turbines | are usually a high-pressure turbine and a low-pressure turbine, which drives the corresponding sections of the compressor. The shafts are usually concentric, thus, the speeds of the high pressure and low-pressure sections can be optimized. In this case the power turbine is separate and is not mechanically coupled. The only connection is via an aerodynamic coupling. In these cases the turbines have three shafts, all operating at independent speeds. The gas generator serves to raise combustion gas products to conditions of around 45-75 psi (3-5 Bar) and temperatures of 1300ºF-1700ºF (704ºC-927ºC) at the exhaust flange. Figure 8-8 shows a cross section of an aeroderivative engine.

Figure 8-8: A cross-section of an aeroderivative gas turbine engine.

Both the power industry and the petrochemical industries use the aircraft-type turbine. The power industry uses these units in a combined cycle mode for power generation, especially in remote areas where the power requirements are less than 100MW. The petrochemical industry uses these types of turbines on offshore platforms especially for gas reinjection and as power plants for these offshore platforms, mostly due to their compactness and the ability to be easily replaced and then sent out to be repaired. The aeroderivative gas turbine also is used widely by gas transmission companies and petrochemical plants, especially for many variable speed mechanical drives. The benefits of the aeroderivative gas turbines are: Favorable installation cost. The equipment involved is of a size and weight that it can be packaged and tested as a complete unit within the

| 291 |

| Centrifugal Compressors | manufacturer’s plant. Generally, the package will include either a generator or a driven pipeline compressor and all auxiliaries and control panels specified by the user. Immediate installation at the job site is facilitated by factory matching and debugging. Adaptation to remote control. Users strive to reduce operating costs by automation of their systems. Many new offshore and pipeline applications today are designed for remote unattended operation of the compression equipment. Jet gas turbine equipment lends itself to automatic control, as auxiliary systems are not complex, water cooling is not required (cooling by oil-to-air exchanges), and the starting device (gas expansion motor) requires little energy and is reliable. Safety devices and instrumentation adapt readily for purposes of remote control and monitoring the performance of the equipment. Maintenance concept. The off-site maintenance plan fits in well with these systems where minimum operating personnel and unattended stations are the objectives. Technicians conduct minor running adjustments and perform instrument calibrations. Otherwise, the aeroderivative gas turbine runs without inspection until monitoring equipment indicates distress or sudden performance change. This plan calls for the removal of the gasifier section (the aero-engine) and sending it back to the factory for repair while another unit is installed. The power turbine does not usually have problems since its inlet temperature is much lower. Downtime due to the removal and replacement of the gasifier turbine is about eight hours.

MEDIUM-RANGE GAS TURBINES Medium-range gas turbines are usually rated between 5 and 15 MW. These units are similar in design to the large heavy-duty gas turbines; their casing is thicker than the aeroderivative casing but thinner than the industrial gas turbines. They usually are split-shaft designs that are efficient in part load operations. Efficiency is achieved by letting the gasifier section (the section which produces the hot gas) operate at maximum efficiency while the power turbine operates over a great range of speeds. The compressor is usually a 10-16 stage subsonic axial compressor, which produces a pressure ratio from about 5:1-15:1. Most American designs use can-annular (about 5–10 combustor cans mounted in a circular ring) or annular-type combustors. Most European designs use side combustors and have lower turbine inlet temperatures compared to their American counterparts. Figure 8-9 shows a medium-range gas turbine.

| 292 |

| Gas Turbines |

Figure 8-9: A medium-sized industrial gas turbine (courtesy Solar Turbine Corp).

The gasifier turbine is usually a 2-3 stage axial turbine with an aircooled first-stage nozzle and blade. The power turbine is usually a singleor two-stage axial-flow turbine. The medium-range turbines are used on offshore platforms and are finding increasing use in petrochemical plants. The straight simple cycle turbine is low in efficiency, but, by using regenerators to consume exhaust gases, these efficiencies can be greatly improved. In process plants, this exhaust gas is used to produce steam. The combined-cycle (air-steam) cogeneration plant has very high efficiencies and is the trend of the future. These gas turbines have, in many cases, regenerators or recuperators to enhance the efficiency of the turbines. Figure 8-10 shows such a new recuperated gas turbine design, which has an efficiency of 38%. The term “regenerative heat exchanger” is used for this system in which the heat transfer between two streams is affected by the exposure of a third medium alternately to the two flows. (The heat flows successively into and out of the third medium, which undergoes a cyclic temperature.) In a recuperative heat exchanger, each element of head-

| 293 |

| Centrifugal Compressors | transferring surface has a constant temperature and, by arranging the gas paths in contraflow, the temperature distribution in the matrix in the direction of flow gives optimum performance for the given heat-transfer conditions. This optimum temperature distribution can be achieved ideally in a contraflow regenerator and approached very closely in a crossflow regenerator.

Recuperator

Compressor

Turbine

Combustor

Figure 8-10: A recuperative medium-sized industrial gas turbine (courtesy Solar Turbines).

SMALL GAS TURBINES Many small gas turbines, which produce below 5 MW, are designed similar to the larger turbines already discussed; however, there are many designs which incorporate centrifugal compressors or combinations of centrifugal and axial compressors, as well as radial-inflow turbines. A

| 294 |

| Gas Turbines | small turbine will often consist of a single-stage centrifugal compressor producing a pressure ratio as high as 6:1, a single side combustor where temperatures of about 1,800ºF (982ºC) are reached, and radial-inflow turbines. Figure 8-11 shows a schematic of such a typical turbine. Air is induced through an inlet duct to the centrifugal compressor, which, rotating at high speed, imparts energy to the air. On leaving, the impeller air, with increased pressure and velocity, passes through a high-efficiency diffuser, which converts the velocity energy to static pressure. The compressed air, contained in a pressure casing, flows at low speed to the combustion chamber, which is a side combustor. A portion of the air enters the combustor head, mixes with the fuel, and burns continuously. The remainder of the air enters through the wall of the combustor and mixes with the hot gases. Good fuel atomization and controlled mixing ensure an even temperature distribution in the hot gases, which pass through the volute to enter the radial inflow turbine nozzles. High acceleration and expansion of the gases through the nozzle guide vane

Figure 8-11: A small radial flow gas turbine cutaway showing the turbine rotor.

passages and turbine combine to impart rotational energy, which is used to drive the external load and auxiliaries on the cool side of the turbine. The efficiency of a small turbine is usually much lower than a larger unit because of the limitation of the turbine inlet temperature and the lower component efficiencies. Turbine inlet temperature is limited because the turbine blades are not cooled. Radial-flow compressors and impellers

| 295 |

| Centrifugal Compressors | inherently have lower efficiencies than their axial counterparts. These units are rugged, and their simplicity in design assures many hours of trouble-free operation. A way to improve the lower overall cycle efficiencies, 18%-23%, is to use the waste heat from the turbine unit. High thermal efficiencies (30-35%) can be obtained, since nearly all the heat not converted into mechanical energy is available in the exhaust, and most of this energy can be converted into useful work. These units, when placed in a combined heat power application, can reach efficiencies of the total process as high as 60% -70%. Figure 8-12 shows an aeroderivative small gas turbine. This unit has three independent rotating assemblies mounted on three concentric shafts. This turbine has a three-stage axial flow compressor followed by a centrifugal compressor, each driven by a single stage axial flow compressor. Power is extracted by a two-stage axial flow turbine and delivered to the inlet end of the machine by one of the concentric shafts. The combustion system comprises a reverse flow annular combustion chamber with multiple fuel nozzles and a spark igniter. This aeroderivative engine produces 4.9 MW and has an efficiency of 32%.

Axial Flow Compressor

Concentric Shafts

Centrifugal Compressor

Axial Flow Turbine

Figure 8-12: A small aeroderivative gas turbine (courtesy Pratt & Whitney Canada Corporation).

| 296 |

| Gas Turbines | MAJOR GAS TURBINE COMPONENTS COMPRESSORS In gas turbines, the centrifugal flow and axial flow compressors, which are continuous flow compressors, are the ones used for compressing the air. Positive displacement compressors such as the gear type units are used for lubrication systems in the gas turbines. AXIAL-FLOW COMPRESSORS An axial-flow compressor compresses its working fluid by first accelerating the fluid and then diffusing it to obtain a pressure increase. The fluid is accelerated by a row of rotating airfoils or blades (the rotor) and diffused by a row of stationary blades (the stator). The diffusion in the stator converts the velocity increase gained in the rotor to a pressure increase. One rotor and one stator make up a stage in a compressor. A compressor usually consists of multiple stages. One additional row of fixed blades (inlet guide vanes) is frequently used at the compressor inlet to ensure that air enters the first-stage rotors at the desired angle. In addition to the stators, additional diffuser at the exit of the compressor further diffuses the fluid and controls its velocity when entering the combustors. In an axial compressor, air passes from one stage to the next with each stage raising the pressure slightly. By producing low-pressure increases on the order of 1.1:1–1.4:1, very high efficiencies can be obtained. The use of multiple stages permits overall pressure increases up to 40:1. The rule of thumb for a multiple stage gas turbine compressor would be that the energy rise per stage would be constant rather than the pressure rise per stage. Figure 8-13 shows a multistage high-pressure axial flow turbine rotor. The turbine rotor depicted in this figure has a low-pressure compressor followed by a high-pressure compressor. There are also two turbine sections. The reason there is a large space between the two turbine sections is that this is a reheat turbine, and the second set of combustors are located between the high pressure and the low pressure turbine sections. The compressor produces 30:1 pressure in 22 stages. The lowpressure increase per stage also simplifies calculations in the design of the compressor by justifying the air as incompressible in its flow through an individual stage.

| 297 |

| Centrifugal Compressors | LP Axial Flow Turbine HP Axial Flow Turbine

HP Axial Flow Compressor LP Axial Flow Compressor Figure 8-13: A high-pressure ratio turbine rotor (courtesy ALSTOM).

Figure 8-14: Variation of flow and thermodynamic properties in an axial flow compressor.

| 298 |

| Gas Turbines | Figure 8-14 shows the pressure, velocity, and total enthalpy variation for flow through several stages of an axial compressor. It is important to note here that the changes in the total conditions for pressure, temperature, and enthalpy occur only in the rotating component where energy is inputted into the system. As seen in Figure 8-13, the length of the blades and the annulus area, which is the area between the shaft and shroud, decreases through the length of the compressor. This reduction in flow area compensates for the increase in fluid density as it is compressed, permitting a constant axial velocity. As stated earlier, an axial-flow compressor operates on the principle of putting work into the incoming air by acceleration and diffusion. R E G E N E R AT O R S Heavy-duty regenerators are designed for applications in large gas turbines in the 5,000-100,000-kW range. The use of regenerators in conjunction with industrial gas turbines substantially increases cycle efficiency and provides an impetus to energy management by reducing fuel consumption up to 30%. Figure 8-15 shows how a regenerator works. In most present-day regenerative gas turbines, ambient air enters the inlet filter and is

Figure 8-15: A typical plate and fin-type regenerator for an industrial gas turbine.

| 299 |

| Centrifugal Compressors | compressed to a pressure ratio between 7 and 15:1, and a temperature of exiting the compressor between 530ºF (277ºC) and 800ºF (427ºC). The air is then piped to the regenerator which heats the air to about 950ºF (510ºC). The heated air then enters the combustor where it is further heated before entering the turbine. After the gas has undergone expansion in the turbine, it is about 1000ºF (537ºC)–1100ºF (590ºC) and essentially at ambient pressure. The gas is ducted through the regenerator where the waste heat is transferred to the incoming air. The gas is then discharged into the ambient air through the exhaust stack. In effect, the heat that would otherwise be lost is transferred to the air, decreasing the amount of fuel that must be consumed to operate the turbine. For a 30,000-kW turbine, the regenerator heats 10 million pounds of air per day. COMBUSTORS All gas turbine combustors perform the same function: they increase the temperature of the high-pressure gas. The gas turbine combustor uses very little of its air (10%) in the combustion process. The rest of the air is used for cooling and mixing. New combustors are also circulating steam for cooling purpose. The air from the compressor must be diffused before it enters the combustor. The velocity of the air leaving the compressor is about 400-600 ft/sec (122–183 m/sec) and the velocity in the combustor must be maintained below 50 ft/sec (15.2 m/sec). Even at these low velocities, care must be taken to avoid the flame to be carried on downstream. However, there are different methods to arrange combustors on a gas turbine. Designs fall into these categories: •

Tubular (side combustors)

•

Can-annular

•

Annular

The combustor is a direct-fired air heater in which fuel is burned almost stoichiometrically with one-third or less of the compressor discharge air. Combustion products are then mixed with the remaining air to arrive at a suitable turbine inlet temperature. Despite the many design differences in combustors, all gas turbine combustion chambers have three features as seen in Figure 8-16:

| 300 |

| Gas Turbines | 1.

a recirculation zone

2.

a burning zone (with a recirculation zone which extends to the dilution region)

3.

a dilution zone

Recirculation Zone

Burning Zone

Dilution Zone

Figure 8-16: A typical combustor can with straight through flow.

The air enters the combustor in a straight through flow or reverse flow. Most aero-engines have straight through flow type combustors. Most of the large frame type units have reverse flow. The function of the recirculation zone is to evaporate, partly burn, and prepare the fuel for rapid combustion within the remainder of the burning zone. Ideally, at the end of the burning zone, all fuel should be burnt so that the function of the dilution zone is solely to mix the hot gas with the dilution air. The mixture leaving the chamber should have a temperature and velocity distribution acceptable to the guide vanes and turbine. Generally, the addition of dilution air is so abrupt that, if combustion is not complete at the end of the burning zone, chilling occurs, which prevents completion. However, there is evidence with some chambers that, if the burning zone is run overrich, some combustion does occur within the dilution region. Figure 8-17 shows the distribution of the air in the various regions of the combustor. The theoretical or reference velocity is the flow of combustor-inlet air through an area equal to the maximum cross section of the combustor

| 301 |

| Centrifugal Compressors | casing. The flow velocity is 25 fps (7.6 mps) in a reverse-flow combustor and between 80 fps (24.4 mps)–135 fps (41.1 mps) in a straight-through flow turbojet combustor. Combustor inlet temperature depends on engine pressure ratio, load and engine type, and whether or not the turbine is regenerative or nonregenerative especially at the low-pressure ratios. The new industrial turbine pressure ratios are between 17:1 and 35:1, which means that the combustor inlet temperatures range from 850ºF (454ºC) to 1200ºF (649ºC). The new aircraft engines have pressure ratios that are in excess of 40:1. Combustor performance is measured by efficiency, the pressure decrease encountered in the combustor, and the evenness of the outlet temperature profile. Combustion efficiency is a measure of combustion completeness. Combustion completeness affects fuel consumption directly, since the heating value of any unburned fuel is not used to increase the turbine inlet temperature. The loss of pressure in a combustor is a major problem, since it affects both the fuel consumption and power output. A pressure loss occurs in a combustor because of diffusion, friction, and momentum. Total pressure loss is usually in the range of 2-8% of static pressure (compressor outlet pressure). The efficiency of the engine will be reduced by an equal percent. The result is increased fuel consumption and lower power output that affects the size and weight of the engine.

Figure 8-17: Air distribution in a typical combustor.

| 302 |

| Gas Turbines | The uniformity of the combustor outlet profile affects the useful level of turbine inlet temperature, since the average gas temperature is limited by the peak gas temperature. The profile factor is the ratio between the maximum exit temperature and the average exit temperature. This uniformity assures adequate turbine nozzle life, which depends on operating temperature. The average inlet temperature to the turbine affects both fuel consumption and power output. A large combustor outlet gradient will work to reduce average gas temperature and consequently reduce power output and efficiency. The traverse number is defined as the peak gas temperature minus mean gas temperature divided by mean temperature rise in nozzle design. Thus, the traverse number must have a lower value, between 0.05 and 0.15 in the nozzle. Equally important are the factors that affect satisfactory operation and life of the combustor. To achieve satisfactory operation, the flame must be self-sustaining, and combustion must be stable over a range of fuel-to-air ratios to avoid ignition loss during transient operation. Moderate metal temperatures are necessary to assure long life. Also, steep temperature gradients, which warp and crack the liner, must be avoided. Carbon deposits can distort the liner and alter the flow patterns to cause pressure losses. Smoke is environmentally objectionable as well as a fouler of heat exchangers. Minimum carbon deposits and smoke emissions also help assure satisfactory operations. The air entering a combustor is divided so that the flow is distributed between three major regions 1) Primary Zone, 2) Dilution Zone, 3) Annular space between the liner and casing. The combustion in a combustor takes place in the primary zone. Combustion of natural gas is a chemical reaction that occurs between carbon, or hydrogen, and oxygen. Heat is given off as the reaction takes place. The products of combustion are carbon dioxide and water. The reaction is stoichiometric, which means that the proportions of the reactants are such that there are exactly enough oxidizer molecules to bring about a complete reaction to stable molecular forms in the products. Normal combustion temperatures range from 3400ºF (1871ºC) to 3500ºF (1927ºC). At this temperature, the volume of nitric oxide in the combustion gas is about .01%. If the combustion temperature is lowered, the amount of nitric oxide is substantially reduced. Velocity is used as a criterion in combustor design, especially with respect to flame stabilization. The importance of air velocity in the primary zone is known. A transition zone is often included before the primary zone so that the high-velocity air from the compressor is diffused

| 303 |

| Centrifugal Compressors | to a lower velocity and higher pressure and distributed around the combustion liner. In the primary zone, fuel-to-air ratios are about 60:1; the remaining air must be added somewhere. The secondary, or dilution, air should only be added after the primary reaction has reached completion. Dilution air should be added gradually so as not to quench the reaction. Flame tubes should be designed to produce a desirable outlet profile and to last a long time in the combustor environment. Adequate life is assured by film cooling of the liner. The air enters the annular space between the liner and casing and is admitted into the space within the liner through holes and slots because of the pressure difference. The design of these holes and slots divides the liner into distinct zones for flame stabilization, combustion, dilution, and provides film cooling of the liner. The liner experiences a high temperature because of heat radiated by the flame and combustion. To improve the life of the liner, it is necessary to lower the temperature of the liner and use a material that has a high resistance to thermal stress and fatigue. The air-cooling method reduces the temperature both inside and outside the surface of the liner. This reduction is accomplished by fastening a metal ring inside the liner to leave a definite annular clearance. Air is admitted into this clearance space through rows of small holes in the liner and is directed by metal rings as a film of cooling air along the liner inside.

TYPICAL COMBUSTOR ARRANGEMENTS Can-annular and annular. In aircraft applications where frontal area is important, either can-annular or annular designs are used to produce favorable radial and circumferential profiles because of the great number of fuel nozzles employed. The annular design is especially popular in new aircraft designs; however, the can-annular design is still used because of the developmental difficulties associated with annular designs. Annular combustor popularity increases with higher temperatures or lowBtu gases, since the amount of cooling air required is much less than in can-annular designs due to a much smaller surface area. The amount of cooling air required becomes an important consideration in low-btu gas applications, since most of the air is used up in the primary zone and little is left for film cooling. Development of a can-annular design requires experiments with only one can, whereas the annular combustor must be treated as a unit and requires much more hardware and compressor flow.

| 304 |

| Gas Turbines | Can-annular combustors can be of the straight-through or reverse-flow design. If can-annular cans are used in aircraft, the straight-through design is used, while a reverse-flow design may be used on industrial engines. Annular combustors are almost always straight-through flow designs. Figure 8-18 shows a typical can annular combustor used in frame type units, with reverse flow.

Figure 8-18: A typical reverse flow can-annular combustor.

Tubular (side combustors). These designs are found on large industrial turbines, especially European designs, and some small vehicular gas turbines. They offer the advantages of simplicity of design, ease of maintenance, and long-life due to low heat release rates. These combustors may be of the “straight-through” or “reverse-flow” design. In the reverse-flow design, air enters the annulus between the combustor can and its housing, usually a hot-gas pipe to the turbine. Reverse-flow designs have minimal length. Figure 8-19 shows one such combustor design.

| 305 |

| Centrifugal Compressors |

Figure 8-19: A typical single can side combustor.

| 306 |

| Gas Turbines | AIR POLLUTION PROBLEMS Smoke. In general, it has been found that much visible smoke is formed in small, local, fuel-rich regions. The general approach to eliminating smoke is to develop leaner primary zones with an equivalence ratio between 0.9 and 1.5. Another supplementary way to eliminate smoke is to supply relatively small quantities of air to those exact, local, over-rich zones. Unburnt Hydrocarbons and Carbon Monoxide. Unburnt Hydrocarbon (UHC) and carbon monoxide (CO) are only produced in incomplete combustion typical of idle conditions. It appears probable that idling efficiency can be improved by detailed design to provide better atomization and higher local temperatures.

Oxides of Nitrogen The main oxides of nitrogen produced in combustion are NO, with the remaining 10% as NO2. These products are of great concern because of their poisonous character and abundance, especially at full-load conditions. The formation mechanism of NO can be explained as follows: •

Fixation of atmospheric oxygen and nitrogen at high-flame temperature.

•

Attack of carbon or hydrocarbon radicals of fuel on nitrogen molecules, resulting in NO formation.

•

Oxidation of the chemically bound nitrogen in fuel.

In 1977 the Environmental Protection Agency (EPA) in the US issued Proposed Rules that limited the emissions of new, modified, and reconstructed gas turbines to: •

75 vppm NOx at 15% oxygen (dry basis)

•

150 vppm SOx at 15% oxygen (dry basis), controlled by limiting fuel sulfur content to less than 0.8% wt.

These standards applied to simple and regenerative cycle gas turbines and to the gas turbine portion of combined cycle steam/electric generating systems. The 15% oxygen level was specified to prevent the NOx ppm level being achieved by dilution of the exhaust with air.

| 307 |

| Centrifugal Compressors | In 1977 it was recognized that there were a number of ways to control oxides of nitrogen:•

Use of a rich primary zone in which little NO formed, followed by rapid dilution in the secondary zone.

•

Use of a very lean primary zone to minimize peak flame temperature by dilution.

•

Use of water or steam admitted with the fuel for cooling the small zone downstream from the fuel nozzle.

•

Use of inert exhaust gas recirculated into the reaction zone.

•

Catalytic exhaust cleanup.

Wet control became the preferred method in the 1980s and most of the 1990s since dry controls and catalytic cleanup were both at very early stages of development. The catalytic converters were used in the 1980s and are still being widely used; however the cost of rejuvenating the catalyst is very high. There has been a gradual tightening of the NOx limits over the years from 75 ppm down to 25 ppm, and now the new turbine goals are 9 ppm. Advances in combustion technology now make it possible to control the levels of NOx production at source, removing the need for ‘wet’ controls. This, of course, opened up the market for the gas turbine to operate in areas with limited supplies of suitable quality water, e.g., deserts or marine platforms. Although water injection is still used, dry control combustion technology has become the preferred method for the major players in the industrial power generation market. DLN (Dry Low NOx) was the first acronym to be coined, but, with the requirement to control NOx without increasing carbon monoxide and unburned hydrocarbons, this has now become DLE (Dry Low Emissions). The majority of the NOx produced in the combustion chamber is called ‘thermal NOx.’ It is produced by a series of chemical reactions between the nitrogen (N2) and the oxygen (O2) in the air that occur at the elevated temperatures and pressures in gas turbine combustors. The reaction rates are highly temperature dependent, and the NOx production rate becomes significant above flame temperatures of about 3300ºF (1815ºC). Figure 8-20 schematically shows flame temperatures and, therefore, NOx production zones inside a “conventional” combustor. This

| 308 |

Next Page

| Gas Turbines |

NOx Production Zone

Figure 8-20: A typical combustor showing the NOx production zone.

design deliberately burned all of the fuel in a series of zones going from fuel-rich to fuel-lean to provide good stability and combustion efficiency over the entire power range. The great dependence of NOx formation on temperature reveals the direct effect of water or steam injection on NOx reduction. Recent research showed an 85% reduction of NOx by steam or water injection with optimizing combustor aerodynamics. In a typical combustor, as shown in Figure 8-20, the flow entering the primary zone is limited to about 10%. The rest of the flow is used for mixing the combusted air and cooling the combustor can. The maximum temperature is reached in the primary or stoichiometric zone of about 4040ºF (2230ºC), and, after mixing the combustion process with the cooling air, the temperature drops down to a low of 2500ºF (1370ºC).

| 309 |

9

Steam Turbines

Steam turbines are some of the most common drives for centrifugal compressors. In chemical plants, steam turbines are widely used, due to the fact that most of these plants need steam or generate it in their processes. Figure 9-1 is a centrifugal compressor driven by a steam turbine. Thus, by using steam turbines, the plant can utilize most of the excess steam thus using energy more efficiently. In the combined cycle utilities plants, combinations of gas turbines and steam turbines are being utilized to obtain high overall efficiencies. In this configuration, the hot gasses from the gas turbine are used to generate the steam for the steam turbine. A steam turbine may be defined as a form of heat engine in which the energy of the steam is transformed into kinetic energy by means of expansion through nozzles, and the kinetic energy of the resulting jet is, in turn, converted into force doing work on rings of blading mounted on a rotating part. The basic idea of steam turbines was conceived as early as 120 BC, yet it was not until 1883 that the first practical steam turbine was developed by De Laval. A typical steam turbine power plant is divided into its heat sources, the boiler or steam generator, and the turbine cycle, which includes the turbine, generator, condenser pumps, and feedwater heaters. The steam turbine operates on the Rankine Cycle.

| 359 |

| Centrifugal Compressors |

Figure 9-1: Steam turbine driving a centrifugal compressor (courtesy MAN Turbomaschinen AG GHH BORSIG).

THE RANKINE CYCLE The Rankine Cycle employs water-steam as the working fluid, and is the most common thermodynamic cycle utilized in the production of electrical power. A schematic of a steam power plant is shown in Figure 9-2. Water enters the boiler feed water pump at point 1 and is pumped isentropically into the boiler. The compressed liquid at 2 is heated until it becomes saturated at 2A, after which it is evaporated to steam at 2B and then superheated to 3. The steam leaves the boiler at 3, expands isentropically in the ideal engine to 4, and passes to the condenser. Circulating water condenses the steam to a saturated liquid at 1, from which state the cycle repeats itself. The thermodynamic diagrams corresponding to the steam power plant in Figure 9-2, showing the thermodynamic states, are shown in the pressure-volume (PV) diagram in Figure 9-3 and the temperature entropy (T-S) diagram in Figure 9-4.

| 360 |

| Steam Turbines |

Figure 9-2: Schematic of a steam turbine power plant.

Figure 9-3: Pressure-volume diagram of a typical steam turbine power plant.

| 361 |

| Centrifugal Compressors |

Figure 9-4: Temperature-entropy diagram of a typical steam turbine power plant.

The work done by the steam turbine Wst is given by Wst = m st (h3 − h4 )

(9-1)

where Ms is the mass flow of the steam and h3, h4 are enthalpies at point 3 and 4. The network produced by the system, Wnet, is the turbine work less the pump work, WP, required to raise the water to the desired pressure and is expressed below: Wnet = Wst - WP

(9-2)

Wnet = Ms (h3 - h4) - WP

(9-3)

The above analysis assumes an ideal isentropic expansion from point 3 to 4. In a steam turbine, the actual process is not isentropic, and some loss does occur. The actual expansion is from point 3 to 4. The pump work

| 362 |

| Steam Turbines | is much smaller than the turbine work and can be neglected when estimating the overall performance and efficiency of steam plants. The energy input requirement to the system, Qin, is given by . Qin = m st (h3 - hB)

(9-4)

The thermal efficiency of the system, η, is then given by: η

= WQnetin

η=

m st (h3 − h4 ) − W p m st ( h3 − hB )

(9-5)

(9-6)

Preheating boiler feed water and the incoming combustion air by using hot exhaust from the boiler, or gas turbine if available, can increase the overall efficiency of the system.

THE REGENERATIVE–REHEAT CYCLE It is evident from the Rankine Cycle (shown in Figure 9-2) that a considerable amount of heat is required to raise the temperature of the water from 2 to 2a. The Rankine Cycle has the disadvantage that the fluid temperature at the pump discharge is much lower than the fluid temperature at the turbine inlet. One way of overcoming this disadvantage is to use the internal system heat rather than the external heat to minimize this difference in the temperatures. This concept is called regenerative heating. In the gas turbines, the regenerative heating is accomplished by using the high temperature exhaust gases. In the steam turbines, intermediate pressure steam rather than exhaust steam is used for heating feed water. As the high-pressure steam progresses through the steam turbine, the steam gets very wet at low pressures. This wet steam is detrimental for a turbine; it results in reduction of efficiency and also nozzle and blade erosion. The reheat cycle involves heating of the steam withdrawn after partial expansion. This idea, combined with regenerative heating for improved thermal efficiency, is common practice in central power plants. A simplified concept of the regenerative reheat steam cycle is depicted in Figure 9-5 and Figure 9-6. The water enters the first pump at point 1 from which it enters the feedwater heater at point 2. In the

| 363 |

| Centrifugal Compressors |

Figure 9-5: Schematic of a regenerative-reheat steam turbine power plant.

Figure 9-6: Temperature-entropy diagram of a regenerative-reheat steam turbine power plant.

| 364 |

| Steam Turbines | feedwater heater, the pressurized condensate is heated by part of the steam extracted from the high-pressure turbine at an intermediate pressure, point 6. The rest of the extracted steam is reheated in the reheater and enters the turbine at point 7. The heated water enters the second pump at point 3, from which it enters the boiler at point 4. The compressed liquid at 4 is heated until it becomes saturated at 4a, after which it is evaporated to steam at 4b, and then superheated to 5. The steam leaves the boiler at 5, expands isentropically in the ideal engine to 6, and in the real case to 6a where it is extracted for regeneration and reheat. The superheated steam now leaves the reheater at 7, expands isentropically in the ideal engine to 8 and in the real case to 8a where it passes to the condenser. Circulating water condenses the steam to a saturated liquid at 1, from which state the cycle repeats itself. H E AT R AT E A N D S T E A M R AT E The heat rate is a modified reciprocal of the thermal efficiency and is in much wider use among steam-power and turbine engineers. The heat rate for a turbine is defined as the heat chargeable in Btu per kilowatthour of horsepower-hour turbine output. Again, the basis upon which the turbine output is taken should be specified. Turbine heat rate should not be confused with the heat rate of the steam-power plant known as the station heat rate. The station heat rate, like station thermal efficiency takes into account all the losses from fuel to switchboard. The heat rate for a straight condensing or non-condensing turbine is: HR =

(h3 − h f 2 )3415 W net

Btu per kw-hr

(9-7)

or HR =

(h3 − h f 2 ) 2544.4 W net

(9-8)

where h3

= enthalpy of throttle steam

hf 2 = enthalpy of liquid water at exhaust pressure Wnet = output on any specified basis, Btu per lb of steam at throttle

| 365 |

| Centrifugal Compressors | Steam rate is defined as the mass rate of steam flow in pounds per hour divided by the power or rate of work development of the turbine in kilo-power-hour. The steam rate, therefore, is the steam supplied per kilowatt-hour or horsepower-hour unit of output. The heat rate may be obtained by multiplying the steam rate by the heat chargeable. STEAM TURBINE The usual turbine consists of four basic parts: the rotor, which carries the blades or buckets; the stator, consisting of cylinder and casing, which are often combined and within which the rotor turns; the nozzles or flow passages for the steam, which are generally fixed to the inside of the cylinder; and the frame or base for supporting the stator and rotor. In small turbines, the cylinder casing and frame are often combined. Several other systems, such as the lubrication systems, steam-piping systems, and in some cases, a condensing system, make up the rest of the turbine. Steam turbines used as drives for centrifugal compressors should have the following major operating criteria: •

Ability to operate over a wide range of steam flows

•

High efficiency over a large operating range

•

Reheat possibilities

•

Fast start up

•

Short installation time

•

Floor-mounted installations

The steam turbine should be able to operate over a wide range of compressor loadings. Thus, this requires a high efficiency over a wide operating range. Rapid start up is desirable. Great care in design must be exercised due to rapid increase in temperatures in the rotor and the shaft. This does not allow for rotor wheels, which use a shrink fit onto the shaft. There have been cases, during rapid start up, of the rotor wheel “walking” on the shaft due to the different growth rates between the shaft and the rotor wheel. Shrink fits should not be used on rotors, which have tip speeds over 700 ft/sec and power requirements of over 1200 HP.

| 366 |

| Steam Turbines | CLASSIFICATIONS OF STEAM TURBINES Several classification schemes for steam turbines are available. The following is a major classification scheme: A. Steam flow direction 1.

Axial flow turbines

2.

Radial flow turbines

3.

Tangential flow turbines

4.

Mixed flow turbines

B. Steam passage between blades 1.

Impulse turbines

2.

Reaction or Parsons turbines

3.

Impulse and Reaction Combination

C. Arc of peripheral admission to the total circumference. 1.

Steam Turbine Nozzles

2.

Full admission turbines

3.

Partial admission turbines

D. Turbine stages in series 1.

Single-stage or simple impulse turbine

2.

Multi-stage turbine

E. Type of stage

F.

1.

Pressure stage, Rateau type impulse

2.

Velocity stage, Curtis type impulse

3.

Pressure and velocity stage combination

General flow arrangement 1.

Single flow

2.

Double flow

3.

Compound flow

4.

Extraction flow

| 367 |

| Centrifugal Compressors | STEAM FLOW DIRECTIONS Steam enthalpy is converted into rotational energy as it passes through a turbine stage. A turbine stage consists of a stationary blade (or nozzle) and a rotating blade (or bucket). Stationary blades convert the potential energy of the steam (temperature and pressure) into kinetic energy (velocity) and direct the flow onto the rotating blades. The rotating blades convert the kinetic energy into impulse and reaction forces, caused by pressure drop, which result in the rotation of the turbine shaft or rotor.

Axial Flow Axial flow signifies steam flow substantially parallel to the axis of rotation, among blades that are set radially. This is the only arrangement used in medium and large turbines and is most commonly used also in small turbines. It provides opportunity for almost any desired degree of expansion of the steam by increase in the length of the blades and the diameter at which they rotate, coupled with increase in the number of rows of blades.

Radial Flow Radial flow is obtained when the steam enters at or near the shaft and flows substantially radially outward among blades, which are placed parallel to the axis of rotation. In one turbine of this type, the Ljongstrom turbine, as seen in Figure 9-7, successive rings of blades are attached alternately to two discs mounted on separate rotors. These rotors are turning in opposite directions on the same axis and driving separate generators.

Tangential Flow Tangential flow is the term applied when the steam enters through a nozzle placed approximately tangent to the periphery and directed into semicircular buckets milled obliquely into the edge of the wheel. Coupled with this is the action of a reversing chamber fitted closely to, but not touching, the periphery of the wheel, which has similar buckets milled into it. These latter chambers receive the steam discharged from the wheel buckets and return it again a number of times to the wheel buckets in the proper direction to produce additional work, the steam following an approximately helical path. Any number of nozzles may be used, each with its reversing chamber, up to that number which will completely fill the periphery. This type of flow is also called helical.

| 368 |

| Steam Turbines |

Figure 9-7: Radial flow or Ljongstrom turbine.

Mixed Flow Mixed flow is the term applied to the flow when it enters in the radial direction and leaves in the axial direction. These types of radial inflow turbines have been widely used with gasses and, in some cases, with steam. S T E A M PA S S A G E B E T W E E N B L A D E S The most common type of flow for large steam turbines is axial flow. The flow is further divided into two categories, impulse and reaction. Impulse and reaction turbine designs are based on the relative pressure drop across the stage. There are two measures for pressure drop, the

| 369 |

| Centrifugal Compressors | pressure ratio and the percent reaction. The percent reaction is the percentage of isentropic enthalpy drop across the rotating element, as a percent of the total stage enthalpy drop. Some manufacturers utilize percent pressure drop across stage to define reaction.

Impulse Turbines Impulse turbines may be defined as a system in which all steam expansion takes place in fixed nozzles and none occurs in passages among moving blades. In actual practice, there must be some pressure drop across the rotating blades to generate flow. A typical HP stage is usually an impulse stage but still has, on the average, about 5% reaction at full load.

Reaction or Parsons Turbine In the turbines so far described, the steam expands only in fixed nozzles and flows through passages between blades arranged in rows, transferring its kinetic energy to these rows of blades and causing them to rotate against resistance. In the widely used reaction turbine proposed and first built by Sir Charles Parsons, the steam decreases in pressure and expands while it is passing through the moving blades as well as in its passage through the fixed nozzles. In a symmetric reaction design, equal pressure drop occurs across the stationary and rotating blades. This amounts to a 50% reaction turbine. The root and tip reactions are often different, especially in long blades, to counter the effect of centrifugal forces on the steam flow. The root is often more impulse, and the tip is nearer 50% reaction.

Impulse and Reaction Combination This is the most common type of steam turbine on the market. The first stages are usually impulse type, while the later stages are usually reaction types (50%). The impulse turbine wheel produces about two times the power output as the 50% reaction type turbine wheel. The reaction turbine, on the other hand, is more efficient. The combination of the first stages being impulse and the later stages being about 50% reaction produces a high power and efficiency turbine in the most compact passage.

| 370 |

| Steam Turbines | ARC OF PERIPHERAL ADMISSION TO THE TOTAL CIRCUMFERENCE FULL ADMISSION TURBINES One aspect of the classification of steam turbines, as mentioned in the classification scheme, is according to the ratio of the arc of peripheral admission to the total circumference. If steam admission is over the full arc we have a full admission turbine. In any design of a steam turbine, two types of losses must be considered. They are the losses encountered by the Mach number and the losses encountered by partial admission. These losses are shown in Figures 9-8 and 9-9, where you can sees that for high Mach numbers, the loss factor increases, but it is more or less constant up to approximately Mach 1. The maximum limit for efficient design is around Mach number 1.15. Excessive mach numbers result in severe shocks, choking of passages and high Mach number losses.

Figure 9-8: Steam turbine losses as a function of Mach number.

| 371 |

| Centrifugal Compressors | Figure 9-9 shows the relationship between efficiency and different configurations for partial entry of the flow. Usually, low admission ratios result in high partial admission losses.

Figure 9-9: Partial entry steam turbine performance properties.

PA RT I A L A D M I S S I O N T U R B I N E S Usually, in the high-pressure end of a steam turbine, the nozzles are located only partially around the circumference (partial admission). At the intermediate stages the increasing volume of steam necessitates that larger circumferential arcs be occupied by nozzles until a full admission situation is reached. It is the balance between the Mach number and the partial entry losses, which finally determines the design. T U R B I N E S TA G E S I N S E R I E S There are two major types of turbines, the impulse turbine and the reaction turbine. The steam volume increases whenever the pressure decreases, but the resulting velocity changes depend on the type of turbine. These velocity changes are distinguishing characteristics of the different types of turbines.

| 372 |

| Steam Turbines | The degree of reaction in an axial-flow turbine is defined as the ratio of the change of enthalpy drop in the rotor to the change in total enthalpy drop across the stage: R=

H rotor H stage

(9-9)

By definition, the impulse turbine has a degree of reaction equal to zero. This degree of reaction means that the entire enthalpy drop is taken in the nozzle and the exit velocity from the nozzle is very high. In practice there must be a pressure drop across the rotating blades to generate flow. Since there is no change in enthalpy in the rotor, the relative velocity entering the rotor equals the relative velocity exiting from the rotor blade. Most steam turbine high pressure (HP) stages are typically impulse stages by design but average a 5% reaction at full load. For a symmetric flow (50% reaction), the enthalpy drop in the rotor is equal to the drop in the stationary part of the turbine. This also leads to equal pressure drop across the stationary and rotating parts. Due to the difference in the turbine blade diameters at the tip and the root of the blade, the reaction percentages are different to counteract the centrifugal forces acting on the steam flow. If this were not the case, too much flow would migrate to the blade tips. In intermediate pressure (IP) and low pressure (LP) turbines, the basic design is reaction; however, in these turbines, about 10% of the reaction is at the root and at the tip, going from about 60% reaction to about 70% reaction at the tips in the LP turbine. S I N G L E - S TA G E O R S I M P L E - I M P U L S E T U R B I N E The simple impulse turbine is one of the simplest forms of a turbine. Here the steam expands from its initial to its final pressure in one nozzle (or one set of nozzles, all working at the same pressure), resulting in a steam jet of high velocity, which enters the blade passages and, by exerting a force on them due to being deflected in direction, turns the rotor. Energy of all forms that remains in the steam after it leaves the single row of blading, is lost. A simple stage impulse turbine is shown in Figure 9-10. The absolute velocity, and the total and static pressure changes, as the steam is accelerated through the nozzle and impinges on the turbine wheel is shown in the figure. The total pressure and temperature remain unchanged in the nozzle, except for minor frictional losses.

| 373 |

| Centrifugal Compressors |

Figure 9-10: A simple impulse steam turbine.

M U LT I - S TA G E I M P U L S E T Y P E T U R B I N E In this type of turbine, there are as many steps as there are chambers, each being called a pressure stage. The resultant steam velocity in each stage is relatively small, allowing reasonably low blade velocities and preventing excessive loss by steam friction. The pressure drops in each stage, and the steam volume increases. The steam velocity is high at exit from the nozzles and is low at exit from the blades. This arrangement is sometimes termed a Rateau Turbine, and the separate stages are called Rateau Stages. VELOCITY AND P R E S S U R E S TA G E C O M B I N AT I O N Turbines with combinations of velocity and pressure staging are commonly used and are of several types. The wheel in each pressure stage might have two (or even three) of rows of blades instead of one. The turbine has as many pressure stages as there are wheels, and each pressure stage has as many velocity stages as there are rows of blades on the wheel in that stage. This arrangement results in a small, short, and low

| 374 |

| Steam Turbines | cost turbine, at more or less sacrifice of efficiency. Commercial turbines of this type are called Curtis turbines, after the original inventor. The individual pressure stages, each with two or more velocity stages, are often called Curtis stages.

Velocity-stage, Curtis Type Impulse Turbine The Curtis type turbine has one set of nozzles with several rows of blades following it. In passing from the nozzle exit through one set of blades, the velocity of the steam is lowered by virtue of the work done on the blades but is still high. It then passes through a row of fixed guide blades, which change the direction of the steam until it flows approximately parallel to the original nozzle direction, discharging it into a second row of blading fixed to the same wheel. This second row again lowers the steam velocity by virtue of the work delivered to the wheel. A second set of guide blades and a third row of moving blades are sometimes used. Figure 9-11 shows a Curtis type stage, a velocity compounded turbine in which there are several moving rows to absorb the kinetic energy coming from the nozzles. The absolute velocity, with the total and static pressure distribution, is also shown in this figure.

Figure 9-11: A Curtis type steam turbine.

| 375 |

| Centrifugal Compressors | P R E S S U R E - S TA G E , R AT E A U T Y P E IMPULSE TURBINE Another impulse turbine is the pressure compound, or Ratteau turbine. In this turbine the work is broken down into various stages. Each stage consists of a nozzle and blade row where the kinetic energy of the jet is absorbed into the turbine rotor as useful work. The steam, which leaves the moving blades, enters the next set of nozzles where the enthalpy decreases further, and the velocity is increased and then absorbed in an associated row of moving blades. The turbine has a series of chambers formed by parallel disc-shaped partitions called diaphragms and has a simple impulse turbine enclosed in it, all wheels being fastened to the same shaft. Each chamber receives the steam in turn through groups of nozzles placed on arcs. The last chamber, in most cases, discharges to the condenser. The pressure drop is divided into the various stages. Figure 9-12 shows a Ratteau turbine.

Figure 9-12: Pressure compounded steam turbine-Ratteau Turbine.

| 376 |

| Steam Turbines | REACTION TYPE TURBINE Figure 9-13 shows a reaction type turbine. In this turbine, the blade passages are shaped so that another convergent nozzle is formed. The nozzle will give an additional force due to the acceleration of the jet, and this is the basis of the reactive principal.

Figure 9-13: Reaction turbine.

GENERAL FLOW ARRANGEMENT Although early machines were one section or cylinder (because of design and manufacturing limitations), most commonly power plant turbines consist of two or more sections designated high pressure (HP), intermediate pressure (IP), and low pressure (LP) in order to improve overall efficiency. The optimum number of stages depends on a number of factors including: •

The amount of available energy from boiler conditions, allowable discharge pressures, and temperatures

| 377 |

| Centrifugal Compressors | •

Stage percent reaction, impulse rotor (near zero reaction) produces double the energy of a 50% reaction rotor, the rotor mean diameter, and rotor speed.

There are five basic types of shaft and casing arrangements: single casing, tandem compound, cross compound, double flow, and extraction steam turbines. Figure 9-14 shows schematics of various steam turbine arrangements. SINGLE-FLOW SINGLE CASING TURBINES In a single flow turbine, the steam enters at one end, flows once through the blading in a direction approximately parallel to the axis, emerges at the other end, and enters the condenser. In turbines with a single casing, all sections are contained within one casing and the steam path flows from throttle to exhaust through that single casing. Figure 914a shows a simple path where steam enters a turbine and is exhausted to the atmosphere or a condenser it also shows an extraction for cogeneration purposes. This is the most common arrangement in small and moderately large turbines.

Figure 9-14: A schematic of arrangements of steam turbines.

| 378 |

| Steam Turbines | COMPOUND-FLOW O R TA N D E M C O M P O U N D T U R B I N E Compound-Flow or Tandem Compound Turbine is the term applied to a machine in which the steam passes in sequence through two or more separate units, expanding in each. The two units arranged in a tandem compound design have both casings on a single shaft and driving the same electrical unit. In most cases the HP exhaust is returned for reheating before entering the IP turbine. Most often the high-pressure and the intermediate-pressure portions are in one casing and the low-pressure portion in another. The IP exhausts into crossover piping and on to one or more low-pressure turbines. In Figure 9-14b, two sections are used, an HP and an LP section. In Figure 9-14c indicates a condition where the split stream is used in a double flow, low-pressure (DFLP) section. Figure 9-15 shows a typical tandem-compound turbine with two casings—one HP and IP and the second, a two-flow LP.

Figure 9-15: A typical compound turbine.

CROSS COMPOUND TURBINE A cross compound design typically has two or more casings, coupled in series on two shafts, with each shaft connected to a generator. In cross compound arrangements, the rotors can rotate at different speeds but cannot operate independently as they are aerodynamically coupled. The cross compound design is inherently more expensive than the tandem

| 379 |

| Centrifugal Compressors | compound design but has a better heat rate, so that the choice between the two is one of economics. Figure 9-14d shows a turbine setup where there are three casings, each with their own generator, and a reheat between the HP and the IP Section. DOUBLE-FLOW TURBINES LP turbines are typically characterized by the number of parallel paths available to the steam. The steam path through the LP turbines is split into parallel flows because of steam conditions and practical limitations on blade length. Typically, several LP flows in parallel are required to handle the large volume of flow rates. The steam enters at the center and divides. The two portions pass axially away from each other through separate sets of blading on the same rotor. This type of unit is completely balanced against end thrust and gives a large area of flow through two sets of lowpressure blading. EXTRACTION FLOW TURBINE The extraction flow stream turbine is the term applied to a turbine where part of the flow is extracted for various reasons such as steam for the plant, for absorption type chillers, or for any other plant process. Figure 9-14a shows a schematic of an extraction type turbine. These turbines maybe back pressure or condensing depending on the application. These types of turbines are used most commonly in petrochemical plants

STEAM TURBINE CHARACTERISTICS Understanding the effect of the steam operating conditions on efficiency and load is very important in operating steam turbines at their optimum operating conditions. The two types of steam turbines are condensing steam turbines and the backpressure steam turbines. Steam inlet temperature and pressure and turbine exhaust pressure and vacuum are the significant operating parameters of a steam turbine. The variations in these parameters affect steam consumption and efficiency. Turbine Steam Inlet Pressure is a major parameter, which affects turbine performance. To obtain the design efficiency, the steam inlet pressure should be maintained. Lowering steam inlet pressure reduces

| 380 |

| Steam Turbines | turbine efficiency and increases steam consumption. A 10% increase in steam pressure will reduce the steam consumption by about 1% in a condensing steam turbine and will reduce the steam consumption by about 4% in a backpressure steam turbine. The effect on efficiency for 10% increase in pressure for a condensing steam turbine is about 1.5 % and 0.45% for a backpressure steam turbine. Turbine steam inlet temperature is another major parameter that affects turbine performance. Reducing steam inlet temperature reduces the enthalpy, which is a function of both the inlet temperature and pressure. At higher steam inlet temperature, the heat extraction by the turbine will also be increased. An increase of about 100ºF (55ºC) will reduce the steam consumption by about 6.6% in a condensing steam turbine and 8.8% in a backpressure turbine. The effect on efficiency for a 100ºF (55ºC) will be an increase of 0.6% in efficiency for a condensing steam turbine and 0.65% in efficiency for a backpressure turbine. It should be noted that the overall efficiency, in most cases, for a condensing steam turbine (30%–35%) is about twice that of a backpressure turbine (18%–20%). In condensing or exhaust backpressure steam turbines, the increase of this backpressure will reduce the efficiency and increase the steam consumption, keeping all other operating parameters. In condensing steam turbines, the condenser vacuum temperature will also increase if the removal of heat from the condenser is reduced. Thus in a water-cooled condenser, if the temperature of the inlet water is increased, the power produced by the turbine is decreased because the backpressure will be increased. In summary, the condensing steam turbines are more efficient and produce more power than backpressure steam turbines. The condensing steam turbine is also more efficient. The cost of a condensing steam turbine is about $25/kw more than a backpressure turbine.

| 381 |

| Centrifugal Compressors | FEATURES AND STRUCTURE OF HP AND IP BLADES HIGH-PRESSURE (HP) AND I N T E R M E D I AT E - P R E S S U R E ( I P ) R O TAT I N G B L A D E S The blades in the HP turbine are small because of the low volumetric flow. Rotating HP blades are usually straight, however, the use of leaned and bowed blades has recently introduced a three dimensional aspect to designs. Shrouds (also called covers or connecting bands) provide a sealing surface for radial steam seals and are used to minimize losses due to leakage. Shrouds also tie the blades together structurally and allow for some control over the damping and natural frequencies of the blades. The shrouds are typically attached either by peened tension or are integral with the blade airflow (or vane). Blades are connected at the root to the rotor or disc by several configurations. Figure 9-16 shows the most common types of root attachments. The choice of type of attachment will depend upon a number of factors. A side-entry, fir tree root design is used in the HP control stage for ease of replacement, if required because of solid particle erosion. For longer blades in the control stage, however, a triple pin construction is sometimes used because the side-entry design has too many modes close to the nozzle wake frequency. The blade roots may be of the serrated or fir tree configuration, inserted into individual axial slots in the disc, a similar serrated or T-shape inserted into a continuous circumferential slot in the disc (this requires a special insertion gap), or may comprise one or several flat “fingers” fitting into circumferential slots in the disc and secured by axially inserted pins. Furthermore, serrated or T-roots may be of male or female type. A particular challenge in HP blading design is the first (control) stage where operation with partial arc admission leads to high dynamic stresses. Design factors, such as choice of leading edge configuration and blade groupings, are chosen to reduce vibratory stresses produced. Blades in IP turbines are very similar in design to those in the HP, with somewhat more twist, plus bowing and leaning to account for greater radial variation in the flow, thus providing flow radial equilibrium.

| 382 |

| Steam Turbines |

Figure 9-16: Typical types of blade roots (courtesy EPRI Turbine Steam Path Damage. Theory and Practice. )

| 383 |

| Centrifugal Compressors | HP AND IP NOZZLES In HP and IP turbines, stationary blades (or nozzles) can be classified into two general design categories: a wheel and diaphragm construction is used for impulse stages and a drum-rotor construction is used for reaction stages. A diaphragm, used in impulse stages, consists of •

nozzles or stationary blades

•

a ring which locates them in the casing

•

a web that extends down between the rotor wheels and supports the shaft packing

Figure 9-17 indicates the typical construction. Diaphragms in the HP and IP are typically of welded construction. In the control (first) stage, nozzles are divided into segments, arranged in separate nozzle “chests” or “boxes”, and each segment has an associated control valve or control valve group. In reaction stages, stationary blades or nozzles are manufactured in a manner similar to that for rotating blades with a root attachment and, in

Figure 9-17: Construction of a typical steam turbine diaphragm (courtesy EPRI Turbine Steam path damage. Theory and Practice).

| 384 |

| Steam Turbines | some cases, a sealing shroud. The blades are fitted by the root attachment on a blade carrier, which is located in the outer casing. Nozzles and diaphragms are typically exposed to pressure differentials, which bend them in the plane perpendicular to the turbine axis. These pressure differentials are highest in the HP, although the shorter blade length limits the bending stresses that develop. F E AT U R E S A N D S T R U C T U R E O F L P B L A D E S Figure 9-18 shows the nomenclature for a rotating LP blade. Rotating LP turbine blades may be “free standing,” that is, not connected to each other in any way. They may be connected in “groups” or “packets” each comprising several blades, generally between 2 and 8, or all blades in the whole row may be “continuously” connected.

Figure 9-18: Nomenclature for a typical LP turbine blade (courtesy EPRI Turbine Steam path damage. Theory and Practice).

| 385 |

| Centrifugal Compressors | Freestanding blades have the following characteristics compared to grouped blades: •

They have less inherent damping at the blade tips

•

Their resonances are more easily defined, (i.e. no mechanical interactions with neighboring blades)

•

They have more aerodynamic interactions

•

They are easier to install and disassemble, as there are no welds or rivets

In connected blades, there are a number of design choices. The connections may consist of shrouds (or bands) over the tips of the blades or of tie-wires (or lashing or lacing wires) located along the blade height. As with the HP blades, connections made at the blade tip are termed shrouding. Shrouds may be inserted over tenons protruding above the blade tips, with tenons riveted down to secure the shrouds. They also may consist of integrally forged stubs in a welded or brazed together assembly. Other types of riveted connections are also used. Tie-wires may consist of either integrally forged stubs welded or blazed together or cylindrical “wires” or rods inserted through a hole (usually in a forged boss) in each blade foil. In order to add mechanical damping, some wire or rod-type lashing wires are left loose in their holes, and there are also some shroud-type connections that merely abut each other and are not permanently attached. Continuous connections must make some provision to accommodate thermal expansion. Shrouds and lacing wires sometimes are introduced to decrease vibratory stresses but can act as “dirt traps.” The airfoil of blades may be of constant section for short blades and constant width, but twisted, for longer ones. The longest blades for the last few rows of the LP are twisted to match the aerodynamics at different radii and improve aerodynamic efficiency. Most recently designs have also begun to be leaned or bowed, thus introducing radial variation as well. Nozzles or stationary blades in LP stages are typically arranged in diaphragms like those for HP and IP impulse stages. However, the construction may be simpler than those in the HP and IP, consisting, for example, of only fixed blades constrained by inner and outer (hub and rim) annular bands. Diaphragms in the LP are of cast or welded construction. In wet stages, diaphragms may be made with hollow blade vanes or

| 386 |

| Steam Turbines | other design features as a means of drawing off moisture that would otherwise lead to liquid droplet erosion. R E Q U I R E D M AT E R I A L C H A R A C T E R I S T I C S Choosing the optimum blade material is an ongoing tradeoff between desirable material characteristics. The demands placed on HP and LP blades emphasize different material characteristics, as seen in Table 9-1. In addition, it is important that blading material be weldable, particularly last stage LP blading, as many designs require that cover blades, tie-wires, and erosion shields be attached by thermal joining. Weldability is also important for blade repairs. It is important to note that up to 1990 many major manufacturers considered the welding of blades and rotors impossible.

Material Characteristics

HP and IP Turbine Blades

Creep Strength and Creep Fatigue Resistance x Tensile Strength x Corrosion Resistance Ductility and Impact Strength x Fatigue Strength x Corrosion Fatigue Resistance Notch Sensitivity Material Damping x Partial Admission Blades Erosion Resistance x Solid Particle Erosion

LP Turbine Blades x x x x x x x x Liquid Drop Erosion

Table 9-1: Important blade material characteristics.

•

Creep strength and creep-fatigue resistance. Creep resistance is important in the first two or three rows of the HP and IP turbines to resist elongation and accumulation of strain at the higher operating temperatures, particularly at stress concentrations such as at the blade-to-rotor attachments.

•

Tensile strength. Tensile strength is required to withstand steady centrifugal and steam bending loads.

| 387 |

| Centrifugal Compressors |

| 388 |

•

Corrosion resistance. Corrosion resistance is important to maintain blade life in the turbine environment.

•

Ductility and impact strength. Ductility is required for three reasons: –

To allow for plastic deformation that can occur from blade rubs or foreign body impact

–

To allow for rivet formation where blades are attached to shrouds

–

To allow localized plastic flow to relieve stress peaks and concentrations, which can occur at the blade root and interblade connections such as lashing wires and shrouds

•

Fatigue strength. Fatigue strength is important to prevent failures from the vibratory stresses imposed by steam flow and system resonance. There are two types of fatigue failures—high cycle low stress and high stress and low cycle failures.

•

Corrosion fatigue resistance. In LP blades, even more important than simple fatigue strength is the resistance of the material to cyclic loads in aggressive or corrosive environments in order to avoid corrosion fatigue.

•

Notch Sensitivity. Notch sensitivity is the effect of stress concentration on fatigue strength. Low notch sensitivity is a desirable characteristic; however, high tensile strength materials that are good for improved strength also lead to high notch sensitivity.

•

Erosion Resistance. The HP and IP stages suffer from erosion by solid particles which are formed by oxides exfoliated from super heater and reheat tubing due to high temperature operation. Liquid erosion occurs mostly in the LP stage where they are subjected to liquid droplet erosion. To avoid this, the amount of water in the steam, especially on the last LP stages, must remain below 10%.

•

Blade Damping. Damping is of primary concern in any blade design. There are three types of damping:

| Steam Turbines | –

Material damping is an inherent part of the blade material and is dependent on the preload stress, dynamic stress, temperature, and frequency.

–

Mechanical or Coulomb friction damping occurs from the relative motion between contacting parts.

–

Aerodynamic damping occurs as a result of work done on the gas stream or by the work done by the gas stream on the rotating airfoil.

B L A D E M AT E R I A L S The most common material for HP and IP rotating and stationary blades and nozzles is 12Cr martensitic stainless steel. Three generic martensitic stainless steels are widely used for turbine blading, most commonly Type AISI 422 for HP blading and Types AISI 403 and AISI 410 in LP blading. There are numerous specific applications materials where turbine manufacturers have customized the generic grade by the addition or deletion of specific alloying elements or by modification of the production or heat-treating process. The final properties of these steels are strongly influenced by tempering temperature. The austenitic stainless steels (AISI series 300) are used in some high temperature applications. The austenitic stainless steels, which have a higher content chromium and tungsten than the martensitic stainless steel, have excellent mechanical properties at elevated temperatures and are typically readily weldable. There is a thermal expansion coefficient difference between martensitic and austenitic stainless steels so that care is required when designing attachment clearances for fitting austenitic blades into martensitic discs. Also there is a potential for stress corrosion cracking when 300 series stainless steels are used in wet steam conditions. The rings and webs of HP and IP nozzle diaphragms are commonly manufactured from stainless steels, although, if the working steam temperature does not exceed 350°C (660°F), the welded diaphragms can be made from carbon steels. Early LP blades materials included cartridge brass (72% Cu, 28% Zn) nickel brass (50% Cu, 10% Ni, 40% Zn), and Monel (66% Ni, 31% Cu, 3% Fe). With the advent of larger turbines, most LP turbine blades have been manufactured from a 12% Cr stainless steel; typically Types AISI 403, 410, or 410-Cb have been chosen depending on the strength required. Types 403

| 389 |

| Centrifugal Compressors | and 410 have better corrosion resistance than Type 422, an important characteristic for use in the wet stages of the LP turbine. There are numerous specifically customized versions of these generic materials, for example, Carpenter H-46 and Jethete M152. Jethete M152 has higher hardness and is thus more resistant to liquid droplet erosion in the LP than Types 403 and 410. So far, it has only been used in LP turbines, but could be used in the HP and IP if needed. European designations for 12% Cr blading alloys include X20CrMoV121 and X20Cr13. More recently, the precipitation hardened stainless steel designated 17-4 PH (AISI 630) was developed by one manufacturer for the last blades of the LP turbine in the largest 3600 rpm machines. It has a nominal composition that is 17% Cr and 4% Ni. The hardening temperature can control a wide range of mechanical properties. Alloy 17-4 PH is somewhat difficult to weld and requires post-weld heat treatment. Titanium alloys, chiefly Ti-6A1-4V (6% aluminum and 4% vanadium), have been used for turbine blades since the early 1960s. The use of titanium in the last few rows of the LP offers a number of advantages over other materials. The advantages to titanium include:

| 390 |

•

Titanium has about half the density of 12Cr steels which allows for longer last stage blades without an increase in centrifugal stresses in the blades and thus an increase in annular area and improved turbine efficiency. The capability of LP turbines to produce power is limited by the long last row of blading and the strength of the rotor to support the blade. The practical limitation for blades constructed of 12% Cr martensitic steel was reached with 840 mm (33.5 in.) blades operating in 3600 rpm machines and 1200 mm (48 in.) blades operating in 3000 rpm machines. In contrast, titanium offers an opportunity to go to 1000 mm (40 in.) and 1350 mm (54 in.) blades for 3600 rpm and 3000 rpm machines respectively. This represents a marked increase in power and makes possible a new generation of LP steam turbines.

•

Titanium has particularly favorable mechanical characteristics in applications involving high stresses at low temperatures. Because titanium has half the density and about half the elastic modulus of steel, the frequencies and mode shapes of titanium blades are very similar to those made of steel. Note however, that the elastic modulus is dependent on the particular titanium composition.

| Steam Turbines | •

Titanium has greater corrosion resistance and, as a result, may have better performance in dry/wet transition phase regions on the LP.

•

Titanium also has excellent resistance to impact and water droplet erosion damage and, in many applications, can be used without erosion shields.

The drawbacks to titanium include: •

Titanium has a higher cost that steel, even though titanium’s lower density means that more blades can be manufactured for a given mass of material which somewhat offsets the higher cost per pound of the material.

•

Titanium is more difficult to machine.

•

Titanium is more difficult to weld. Titanium requires a high state of cleanliness and an inert welding atmosphere.

•

Titanium has poor resistance to sliding wear, which can allow fretting corrosion in some conditions, although fretting has not been found to be as much of a problem as was once anticipated.

•

Titanium has lower internal material damping than stainless steel.

•

A major disadvantage of titanium for blades is that recent high cycle fatigue studies have shown an endurance limit in air and in steam smaller than for 12% Cr stainless steels. Shot peening has been used to restore the fatigue life lost after machining, production, or repair processes.

Duplex stainless steels are those stainless steels that contain very high levels of chromium and about equal amounts of ferrite and austenite. They have been evaluated for use for LP blading, primarily in Europe. There are a variety of types of duplex stainless steels with ferrite contents ranging from about 45-75. The duplex stainless steels have excellent corrosion fatigue characteristics. Their primary drawback for blading may be somewhat lower yield strength than Type 403/410 in general, (although the specific characteristics for the Ferralium alloy indicate that good yield and tensile strength can be achieved), and they do show some long time service embrittlement at temperatures about 300°C (570°F).

| 391 |

| Centrifugal Compressors | S U R FA C E T R E AT M E N T S Coatings of surface hardening are frequently used to improve the surface characteristics of turbine blades. Such protective schemes are most commonly used to improve erosion resistance in susceptible locations. LP blade coatings and surface treatments are used for improved performance against environmentally assisted mechanisms and liquid droplet erosion. MECHANICAL EFFICIENCY The mechanical efficiency of a turbine is the ratio of the brake output to the internal output. The mechanical efficiency is an index of the external losses. TURBINE EFFICIENCY The engine efficiency is the ratio of the real output of the turbine to the ideal output. The engine efficiency is primarily of interest to the designer as a means of comparing the real turbine with the ideal. Just as in the case of thermal efficiencies, the output basis upon which the engine efficiency is determined must be designated.

ADVANTAGES AND DISADVANTAGES OF STEAM TURBINES A steam drive offers several advantages as a prime mover for the operation of a centrifugal compressor. One big advantage is that the process involves two-phase fluid, and the working fluid being pumped from low pressure to high pressure is liquid. The work required to bring the water to the desired pressure (Point B) is much less than the work required to raise the pressure of gas by compressing it. In fact, the work required to pump the water is frequently neglected when determining the output and efficiency of large steam turbines. A second advantage of steam drives is its ability to burn a wide variety of fuels. Since the working fluid does not come in direct contact with the combustion products, any fuel ranging from natural gas to heavy residual fuel and solid fuels, such as coal or refuse, can be burned in the boiler.

| 392 |

| Steam Turbines | A third advantage, especially for small steam turbines, is to use steam, which is usually obtained as a by-product of the process. This increases the overall efficiency of the plant. The disadvantage of the steam turbine is the large amount of equipment required, since the water is heated indirectly in the furnace and the heat exchangers must be very large to raise the water to the desired temperature. The boiler requires a large supporting structure and foundation. In order to obtain high efficiencies with a Rankine cycle, the steam must be condensed at the turbine exhaust. The steam condenser requires a large surface area to cool the exhaust steam. The condenser also requires cooling water with its accompanying cooling towers. The medium and small steam turbine drives are usually back pressure turbines. While the use of a boiler offers the advantage of multiple fuel capabilities, the use of some of these fuels presents problems. If heavy fuels are burned, the liquids must be heated prior to introduction into the burners. The use of solid fuels creates even more problems, and the handling equipment can be a sizeable portion of the overall plant equipment. Fuels such as Bunker C oil and coal cause corrosion and fouling problems in the heat exchanger due to fuel impurities.

REFERENCES Cotton, K. C. Evaluating and Improving Steam Turbine Performance. Cotton Fact, Inc. Rexford, NY. 1993. Craig, H. R. M., and Kalderon D. “Research and Development for Large Steam Turbines.” Proc. American Power Conference. 1973. Craig, H. R. M., and Hobson G. “The Development of Long Last-Stage Turbine Blades.” GEC J. of Science and Technology, Vol. 40, No. 2. 1973. pp. 65-71. Leyzerovich, A. Large Power Steam Turbines, Volume 1: Design and Operation, Volume 2: Operations. PennWell Books. Tulsa OK. 1997. McCloskey, T. H., et al. Turbine Steam Path Damage: Theory & Practice Volume 1 Turbine Fundamentals. EPRI. 1999 McCloskey, T. H. et al., Turbine Steam Path Damage: Theory & Practice Volume 2 Damage Mechanisms. EPRI. 1999 Petrovic, M., and Riess, W. “Off-Design Flow Analysis of LP Steam Turbines”. 2nd. Conference on Turbomachinery-Fluid Dynamics and Thermodynamics. Amsterdam. 1997.

| 393 |

| Centrifugal Compressors | Petrovic, M., and Riess, W. “Off-Design Flow Analysis and Performance Prediction of Axial Turbines.” ASME Paper 97-GT-55. 1997. Sanders, W. P., “Turbine Steam Path Engineering for Operations and Maintenance Staff.” Turbo-Technic Services Incorporated. Toronto, Ontario, Canada. December, 1998. Trumpler, W .E. and Owens, H. M. “Turbine Blade Vibration and Strength.” Trans. ASME, Volume 77, 1955, pp. 337-341.

| 394 |

10 Electric Motors

MOTORS Electric motors have not always been used as the first choice for a compressor. This concept changed in the 1980s as the cost of fuel increased. Electric motors have high efficiency and reliability and are available to drive almost any configuration of centrifugal compressor. Many plants have moved from small steam turbine drives to electric motor drives. Since the mid-1990s large electric motors have become popular. In addition to the considerations of heat balance and investment costs, a third factor, the availability of feedstock, must be considered. Hydrocarbons are needed for many processes thus the burning of the hydrocarbons could be unproductive. Electric power generated from coal, hydro, or nuclear energy can be purchased with a resultant saving of feedstock to make more products. All electric motors operate on the same basic principal, regardless of type or size. When a wire carries electric current in the presence of a magnetic field (at least partially perpendicular to the current), a force on the wire is produced perpendicular to both the current and the magnetic field. In a motor, the magnetic field radiates either in toward or outward from the motor axis (shaft) across the air gap, which is the annular space between the rotor and stator. Current-carrying conductors parallel to the axis (shaft) then have a force on them tangent to the rotor circumference. The force on the wire opposes an equal force (or reaction) on the

| 395 |

| Centrifugal Compressors | magnetic field. It makes no difference whether the magnetic field is created in the rotor or the stator; the net result is the same—the shaft rotates. Within these basic principles there are many types of electric motors. Each has its own individual operating characteristics suited for specific drive applications. When several types are suitable, selection is based on initial installed cost and operating costs, including maintenance and consideration of reliability. Figure 10-1 shows speed-horsepower application zones for motors.

Figure 10-1: A range of speed and horsepower for various types of motors.

| 396 |

| Electric Motors | MOTOR SPEED AND TYPE MOTOR SPEEDS Centrifugal compressors of higher speeds are being used more frequently than ever before. The top motor speed available is 3600 rpm on a 60-hertz system. To meet requirement speeds higher than 3600 rpm, it is necessary to use step-up gears. The introduction of a step-up gear permits a normal induction motor speed of 1800 rpm. For 500 to 20,000-hp drives, 1800 rpm is the lowest first-cost induction motor. Similarly, the lowest first cost synchronous motor speed is 1200 rpm in sizes from 5000 hp and up. There is really no size limit when building a motor. Synchronous speeds are calculated by the same relationship given for induction motors. Speeds above the limits given are obtained through step-up gears. Large high-speed centrifugal compressors are often driven through step-up gears to obtain the desired compressor speed. Two-pole (3600 r/min at 60 Hz) synchronous motors can be built but are uneconomical in comparison with geared drives. MOTOR TYPES There are three types of motors available for compressor drives: •

Induction

•

Synchronous

•

Wound Rotor Induction

A LT E R N AT I N G - C U R R E N T ( S Q U I R R E L - C A G E ) INDUCTION MOTORS These motors are by far the most common constant-speed drives. They are relatively simple in design and, therefore, both low in cost and highly reliable. Squirrel-cage induction motors are simple, rugged, and reliable and have no rotating windings, slip rings, or commutators generally associated with other motors. They are utilized with all compressors and may be belted, geared, direct coupled, or flange mounted. In the latter, the motor stator is fastened to the compressor frame and the rotor is

| 397 |

| Centrifugal Compressors | mounted on the compressor shaft. The impeller of small dynamic compressors is often mounted directly on the motor shaft. The induction motor is the first choice for drives from 500 to 5000 hp because it has one insulated stator winding and one uninsulated, shorted rotor winding. For 1200-, 1800-, and 3600-rpm compressors no speed increaser is required. The 1800-rpm motor is the least expensive of the three speeds; therefore, this speed is usually selected for higher speeds using step-up gears, as shown in Figure 10-2. The induction motor has reasonable efficiency, but its lagging power factor is often a disadvantage, particularly in large sizes. Its current inrush is apt to be quite high. They vary in price from about $70/kW for medium size motors (10kW–200kW) to about $35/kW for larger motors. The power output of a motor is given by the following relationship:

Figure 10-2: Squirrel cage induction motor drive through a high speed gear driving a centrifugal compressor (courtesy MAN turbomaschinen GHH BORSIG).

| 398 |

| Electric Motors | Three Phase: P( kW ) = 0.00173 xVxIxPFxη

(10-1)

Single Phase: P( kW ) = 0.001xVxIxPFxη

(10-2)

where: V = Voltage I = line current (Amps) PF = Power factor η= Motor Efficiency The allowable power factor in any given installation is a function of the reactive components (leakage paths) to the real components The typical three-phase squirrel-cage motor has stator windings, which are connected to the power source. The rotor is a cylindrical magnet structure mounted on the shaft with slots in the surface, parallel (or slightly skewed) to the shaft; either bars are inserted into these slots or molten metal is cast in place and connected by a short-circuiting end ring at both ends of the rotor. The name “squirrel-cage” derives from this rotor-bar construction. In operation, current passing through the stator winding creates a rotating magnetic field, which cuts the rotor winding unless the rotor is turning in exact synchronism with the stator field. This cutting action induces a voltage, and hence a current, in the rotor which in turn reacts with the magnetic field to produce torque. The typical medium-sized squirrel-cage motor is designed to operate at 2 to 3% slip (97-98% of synchronous speed). The synchronous speed is determined by the power-system frequency and the stator-winding configuration. If the stator is wound to produce one north and one south magnetic pole, it is a two-pole motor; there is always an even number of poles (2, 4, 6, 8, etc.). The synchronous speed (n) is n = 120 f / p

(10-3)

where f = frequency, Hz p = number of poles

| 399 |

| Centrifugal Compressors | The actual operating speed will be slightly less by the amount of slip. Slip depends upon motor size and application. Typically, the larger motor, the less slip; an ordinary 7460-W (10 hp) motor may have 21⁄2% slip, whereas motors over 746 kW (1000 hp) may have less than 1⁄2%. High-slip motors (as much as 13% slip) are used for applications with high inertia that require a high starting torque. S Y N C H R O N O U S A LT E R N AT I N G - C U R R E N T M O T O R S The synchronous motor is a fixed-speed machine. Some consider it less reliable than the squirrel cage induction motor because it has a wound rotor to which direct current excitation must be applied, usually through rotating slip rings. Consequently, a small direct current supply (motorgenerator set or rectifier) is required, together with more elaborate starting equipment. Synchronous motors may be flange-mounted, coupled, geared, or mounted (engine type) on the compressor shaft. The synchronous motor on the compressor drive usually has a relatively low current inrush. It frequently is selected for its high or leading power factor. Synchronous motors, because of their more complicated design and the necessity for a field power supply, are typically applied only in largehorsepower ratings (several hundred horsepower and larger); synchronous motors over 60,000 kW (80,000 hp) have been built. Figure 10-3 is a three-section centrifugal compressor driven by a large synchronous motor (15,000 HP). With their latitude in size and characteristics and their important inherent high power factor and efficiency, synchronous motors are applied to a wide variety of drives. Synchronous motors present centrifugal compressor manufacturers with a problem unique among the various drivers available for their machinery. Synchronous machines produce an oscillating torque at twice slip frequency on starting. In case of inadequately designed shaft system of coupled compressor or if tile motor malfunctions, this torsional excitation has destructive effect. In order to avoid failure, it is important to carry out two things. First, a rigorous analysis has to be made at the design stage of the torsional effects during the start up. A mathematical model has been developed which represents the electrical power feed network for the motor, an equivalent circuit description of the motor itself, and a non-linear characterization of the shaft and coupling connecting the compressors to the motors. The next step is carried out during field installation. Instantaneous magnitude of the excitation torque and its resultant shaft stress are two

| 400 |

| Electric Motors |

Figure 10-3: Synchronous motor driving a three-section centrifugal compressor (courtesy Enterprise Products LLP).

important parameters to be measured during the field installation. For this, a torsional acceleration monitor attached to the uncompleted synchronous motor is used to gauge the axial torques being applied to the compressor on startup. Shaft stress levels can be measured using a strain measurement system based on telometry. The two instrumentation packages are used to insure safe operation, detect motor starting malfunctions, and verify the analytical design of the shaft/coupling system. These motors run in exact clock synchronism with the power system. For most modern power systems, these are truly constant-speed motors. The use of synchronous motors requires a detail calculation of the torsional dynamics on the rotor, since the high acceleration causes some very high torsional torques. In the conventional synchronous motor, a rotating magnetic field is developed by the stator currents as in induction motors. The rotor, however, is different, consisting typically of pairs of electromagnets (poles) spaced around the rotor periphery. The rotor field corresponds to the field produced by the AC stator having the same number of poles. The rotor or field coils are supplied with direct current; the magnetic field is therefore stationary with respect to the rotor structure. Torque is developed by the interaction of the rotor magnetic field and the stator current (in-phase component). Under

| 401 |

| Centrifugal Compressors | no-load conditions and with appropriate DC field current, rotor and stator magnetic-field centers coincide. The voltage applied to the stator winding is balanced by an opposing voltage generated in the stator by the rotor field (induced) and no AC power current flows. As load is applied, the rotor tends to decelerate momentarily, causing a shift of rotor position with respect to the AC field. This shift produces a difference between the applied and induced voltage; the voltage difference causes current to flow; the current reacts with the rotor magnetic flux, producing torque. Synchronous motors should not be started with the DC field applied. Instead they are started as inductions motors; bars, acting like a squirrelcage rotor, are embedded in the field-pole surface and connected by end rings at both ends of the rotor. These damper bars also serve to damp out oscillations under normal running conditions. When the motor is at approximately 95 percent speed (depending upon application and motor design), direct current is applied to the field and the motor pulls into step (synchronism). Because the damper bars do not affect the synchronous-speed characteristics, they are designed for starting performance. This provides flexibility in the accelerating characteristics to meet specific application requirements without affecting running efficiency and other synchronousspeed characteristics. The rotor design of a squirrel-cage motor, on the other hand, must be a compromise between starting and running performance. The DC field is usually shorted by a resistor during starting and contributes accelerating torque, particularly near synchronous speed. The synchronous motor is usually the primary choice for very large ratings because of price when power factor and efficiency are prime considerations. These motors have three windings on the rotor—one insulated and one un-insulated amortisseur winding that is used for starting only. The insulated synchronous motor rotor winding requires a DC power supply that is usually supplied from a brushless exciter mounted on the motor shaft. Synchronous motors are available at the same speeds as induction motors; however, in practice only the 1200-rpm speed is used. Step-up gears are always used to match the compressor speed requirements. The control and operation of synchronous motors require more devices, but these are well understood and are very reliable. The system is such that the operator controls the machine in the same way he controls an induction motor. The synchronous drive is more efficient; it can operate at 1.0 pf or 0.8 pf and is lower in first cost in the larger horsepower sizes.

| 402 |

| Electric Motors | For very large drives above 30,000 hp, the use of non-self-starting synchronous motors at 3600 rpm is possible. These are synchronous generators that must be brought up to speed with a starting turbine or motor and synchronized with the system and operated as a motor. Obviously, it is a requirement on these drives that the compressor be unloaded to 15 to 20 percent of full load to keep the starting equipment size reasonable. WOUND-ROTOR INDUCTION Wound-rotor induction motors have not been used for compressor drives because of their cost and high losses when resistors are used for speed control. One exception is the very large drives used in wind tunnels where there is no other way to control the compressor output except by speed control. The increased cost of energy and the availability of thyristors in large power sizes is causing a new interest in the wound-rotor motor because, instead of discarding the slip power as heat at lower than synchronous speed, this power is “recycled” by solid-state rectifier inverter techniques and put back into the power system. On applications where speed control is required (and particularly where the load cycle requires operation at reduced speed for long periods), this drive results in relatively high efficiency over the entire speed range.

Cost Recent changes in motor prices have altered the specific relationships of synchronous versus induction motor prices, but a general relationship continues to exist as shown in Table 10-1 Geared Drives

5,000–15,000 HP 6,000–50,000

Lowest First Cost Induction 1800 RPM X

Synchronous 1200 RPM X

Table 10-1: Synchronous versus induction motor prices.

The overlap between 6000 hp and 15,000 hp is intentional because specific drive characteristics and requirements will change the choice.

| 403 |

| Centrifugal Compressors | POWER-FACTOR CORRECTION Conventional synchronous-motor power factors are either 100 or 80% leading. Leading-power factor machines are used frequently to correct for the lagging power factor of the remaining plant load. Even 100% powerfactor motors can be operated leading at reduced loads. An advantage of synchronous motors over capacitors is their inherent tendency to regulate power-system voltage; as voltage drops, more leading reactive power is delivered to the power system, and conversely as voltage rises, less reactive power is delivered, in contrast to capacitors for which the reactive power decreases directly in proportion to the voltage drop squared. The amount of leading reactive power delivered to the system depends on dc field current, which is readily adjustable. Field current is an important control element. It controls not only the power factor but also the pullout torque (the load at which the motor pulls out of synchronism). For example, field forcing can prevent pullout on anticipated high transient loads or voltage dips. Loads with known high transient torques are driven frequently with 80% power-factor synchronous motors. The needed additional field supplies both additional pullout torque and power-factor correction for the power system. When high pullout torque is required, the leading power-factor machine is often less expensive than a unity-power-factor motor with the same torque capability. CORRECTION MOTOR SIZE SELECTION Nothing is more useless than a motor installed on its foundation that is too small to start the machine to which it is coupled. There is no excuse for this problem because the widespread use of computers by machine designers makes it practical to predict accurately the brake horsepower requirements of the compressor. It is easy to avoid this problem because motors are available in standard ratings in the larger horsepower sizes in average increments of about 15%. Therefore, a practice of selecting the standard motor rating that matches or exceeds the brake horsepower of the driven machine will result in 0 to 15% margin. The practice of adding 10% to the calculated brake horsepower and then using the next larger standard horsepower motor will result in margins of 10 to 25%. With this wide selection of standard ratings, it is easy to find a good match between the driven machine size and the motor rating without the use of continuous overload capabilities.

| 404 |

| Electric Motors | Past practice has been to use motors with an overload capability of 115%. This can reduce investment cost if the machine is to be operated at 115% load. The motor life, however, will be shortened because the temperature rise at 115% load is 90°C. Conversely, the life of the same motor will be increased if this motor is operated at 100% load or approximately 70°C rise. CORRECT MOTOR TORQUE SELECTION Torque selection would be easier if a simple match between full-load brake horsepower and the motor rating were the whole story. To get the benefits of the motor’s simplicity, the plant designer must understand its peculiarities. When a motor is starting, it draws four to seven times its fullload current. This depresses the voltage, particularly on large motors, and results in a reduction of motor starting torque by the square of the voltage drop. If the voltage drops to 80%, the torque drops to 64% [(80%)2 = 64%]. This will be no problem if the designer has planned for this condition and matched the motor speed torque curve at reduced voltage to that required by the compressor.

Figure 10-4: A typical motor vs. driven compressor speed–torque characteristic curve.

| 405 |

| Centrifugal Compressors | It is good practice to apply a motor with torque capability at the voltage, to which the system drops when the motor is being started, that is at least 10% greater than the torque required by the driven machine at all points between zero and full speed, as seen in Figure 10-4. A motor with special torque characteristics, a stiffer power system, or a method of unloading the driven machine may be required. On some very large motors, all three of these may be required to ensure a drive that will operate. M O T O R V O LTA G E A N D S TA R T I N G M E T H O D Selection of the proper motor voltage is usually determined by the available electrical system and its size. The most frequently used voltage for 500-4000 hp motors is 4160 volts. On large electrical systems, motors 4000 hp and larger may have to be rated 6600 or 13,200 volts because of available switchgear. When motors are started, they draw high inrush current that depresses the line voltage. The amount of voltage drop depends upon the size of the electrical system. Standards usually indicate this voltage dip should be less than 10%; however, there are many good systems where the motor driving a compressor is the largest in the plant and where the voltage dip is 20 to 30% or more. As pointed out previously, motor torque varies as the square of the applied voltage. This reduced available torque usually requires that motordriven compressors be started with the inlet valves or guide vanes closed. Most compressors will operate under these conditions for a short time (usually 60 to 120 seconds) without damage from heating or surge until the motor reaches full speed. In many cases, the compressor breakaway torque is 15 to 20%, and as the compressor comes up to speed, it evacuates itself until the torque at full speed is 15 to 20%. These typical curves are shown in Figure 10-5. With low-starting torque requirements, it is possible on large ratings to design special low-inrush low-torque motors that will cause less voltage drop when starting. The three commonly used methods of reduced-voltage starting in use at present are: reactor start, autotransformer start, and unit transformer start. The most popular of the three is the unit (or captive) transformer start because it does not require extra switching, can be designed to meet individual system requirements simply by selecting the transformer impedance, and can permit the use of medium or small motors on 13,800

| 406 |

| Electric Motors | volt systems without penalizing motor characteristics. Reduced voltage starting is never a drive requirement, but it is a requirement to prevent or limit the electrical system disturbance.

Figure 10-5: Typical compressor speed–torque curve.

INERTIA LOAD (WK2) OF THE COMPRESSOR The energy required to accelerate the inertia of the driven machine shows up as heat in the motor rotor as the machine comes up to speed in exactly the same way that the kinetic energy of a moving automobile is converted to heat in the brake linings when the automobile is stopped. Frequent stops or too heavily loaded stops of an automobile will cause the brake linings to overheat and fail. In a similar fashion, starting the motor too frequently or starting a driven machine with load inertia larger that the motor was designed for will cause the motor rotor winding to overheat and fail. Here the analogy stops for, if the motor is properly matched to the load inertia, the rotor winding will last indefinitely. Industry standards have been established for the load WK2 capability of the standard motor.

| 407 |

| Centrifugal Compressors | These values are high enough for most centrifugal compressor drives. Many high-speed compressors require step-up gears. The WK2 at the motor shaft is proportional to the square of the gear ratio. For this reason, special motor characteristics may be required. MOTOR BURNOUTS Many motor windings are lost every year because of sudden faults, such as foreign objects entering the motor and puncturing the winding insulation. The overwhelming majority of winding failures, however, are caused by burnouts resulting from sustained over current. Winding failure caused by over current is not a “sudden death” proposition; insulation does not immediately break down upon arrival at a critical temperature. Insulation failure results from over temperature for a period of time. A rule of thumb commonly applied to motor insulation life is that, for every 18ºF (10ºC) rise above rated temperature, winding life will be reduced by 1⁄2. Consider the example of a 3-phase motor with Class B insulation, service factor of 1.0, and rated temperature rise of 144ºF (80ºC) (by resistance), temperature rise will reasonably conform to the formula: Percent Motor Load  Temp. Increase = Rated Rise   100 

(10-4)

Modifying the momentum equation by substituting the skin friction, the following relationship is obtained for the meridional plane: Assume that this motor is designed for a useful life of 100,000 hours of operation at rated load design ambient temperature. Suppose that, upon installation, the motor is operated at 120% of rated load. Temperature rise above ambient will be: 2

 120  Temp Increase = 80º C   = 115º C  100 

(10-5)

This is approximately 35ºC above rated rise. To determine life expectancy of the motor in operation at 120% of rated load, this formula can be applied:  1  Life = Rated Life N  2 

| 408 |

(10-6)

| Electric Motors | where: N=

Actual º C Rise − Rated º C Rise 10º C

(10-7)

for this example: 115°C - 80°C N = --------------------- = 3.5 10°C

(10-8)

Therefore the motor life expectancy is:  1  Life Expectancy = 100000  3.5  = 9,300 hours 2 

(10-9)

A study of motor winding failures conducted by Underwriters’ Laboratories found that the motor burnouts fell within seven categories: 1.

Aged insulation was given as reason for insulation failure in 11.6% of the failures. In these cases, the motor had outlived its design life span.

2.

Bearing failure resulting in locked rotor or excessive drag caused 19.6% of the failures.

3.

One or two phases were burned out in 23.7% of the failures. This resulted from phase voltage unbalance or a single-phasing condition.

4.

Dirt or other foreign material causing excessive friction and reduced ventilation accounted for 9.5% of the failures.

5

Insulation attack by oil or moisture caused 12.9% of the failures.

6.

Extreme heat resulted in roasted windings on 13.3% of the failed motors.

7.

A variety of other reasons accounted for the remaining 8.4% of the burnouts.

Only those motors whose insulation failed because of old age can be considered to have died of natural causes. Failure of the remaining 88.4% of the motors might have been forestalled by proper motor application and proper motor operation.

| 409 |

| Centrifugal Compressors | APPLICATION OF INDUCTION MOTORS One of the most important considerations in applying induction motors is the way a motor starts and behaves while starting its load and bringing it up to speed. Lack of understanding this behavior has lead to many motor failures. The characteristic of the motor that distinguishes it from turbine drivers is that it instantly develops large starting torques as soon as the motor is energized. No driven equipment, including the motor itself, is capable of coming up to speed as fast as the torque is applied. The results are thermal strain in the motor and mechanical strain on the shafts, couplings, and load. Under accelerating conditions, the most critical parts of a motor are the rotor bars and end rings. These parts heat up most rapidly. During acceleration, the motor current (ampere load) is 5–6 times the normal full load value and remains at this high value until the motor is at approximately 80–90% speed. This entire power input is converted into heating the motor stator and rotor and to shaft horsepower output. Since heating is a function of the motor current squared, heating of the motor is 25–30 times normal. No attempt can be made in calculating the actual heat rise in the motor without having the complete motor design data and the load torque requirements. As a general rule, the motor rotor will reach maximum allowable temperature in 10–20 seconds and the stator windings in slightly longer periods of time. The rotors of most 480 volt motors are of cast construction and are not as subject to the effects of overheating as the brazed bar and ring constructed rotors of 2400 volt motors. Because of the complexity of calculating safe temperature rise related to accelerating time, NEMA standards have specified the number of safe starts of 250-500 Hp motors as follows: N U M B E R O F S TA R T S Squirrel-cage induction motors rated 250 to 500 horsepower inclusive as given in par. D of MG 1-10.32 shall be capable of making two starts in succession, coasting to rest between starts, with the motor initially at ambient temperature or of making one start with the motor initially at ambient temperature not exceeding its rated load operating temperature, provided the WK2 of the load, the load torque during acceleration, the applied voltage, and the method of starting are those for which the motor was designed.

| 410 |

| Electric Motors | If additional starts are required, it is recommended that none be made until all conditions affecting operation have been thoroughly investigated and the apparatus examined for evidence of excessive heating. It should be recognized that the number of starts should be kept to a minimum since the life of the motor is affected by the number of starts. Motors larger than 500 hp generally have longer accelerating times and consequently the number of safe starts is less than that of smaller motors. As a guide, the following rules should be followed: •

Not more than 3 starts/hr for motors 200 hp or less

•

Not more than 2 starts/hr for motor 200-500 hp

•

Not more than 1 start/hr for motors 500 hp and larger

Following these simple rules will do much to prolong motor life and reduce maintenance costs. Figure 10-6 provides a nomogram, which gives locked rotor current for standard induction motors. All vertical scales may be multiplied by 10 or 100. The determination of locked-rotor current is referred to by NEMA code.

Figure 10-6: Nomogram for locked motor currents for induction motors.

| 411 |

| Centrifugal Compressors | MOTOR ENCLOSURES Motor winding life can be substantially increased by the proper choice of enclosures: •

Open drip-proof, for use indoors. Relatively clean locations.

•

Weather-protected for use in outdoor locations. This enclosure can easily be modified to include filters when used in extremely dirty locations.

•

Totally Enclosed Water Air Cooled (TEWAC)—For use in outdoor locations. Since the heat from the motor losses is taken away by the water, motors having this enclosure are well suited for installation in small rooms.

•

Totally Enclosed Fan Cooled (TEFC)—These machines are arranged so that the heat is transferred from the motor to the ambient by utilizing an air-to-air heat exchanger either mounted on the frame of the motor or built as a part of the motor with fins cast into the end shields or frames.

•

Hazardous Locations—This requirement must be met by pressurizing the motor enclosure with clean air or inert gas. Explosion proof motors are not generally above 500 hp.

REFERENCES Daugherty R. H. “Chapter 29 Electric Motors and Auxiliaries.” Perry’s Chemical Engineers’ Handbook 7th Edition. 1997. Hargett, Y. S. “Large Steam Turbine Driven Generators.” Large Steam Turbine Generator Department–Schenectady, N.Y. Nippes P. I. Synchronous Machinery, The Electric Power Engineering Handbook. CRC Press LLC. 2000. Wright, J. “A Practical Solution to Transient Torsional Vibration in Synchronous Motor Drive Systems.” Amer. Soc. of Mech. Eng., Pub. 75DE-15.

| 412 |

11

Rotor Dynamics, Bearings, Lubrication Couplings, and Gears

ROTOR DYNAMICS The present trend in rotating equipment is toward increasing design speeds, which increases operational problems from vibration—hence, the importance of vibration analysis. A thorough appreciation of vibration analysis will aid in the diagnosis of rotor dynamics problems. D E S I G N C O N S I D E R AT I O N S Design of rotating equipment for high-speed operation requires careful analysis. The discussion in the preceding section presents elementary analysis of such problems. Once a design is identified as having a problem, it is an altogether different matter to change this design to cure the problem. The following paragraphs discuss some observations and guidelines based on the analysis presented in the previous sections. N AT U R A L F R E Q U E N C Y This parameter for a single degree of freedom is given by ϖn =

k m

(11-1)

| 413 |

| Centrifugal Compressors | Increasing the mass reduces ϖn , and increasing the spring constant k increases it. From a study of the damped system, the damped natural frequency, as shown below is lower than ϖn: ϖd = ϖn

(1 − ζ ) 2

(11-2)

UNBALANCES All rotating machinery is assumed to have an unbalance. Unbalance produces excitation at the rotational speed. The natural frequency of the system ϖn is also known as the critical shaft speed. From the study of the forced-damped system, the following conclusions can be drawn: •

the amplitude ratio reaches its maximum values at ϖd = ϖn

•

(1 − ζ ) 2

the damped natural frequency ϖn does not enter the analysis of the forced-damped system.

The more important parameter is ϖn, the natural frequency of the undamped system. In the absence of damping, the amplitude ratio becomes infinite at ϖ = ϖn. For this reason, the critical speed of a rotating machine should be kept away from its operating speed. Small machinery involves small values of mass m and has large values of the spring constant k (bearing stiffness). This design permits a class of machines, which are small in size and of low speed in operation, to operate in a range below their critical speeds. This range is known as subcritical operation, and it is highly desirable if it can be attained economically. The design of large rotating machinery centrifugal compressors, gas and steam turbines, and large electrical generators pose a different problem. The mass of the rotor is usually large, and there is a practical upper limit to the shaft size that can be used. Also, these machines operate at high speeds. This situation is resolved by designing a system with a very low critical speed in which the machine is operated above the critical speed. This is known as supercritical operation. The main problem is that during

| 414 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | startup and shutdown, the machine must pass through its critical speed. To avoid dangerously large amplitudes during these passes, adequate damping must be located in the bearings and foundations. The natural structural frequencies of most large systems are also in the low-frequency range, and care must be exercised to avoid resonant couplings between the structure and the foundation. The excitation in rotating machinery comes from rotating unbalanced masses. These unbalances result from four factors: 1.

An uneven distribution of mass about the geometric axis of the system. This distribution causes the center of mass to be different from the center of rotation.

2.

A deflection of the shaft due to the weight of the rotor, causing further distance between the center of mass and the center of rotation. Additional discrepancies can occur if the shaft has a bend or a bow in it.

3.

Static eccentricities are amplified due to rotation of the shaft about its geometric center.

4.

If supported by journal bearings, the shaft may describe an orbit so that the axis of rotation itself rotates about the geometric center of the bearings.

These unbalance forces increase as a function of the square of the speed (ω2), making the design and operation of high-speed machinery a complex and exacting task. Balancing is the only method available to tame these excitation forces. The high-speed turbomachines must operate in a region away from any critical speed. The amplification factor used to indicate the severity of the critical speed is given by the relationship AF =

Critical Speed Peak Width of the “half − Power” point

(11-3)

AF =

N C1 N 2 − N1

(11-4)

where (N2 – N1) are the RPMs corresponding to the .707 peak critical amplitude.

| 415 |

| Centrifugal Compressors | The amplification factor should be below 8 and preferably below 5. A rotor response plot is shown in Figure 11-1. The operational speed for units operating below the first critical speed should be at least 20% below the critical speed. For units operating above their first critical speed, the operational speed must be at least 15% above the critical speed and/or 20% below any critical speed. The preferred bearings for the various types of installation are tilting-shoe radial bearings and the self-equalizing tilting pad/thrust bearings. Radial and thrust bearings should be equipped with embedded temperature sensors to detect pad surface temperatures.

Figure 11-1: Rotor response plot. (Figure 7 of Standard 617, Centrifugal Compressors for General Refinery Services, 4th Edition, 1979, reprinted by courtesy of the America Petroleum Institute).

| 416 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | APPLICATION TO ROTATING MACHINES RIGID SUPPORTS The simplest model of a rotating machine consists of a large disc mounted on a flexible shaft with the ends mounted in rigid supports. The rigid supports constrain a rotating machine from any lateral movement but allow free angular movement. A flexible shaft operates above its first critical. Figures 11-2 and 11-3 show such a shaft. The mass center of the disc “e” is displaced from the shaft centerline or geometric center of the disc due to manufacturing and material imperfections. When this disc is rotated at a rotational velocity ω, the mass causes it to be displaced so that the center of the disc describes an orbit of radius δr from the center of the bearing centerline. If the shaft flexibility is represented by the radial stiffness (Kr), it will create a restoring force on the disc of Kr δr that will balance the centrifugal force equal to ϖ2 (δr + e). Equating the two forces obtains K rδ = mϖ 2 (δ r + e )

(11-5)

Therefore, mϖ 2 e (ϖ / ϖ n ) e = K r − mϖ 2 1 − (ϖ / ϖ n ) 2

δr =

(11-6)

where ϖn = Kr /m, the natural frequency of the lateral vibration of the shaft and disc at zero speed. The previous equation shows that when ϖ < ϖn, δr is positive. Thus, when operating below the critical speed, the system rotates with the center of mass on the outside of the geometric center. Operating above the critical speed (ϖ > ϖn), the shaft deflection δr tends to infinity. Actually, this vibration is damped by outside forces. For very high speeds (ϖ >> ωn), the amplitude δr equals e, meaning that the disc rotates about its center of gravity.

| 417 |

| Centrifugal Compressors |

Figure 11-2: Rigid supports.

Figure 11-2: Flexible supports.

| 418 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | FLEXIBLE SUPPORTS The previous section discussed the flexible shaft with rigid bearings. In the real world, the bearings are not rigid but possess some flexibility. If the flexibility of the system is given by Kb, then each support has a stiffness of Kb/2. In such a system the flexibility of the entire lateral system can be calculated by the following relationship: 1 1 1 K +Kr = + = b Kt Kr Kb Kr K b Kt =

Kr Kb Kb + K r

(11-7) (11-8)

Therefore, the natural frequency Kr Kb Kb + Kr m

ϖ nt =

Kt = m

ϖ nt =

Kr Kb x m Kb + K r

ϖ nt = ϖ n

Kb Kb + K

(11-9)

r

It can be observed from the previous expression that when Kb « Kr (very rigid support), then ϖnt = ϖn or the natural frequency of the rigid system. For a system with a finite stiffness at the supports, or Kb < Kr, then ϖn is less than ϖnt . Hence, flexibility causes the natural frequency of the system to be lowered. Plotting the natural frequency as a function of bearing stiffness on a log scale provides a graph, as shown in Figure 11-4. When Kb « Kr, then ϖnt = ϖn Kb /K. Therefore, ϖnt is proportional to the square root of Kb, or log ϖnt is proportional to one-half log Kb. Thus, this relationship is shown by a straight line with a slope of 0.5 in Figure 114. When Kb « Kr, the total effective natural frequency is equal to the natural rigid-body frequency. The actual curve lies below these two straight lines, as shown in Figure 11-4. The critical speed map shown in Figure 11-4 can be extended to include the second, third, and higher critical speeds. Such an extended

| 419 |

| Centrifugal Compressors | critical speed map can be very useful in determining the dynamic region in which a given system is operating. One can obtain the locations of a system’s critical speeds by superimposing the actual support versus the speed curve on the critical speed map. The intersection points of the two sets of curves define the locations of the system’s critical speeds. When the previously described intersections lie along the straight line on the critical speed map with a slope of 0.5, the critical speed is bearing controlled. This condition is often referred to as a “rigid-body critical.” When the intersection points lie below the 0.5 slope line, the system is said to have a “bending critical speed.” It is important to identify these points, since they indicate the increasing importance of bending stiffness over support stiffness. Figures 11-5 and 11-6 show vibration modes of a uniform shaft supported at its ends by flexible supports. Figure 11-5 shows rigid supports and a flexible rotor. Figure 11-6 shows flexible supports and rigid rotors. To summarize the importance of the critical speed concept, one should bear in mind that it allows an identification of the operation region of the rotor-bearing system, probably mode shapes, and approximate locations of peak amplitudes.

Figure 11-4: Critical speed map.

| 420 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Figure 11-5: Rigid supports and a flexible rotor.

Figure 11-6: Flexible supports and rigid rotors.

FORCES ACTING ON A ROTOR BEARING SYSTEM There are many types of forces that act on a rotor-bearing system. The forces can be classified into three categories: (1) casing and foundation forces, (2) forces generated by rotor motion, and (3) forces applied to a rotor. Table 11-1 by is a compilation of these forces.

| 421 |

| Centrifugal Compressors | Source of Force

Description

Application

1. Forces transmitted to foundations, casing, or bearing pedestals

Constant, unidirectional force Constant force, rotational Variable, unidirectional

Constant liner acceleration Rotation in gravitational or magnetic field Impressed cyclic ground or foundation-motion Airblast, explosion, or earthquake. Nearby unbalanced machinery Blows, impact Present in all rotating machinery

Impulsive forces Random forces 2. Forces generated by rotor motion

Rotating unbalance; residual, or bent shaft Coriolis forces Elastic hysteresis of rotor

Coulomb friction

Fluid friction Hydrodynamic forces, static. Hydrodynamic forces, dynamic. Dissimilar elastic beam Stiffness reaction forces Gyroscopic moments 3. Applied to rotor

Drive torque Cyclic forces Oscillating torques Transient torques Heavy applied rotor force Gravity Magnetic field: stationary or rotating Axial forces

Table 11-1: Forces acting on rotor bearing systems.

| 422 |

Motion around curve of varying radius. Space applications Rotary-coordinated analyses Property of rotor material which appears when rotor is cyclically deformed in bending, torsionally or axially Construction damping arising from relative motion between shrinkfitted assemblies Dry-friction bearing whirl Viscous shear of bearings Fluid entrainment in turbomachinery Windage Bearing load capacity Voute pressure forces Bearing stiffness and damping properties Rotors with differing rotor lateral stiffness. Slotted rotors, electrical machinery, Keyway. Abrupt speed change conditions Significant in high-speed flexible rotors with discs Accelerating or constant-speed operation Internal combustion engine torque and force components Misaligned couplings, Propellers, Fans Internal combustion engine drive Gears with indexing or positioning errors Drive gear forces Misaligned 3-or-more rotor-bearing assembly Non-vertical machines Non-spatial applications Rotating electrical machinery Turbomachine balance piston Cyclic forces from propeller, or fan Self-excited bearing forces Pneumatic hammer

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | FORCES TRANSMITTED TO CASING AND F O U N D AT I O N S These forces can be due to foundation instability, other nearby unbalanced machinery, piping strains, rotation in gravitational or magnetic fields, or excitation of casing or foundation natural frequencies. These forces can be constant or variable with impulse loadings. The effect of these forces on the rotor-bearing system can be great. Piping strains can cause major misalignment problems and unwanted forces on the bearings. Operation of reciprocating machinery in the same area can cause foundation forces and unduly excite the rotor of a turbomachine. F O R C E S G E N E R AT E D B Y R O T O R M O T I O N These forces can be classified into two categories: (1) forces due to mechanical and material properties and (2) forces caused by various loadings of the system. The forces from mechanical and material properties are unbalanced and are caused by a lack of homogeneity in materials, rotor bow, and elastic hysterisis of the rotor. The forces caused by loadings of the system are viscous and hydrodynamic forces in the rotorbearing system and various blade-loading forces, which vary in the operational range of the unit. FORCES APPLIED TO A ROTOR Rotor-applied forces can be due to drive torques, couplings, gears, misalignment, and axial forces from piston and thrust unbalance. They can be destructive, and they often result in the total destruction of a machine. R O T O R B E A R I N G S Y S T E M I N S TA B I L I T I E S Instabilities in rotor-bearing systems may be the result of different forcing mechanisms. One can divide these instabilities into two general yet distinctly different categories: (1) the forced or resonant instability dependent on outside mechanisms in frequency of oscillations and (2) the self-excited instabilities that are independent of outside stimuli and independent of the frequency. Table 11-2 is the characterization of the two categories of vibration stimuli.

| 423 |

| Centrifugal Compressors |

Frequency / rpm relationship Amplitude/rpm relationship Influence of damping

System Geometry

Vibration frequency

Forced or Resonant Vibration

Self-Excited or Instability Vibration

NF = Nrpm or N or Rational fraction Peak in narrow bands of rpm Additional damping reduces amplitude No change in rpm at which it occurs Lack of axial sym. External forces

Constant and relatively independent of rotating speed Blossoming at onset and continue to increase with increasing rpm Additional damping may defer to a higher rpm. Will not materially affect amplitude

At or near shaft Critical or natural Frequency 1. Critical freq. above running speed 2. Axisymmetric 3. Damping

Independent of symmetry Small deflection to an axisymmetric system Amplitude will self-propagate Same 1. Operating rpm below onset 2. Eliminates instability 3. Introduce damping

Table 11-2: Characteristics of forced and self-excited vibration.

F O R C E D ( R E S O N A N T ) V I B R AT I O N In forced vibration the usual driving frequency in rotating machinery is the shaft speed or multiples of this speed. This speed becomes critical when the frequency of excitation is equal to one of the natural frequencies of the system. In forced vibration, the system is a function of the frequencies. These frequencies can also be multiples of rotor speed excited by frequencies other than the speed frequency such as blade passing frequencies, gear mesh frequencies, and other component frequencies. Figure 11-7 shows that for forced vibration, the critical frequency remains constant at any shaft speed. The critical speeds occur at one-half, one, and two times the rotor speed. The effect of damping in forced vibration reduces the amplitude, but it does not affect the frequency at which this phenomenon occurs. Typical forced vibration stimuli are as follows: •

| 424 |

Unbalance. This stimulus is caused by material imperfections, tolerances, etc. The mass center of gravity is different from the geometric case, leading to a centrifugal force acting on the system.

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | •

Asymmetric flexibility. The sag in a rotor shaft will cause a periodic excitation force twice every revolution.

•

Shaft misalignment. This stimulus occurs when the rotor centerline and the bearing support line are not true. Misalignment may also be caused by an external piece such as the driver to a centrifugal compressor. Flexible couplings and better alignment techniques are used to reduce the large reaction forces.

Figure 11-7: Characteristic of forced vibration or resonance in rotating machinery. (Ehrich, F. F., “Identification and Avoidance of Instabilities and Self-Excited Vibrations in Rotating Machinery,” Adopted from ASME Paper 72-DE-21, General Electric Co., Aircraft Engine Group, Group Engineering Division, May 11, 1972.).

| 425 |

| Centrifugal Compressors | PERIODIC LOADING This type of loading is caused by external forces that are applied to the rotor by gears, couplings, and fluid pressure, which is transmitted through the blade loading. S E L F - E X C I T E D I N S TA B I L I T I E S The self-excited instabilities are characterized by mechanisms, which whirl at their own critical frequency, independent of external stimuli. These types of self-excited vibrations can be destructive since they induce alternating stress, which leads to fatigue failures in rotating equipment. The whirling motion that characterizes this type of instability generates a tangential force normal to the radial deflection of the shaft and a magnitude proportional to that deflection. The type of instabilities that fall under this category are usually called whirling or whipping. At the rotational speed where such a force is started, it will overcome the external stabilizing damping force and induce a whirling motion of ever-increasing amplitude. Figure 11-8 shows the onset speed. The onset speed does not coincide with any particular rotation frequency. Also, damping results from a shift of this frequency, not in the lowering of the amplitude as in forced vibration. Important examples of such instabilities include hysteretic whirl, dry-friction whip, oil whip, aerodynamic whirl, and whirl due to fluid trapped in the rotor. In a self-excited system, friction or fluid energy dissipations generate the destabilizing force.

Hysteretic Whirl This type of whirl occurs in flexible rotors and results from shrink fits. When a radial deflection is imposed on a shaft, a neutral strain axis is induced normally to the direction of flexure. From first-order considerations, the neutral-stress axis is coincident with the neutral-strain axis, and neutral-stress axes are displaced so that the resultant force is not parallel to the deflection. The tangential force normal to the deflection causes whirl instability. As whirl begins, the centrifugal force increases, causing greater deflections—which result in greater stresses and still greater whirl forces. This type of increasing whirl motion may eventually be destructive, as seen in Figure 11-9. Some initial impulse unbalance is often required to start the whirl motion. Newkirk has suggested that the effect is caused by interfaces of joints in a rotor (shrink fits) rather than defects in rotor material. This type

| 426 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | of whirl phenomenon occurs only at rotational speeds above the first critical. The phenomenon may disappear and then reappear at a higher speed. Some success has been achieved in reducing this type of whirl by reducing the number of separate parts, restricting the shrink fits, and providing some lockup of assembled elements.

Figure 11-8: Characteristics of instabilities or self-excited vibration in rotating machinery. (Ehrich, F. F., “Identification and Avoidance of Instabilities and SelfExcited Vibrations in Rotating Machinery,” Adopted from ASME Paper 72-DE-21, General Electric Co., Aircraft Engine Group, Group Engineering Division, May 11, 1972.).

| 427 |

| Centrifugal Compressors |

Figure 11-9: Hysteretic whirl. (Ehrich, F. F., “Identification and Avoidance of Instabilities and Self-Excited Vibrations in Rotating Machinery,” Adopted from ASME Paper 72-DE-21, General Electric Co., Aircraft Engine Group, Group Engineering Division, May 11, 1972.).

Dry-Friction Whirl This type of whip is experienced when the surface of a rotating shaft comes into contact with an unlubricated stationary guide. The effect takes place because of an unlubricated journal, contact in radial clearance of labyrinth seals, and loss of clearance in hydrodynamic bearings. Figure 11-10 shows this phenomenon. When contact is made between the surface and the rotating shaft, the coulomb friction will induce a tangential force on the rotor. This friction force is approximately proportional to the radial component of the contact force, creating a condition for instability. The whirl direction is counter to the shaft direction.

| 428 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Figure 11-10: Dry friction whirl. (Ehrich, F. F., “Identification and Avoidance of Instabilities and Self-Excited Vibrations in Rotating Machinery,” Adopted from ASME Paper 72-DE-21, General Electric Co., Aircraft Engine Group, Group Engineering Division, May 11, 1972.).

Oil Whirl This instability begins when fluid entrained in the space between the shaft and bearing surfaces begins to circulate with an average velocity of one-half of the shaft surface speed. Figure 11-11 shows the mechanism of oil whirl. The pressures developed in the oil are not symmetric about the rotor. Because of viscous losses of the fluid circulating through the small clearance, higher pressure exists on the upstream side of the flow than on the downstream side. Again, a tangential force results. A whirl motion exists when the tangential force exceeds any inherent damping. It has been shown that the shafting must rotate at approximately twice the critical speed for whirl motion to occur. Thus, the ratio of frequency to

| 429 |

| Centrifugal Compressors | rpm is close to 0.5 for oil whirl. This phenomenon is not restricted to the bearing, but it also can occur in the seals. The most obvious way to prevent oil whirl is to restrict the maximum rotor speed to less than twice its critical. Sometimes oil whip can be reduced or eliminated by changing the viscosity of the oil or by controlling the oil temperature. Bearing designs that incorporate grooves or tilting pads are also effective in inhibiting oil-whirl instability.

Aerodynamic Whirl Although the mechanism is not clearly understood, it has been shown that aerodynamic components, such as compressor wheels and turbine

Figure 11-11: Oil friction whirl. (Ehrich, F. F., “Identification and Avoidance of Instabilities and Self-Excited Vibrations in Rotating Machinery,” Adopted from ASME Paper 72-DE-21, General Electric Co., Aircraft Engine Group, Group Engineering Division, May 11, 1972.).

| 430 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | wheels, can create cross-coupled forces due to the wheel motion. Figure 11-12 is one representation of how such forces may be induced. The acceleration or deceleration of the process fluid imparts a net tangential force on the blading. If the clearance between the wheel and housing varies circumferentially, a variation of the tangential forces on the blading may also be expected, resulting in a net destabilizing force, as shown in Figure 11-12. The resultant force from the cross coupling of angular motion and radial forces may destabilize the rotor and cause a whirl motion. The stiffness that results from the previous quantification may be used in a critical-speed program in much the same manner as bearing coefficients.

Figure 11-12: Aerodynamic cross coupling. (Ehrich, F. F., “Identification and Avoidance of Instabilities and Self-Excited Vibrations in Rotating Machinery,” Adopted from ASME Paper 72-DE-21, General Electric Co., Aircraft Engine Group, Group Engineering Division, May 11, 1972.).

| 431 |

| Centrifugal Compressors | Whirl from Fluid Trapped in the Rotor This type of whirl occurs when liquids are inadvertently trapped in an internal rotor cavity. The mechanism of this instability is shown in Figure 11-13. The fluid does not flow in a radial direction but flows in a tangential direction. The onset of instability occurs between the first and second critical speeds. Table 11-3 is a handy summary for both avoidance and diagnosis of self-excitation and instabilities in rotating shafts.

Figure 11-13: Whirl from fluid trapped in the rotor. (Ehrich, F. F., “Identification and Avoidance of Instabilities and Self-Excited Vibrations in Rotating Machinery,” Adopted from ASME Paper 72-DE-21, General Electric Co., Aircraft Engine Group, Group Engineering Division, May 11, 1972.).

| 432 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Type of Instability

Onset

Frequency Response

Forced Vibration Unbalance

Any speed

Nf = N

Shaft misalignment

Any speed

Nf = 2N

Self-Excited Vibration Hysteretic whirl

N > N1

Hydrodynamic whirl (Oil Whip) Aerodynamic whirl

N > 2N1

Nf ~~ N1 Nf = .5N Nf < .5N

N > N1

Nf = N1

Dry-friction whirl

Any speed

Nf1 = -nN

Entrained fluid

N1 < N < 2N

Nf = N1 .5N < Nf < N

Caused by Nonhomogeneous material Driver and driven equipment misaligned Shrink fits and built-up parts Fluid film bearings and seals Compressor or turbine, tip clearance effects, balance pistons Shaft in contact with stationary guide Liquid or steam entrapped in rotor

Table 11-3: Characteristics of rotor instabilities.

CAMPBELL DIAGRAM The Campbell Diagram is an overall (or bird’s-eye) view of regional vibration excitation that can occur on an operating system. The Campbell diagram can be generated from machine design criteria or from machine operating data. A typical Campbell diagram plot is shown in Figure 11-14. The rotational speed is along the X-axis. The system frequency is along the Y-axis. The speed lines are engine-order lines: one-half engine order, one times engine order, two times engine order, three times engine order, four times engine order, five times engine order, ten times engine order, etc. This form of design study is necessary, especially when designing an axial compressor to determine if a natural blade frequency is excited by a running frequency, its harmonics, or sub harmonics. For example, take the second-stage blade of a hypothetical compressor. Its first flexural natural frequency is calculated and found to be 200 Hz. From the Campbell diagram figure, it is apparent that a forcing frequency of 12,000 rpm produced by operating the compressor at 12,000 rpm will excite the 200Hz first flexural frequency of the blade (200 Hz x 60 – 12,000 rpm). Also, there are five inlet guide vanes ahead of the second-stage blade row. Operating the compressor at 2400 rpm will excite the 200 Hz natural frequency of the blade (200 Hz x 60 = 5 x 2400 rpm).

| 433 |

| Centrifugal Compressors |

Figure 11-14: Campbell diagram.

| 434 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | Following a calculation of the blade natural frequency and a Campbell diagram study of possible excitation sources, it is usual practice to check for the natural frequency band spread by testing the blades on a shaker table. This natural frequency band spread plotted on the Campbell diagram now indicates that operating the compressor between 11,700 rpm and 12,600 rpm should be prohibited. When there are several blade rows on the compressor and several sources of excitation, the designer can be confronted with the difficult task of designing the blade and guide vane rows to meet structural and aerodynamic criteria. Natural blade frequency will be affected by rotational and aerodynamic loading, and it needs to be factored in. In most axial compressors, there are specific operational speed ranges, which are restricted to avoid blade failure from fatigue. To ensure that blade stress levels are within the fatigue life requirements of the compressor, it is usual practice to strain-gauge the blading on one or two prototype machines, measure the stress levels, and generate a Campbell diagram showing the plotted test data. To measure data, an impeller can also be mounted on a shaker table with a variable frequency output (0–10,000 Hz). Accelerometers can be mounted at various positions on the impeller to obtain the frequency responses in conjunction with a spectrum analyzer, as shown in Figure 11-15.

Figure 11-15: Accelerometer locations on impeller tested.

| 435 |

| Centrifugal Compressors | Initially, tests are run to identify the major critical frequencies of the impeller. Mode shapes are then determined visually at each of the critical frequencies. To obtain these mode visualizations, salt is sprinkled evenly on the disc surface. The shaker is maintained at a particular frequency, at which value a given critical frequency is excited for a certain length of time so that the salt particles display the mode shape. The salt accumulates in the nodal regions. Photographs are taken at lower values of these critical frequencies. Photography allows a qualitative identification of the appropriate mode shapes corresponding to each frequency. Figure 11-16 shows an impeller with the mode shapes. The next step in the testing procedure is to record accelerometer readings at various disc, blade, and shroud locations at lower critical frequencies. The objective of this test is to quantitatively identify the high and low excitation regions. For this test, a six- or five-blade region is considered sufficiently large to be representative of the entire impeller. The results of these tests are plotted on a Campbell Diagram, as shown for one such impeller in Figure 11-17. Lines of excitation frequencies are then drawn vertically on the Campbell Diagram, and a line corresponding to the design speed is drawn horizontally. Where the lines of excitation frequencies and multiples of running speed intersect near the line of design rpm,

Figure 11-16: Impeller showing nodal points.

| 436 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | a problem area may exist. If, for instance, an impeller has 20 blades, a design speed of 3000 rpm (50 HZ), and a critical frequency of 1000 Hz, the impeller is very likely to be severely excited, since the critical is exactly 20N. On a Campbell Diagram the previous example will correspond to an exact intersect of the running speed line, 1000 Hz frequency line, and the line of slope 20N. A shrouded centrifugal compressor impeller was tested containing 12 blades and a design speed of 3000 rpm. The 12-bladed impeller’s first excitation mode occurred at a frequency of 150 Hz, resulting in a singleumbrella mode occurring at the contact point between the two back shrouds. At 350 Hz, a coupled mode existed. At these two frequencies, it is the back shroud that is the exciting force. At 450 Hz, a two-diameter mode existed. This mode is characterized by four nodal radial lines and, in many instances, can be the most troublesome mode. This mode is excited by the front shroud and the impeller eye. A double-umbrella mode occurred at 600 Hz. At the last two frequencies, the blade eye experienced high excitation. The Campbell Diagram, Figure 11-17, showed that at design speed this frequency coincided with the 12N line. This coincidence is undesirable, since the number of blades is 12 and may be the exciting force needed to cause a problem. At 950 Hz, a three-diameter mode

Figure 11-17: Campbell diagram of tested impeller.

| 437 |

| Centrifugal Compressors | existed, and at 1100 Hz a four-diameter mode existed. At 1100 Hz, the blade-tip frequency is the predominant forcing function. This impeller seemed to be in trouble at 600 Hz, since this frequency coincided with the number of blades, and should be increased to 15, or the blades should be made out of a thicker stock. This type of analysis is useful mostly in the design stages so, if a problem exists, the machine can be run at a different speed to avert a catastrophe. B E A R I N G A N D S H A F T I N S TA B I L I T I E S One of the most serious forms of instability encountered in journal bearing operation is known as “half-frequency whirl.” “Whirl” is the phenomenon where the shaft center moves in a circular motion in the bearing cavity. There are many types of “Whirling motion,” and most of them are in the direction of rotation, except the “Coulomb Whirl” which is caused by the shaft contacting the bearing surface. The “Oil Whirl” is a selfexcited vibration and characterized by the shaft center orbiting around the bearing center at a frequency of approximately half of the shaft rotational speed, as shown in Figure 11-18.

Figure 11-18: The showing of formation of oil whirl in a rotor system.

| 438 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Figure 11-19: Severity charts: (a) displacement, (b) velocity, (c) acceleration.

| 439 |

| Centrifugal Compressors | As the speed is increased, the shaft system may be stable until the “whirl” threshold is reached. When the threshold speed is reached, the bearing becomes unstable, and further increase in speed produces more violent instability until eventual seizure results. Unlike an ordinary critical speed, the shaft cannot “pass through,” and the instability frequency will increase and follow that half ratio as the shaft speed is increased. This type of instability is associated primarily with high-speed, lightly loaded bearings. At present, this form of instability is well understood, can be theoretically predicted with accuracy, and avoided by altering the bearing design. It should be noted that the tilting pad journal bearing is almost completely free from this form of instability. However, under certain conditions, the tilting pads themselves can become unstable in the form of shoe (pad) flutter, as mentioned previously. All rotating machines vibrate when operating, but the failure of the bearings is mainly caused by their inability to resist cyclic stresses. The level of vibration a unit can tolerate is shown in the severity charts in Figure 11-19. These charts are modified by many users to reflect the critical machines in which they would like to maintain much lower levels. Care must always be exercised when using these charts, since different machines have different size housings and rotors. Thus, the transmissibility of the signal will vary.

BEARINGS The bearings in turbomachines provide support and positioning for the rotating components. Radial support is generally provided by journal or roller bearings, and axial positioning is provided by thrust bearings. Some engines, mainly aircraft jet engines, use ball or roller bearings for radial support, but nearly all-industrial turbomachinery use journal bearings. A long service life, a high degree of reliability, and economic efficiency are the chief aims when designing bearing arrangements. To reach these criteria, design engineers examine all the influencing factors:

| 440 |

•

Load and speed

•

Lubrication

•

Temperatures

•

Shaft arrangements

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | •

Life

•

Mounting and dismounting

•

Noise

•

Environmental Conditions

JOURNAL BEARINGS The heavy frame type gas turbines use journal bearings. Journal bearings may be either full round or split; the lining may be heavy, as in large-size bearings for heavy machinery, or thin, as used in precision insert-type bearings in internal combustion engines. Most sleeve bearings are of the split type for convenience in servicing and replacement. Often in split bearings, where the load is entirely downward, the top half of the bearing acts only as a cover to protect the bearing and to hold the oil fittings. Figure 11-20 shows a number of different types of journal bearings. A description of a few of the pertinent types of journal bearings follows: •

Plain Journal. Bearing is bored with equal amounts of clearance (on the order of one and one-half to two thousands of an inch per inch of journal diameter) between the journal and bearing.

•

Circumferential grooved bearing. Normally has the oil groove at half the bearing length. This configuration provides better cooling but reduces load capacity by dividing the bearing into two parts.

•

Cylindrical bore bearings. Another common bearing type used in turbines. It has a split construction with two axial oil-feed grooves at the split.

•

Pressure or pressure dam. Used in many places where bearing stability is required, this bearing is a plain journal bearing with a pressure pocket cut in the unloaded half. This pocket is approximately 1⁄32 of an inch deep with a width 50% of the bearing length. This groove channel covers an arc of 135 and terminates abruptly in a sharp edge dam. The direction of the rotation is such that the oil is pumped down the channel towards the sharp edge. Pressure dam bearings are for one direction of rotation. They can be used in conjunction with cylindrical bore bearings, as shown in Figure 11-20.

| 441 |

| Centrifugal Compressors |

Figure 11-20: Comparison of general bearing types.

| 442 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | •

Lemon bore or elliptical. This bearing is bored with shims at the split line, which are removed before installation. The resulting bore shape approximates an ellipse with the major axis clearance. Elliptical bearings are for both directions of rotation.

•

Three-lobe bearing. The three-lobe bearing is not commonly used in turbomachines. It has a moderate load carrying capacity. It is restricted to one direction of rotation.

•

Tilting pad bearings. This bearing is the most common bearing type in today’s machines. It consists of several bearing pads posed around the circumference of the shaft. Each pad is able to tilt to assume the most effective working position. Its most important feature is self-alignment when spherical pivots are used. This bearing also offers the greatest increase in fatigue life because of the following advantages: –

Self aligning for optimum alignment and minimum limit

–

Thermal conductivity backing material to dissipate heat developed in oil film

–

A thin babbitt layer can be centrifugally cast with a uniform thickness of about 0.005 of an inch. Thick babbitt greatly reduces bearing life. Babbitt thickness in the neighborhood of .01 reduces the bearing life by more than half

–

Oil film thickness is critical in bearing stiffness calculations. In a tilting-pad bearing one can change this thickness in a number of ways: (a) change in the number of pads, (b) load can be directed on or in-between the pads, (c) changing axial length of pad.

The previous list contains some of the most common types of journal bearings. They are listed in the order of growing stability. All of the bearings designed for increased stability are obtained at higher manufacturing costs and reduced efficiency. The anti-whirl bearings all impose a parasitic load on the journal, which causes higher power losses to the bearings and, in turn, requires higher oil flow to cool the bearing. Many factors enter into the selection of the proper design for bearings. Some of the factors, which affect bearing design follow:

| 443 |

| Centrifugal Compressors | •

Shaft speed range

•

Maximum shaft misalignment that can be tolerated

•

Critical speed analysis, and the influence of bearing stiffness on this analysis

•

Loading the compressor impellers

•

Oil temperatures and viscosity

•

Foundation stiffness

•

Axial movement that can be tolerated

•

Type of lubrication system and its contamination

•

Maximum vibration levels that can be tolerated

T I LT I N G PA D B E A R I N G S Normally, the tilting pad journal bearing is considered when shaft loads are light because of its inherent ability to resist oil whirl vibration. However, this bearing, when properly designed, has a very high loadcarrying capacity. It has the ability to tilt to accommodate the forces being developed in the hydrodynamic oil film and, therefore, operates with an optimum oil film thickness for the given load and speed. This ability to operate over a large range of load is especially useful in high-speed gear reductions with various combinations of input and output shafts. Another important advantage of the tilting-pad journal bearing is its ability to accommodate shaft misalignment. Because of its relatively short length-to-diameter ratio, it can accommodate minor misalignment quite easily. As shown earlier, bearing stiffness varies with the oil-film thickness so that the critical speed is directly influenced to a certain degree by oil-film thickness. Again, in the area of critical speeds, the tilting-pad journal bearing has the greatest degree of design flexibility. There are sophisticated computer programs that show the influence of various load and design factors on the stiffness of tilting-pad bearings journal bearings. The following variations are possible in the design of tilting-pad bearings: •

| 444 |

The number of pads can be varied from three to any practical number.

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | •

The load can be placed either directly on a pad or to occur between pads.

•

The unit loading on the pad can be varied by either adjusting the arc length or the axial length of the bearing pad.

•

A parasitic preload can be designed into the bearing by varying the circular curvature of the shaft.

•

An optimum support point can be selected to obtain a maximum oil-film thickness.

On a high-speed rotor system, it is necessary to use tilting-pad bearings because of the dynamic stability of these bearings. A high-speed rotor system operates at speeds above the first critical speed of the system. It should be understood that a rotor system includes the rotor, the bearings, the bearing support system, seals couplings, and other items attached to the rotor. The system’s natural frequency is, therefore, on the stiffness and damping effect of these components. Commercial, multipurpose tilting-pad bearings are usually designed for multidirectional rotation so that the pivot point is at pad midpoint. However, the design criteria generally applied for producing maximum stability and load-carrying capacity locates the pivot at two-thirds of the pad arc in the direction of rotation. Bearing load is another important design criterion for tilting-pad bearings. Bearing preload is bearing assembly clearance divided by machined clearance C' = ---------------------------------------------------------------------Concentric Pivot Film Thickness Preload Ratio = -----C Machined Clearance

(11-10)

A preload of 0.5–1.0 provides for stable operation because a converging wedge is produced between the bearing journal and the bearing pads. The variable C is an installed clearance and is dependent upon radial pivot position. The variable C is the machine clearance and is fixed for a given bearing. Figure 11-21 shows two pads of a five-pad tilting-pad bearing where the pads have been installed such that the preload ratio is less than 1, and Pad 2 has a preload ratio of greater than 1.0. The solid line in Figure 11-21 represents the position of the journal in the concentric position. The dashed line represents the journal in a position with a load applied to the bottom pads.

| 445 |

| Centrifugal Compressors |

Figure 11-21: Tilting-pad bearing preload.

Unloaded pads are also subject to flutter, which leads to a phenomenon known as “Leading-edge lockup.” Leading-edge lockup causes the pad to be forced against the shaft, and it is then maintained in that position by the frictional interaction of the shaft and the pad. Therefore it is of prime importance that the bearings be designed with preload, especially for low viscosity lubricants. In many cases, manufacturing reasons and the ability to have two-way rotation cause many bearings to be produced without preload. Bearing designs are also affected by the film from a laminar to a turbulent region. The transition speed (Nt) can be computed using the following relationship: v Nt = 1.57x103 --------------(11-11) √DC3 where: v = viscosity of fluid D = diameter (inches) C = diametrical clearance (inches) Turbulence creates more power absorption, thus increasing oil temperature that can lead to severe erosion and fretting problems in bearings. It is desirable to keep the oil discharge temperature below 170ºF (76.7ºC), but with high-speed bearings, this ideal may not be possible. In those cases it is better to monitor the temperature difference between the oil entering and leaving, as shown in Figure 11-22.

| 446 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Figure 11-22: Discharge temperature criteria.

B E A R I N G M AT E R I A L S In all the time since Isaac Babbitt patented his special alloy in 1839, nothing has been developed that encompasses all of its excellent properties as an oil-lubricated bearing surface material. Babbitts have excellent compatibility and non-scoring characteristics and are outstanding in embedding dirt and conforming to geometric errors in machine construction and operation. They are, however, relatively weak in fatigue strength, especially at elevated temperatures and when the babbitt is more than about 0.015 in. (0.38 mm) thick, as seen in Figure 11-23. In general, the selection of a bearing material is always a compromise, and no single composition can include all desirable properties. Babbitts can tolerate momentary rupture of the oil film and may well minimize shaft or runner damage in the event of a complete failure. Tin babbitts are more desirable than the lead based materials, since they have better corrosion resistance, fewer tendencies to separate from the shaft or runner, and are easier to bond to a steel shell. Application practices suggest a maximum design temperature of about 300ºF (148.9ºC) for babbitts, and designers will set a limit of about 50ºF (10ºC) less. As temperatures increase, there is a tendency for the metal to creep under the softening influence of the rising temperature. Creep can occur with generous film thickness and can be observed as ripples on the bearing surface where flow took place. With tin babbitts, observation has

| 447 |

| Centrifugal Compressors |

Figure 11-23: Babbit fatigue characteristics.

shown that creep temperature ranges from 375ºF (190.5ºC) for bearing loads below 200 psi (13.8 Bar.) to about 260ºF-270ºF (126.7ºC-132.2ºC) for steady loads of 1000 psi (69 Bar). This range will be improved by using very thin layers of babbitt, such as in automotive bearings. THRUST BEARINGS The most important function of a thrust bearing is to resist the unbalanced force in a machine’s working fluid and to maintain the rotor in its position (within prescribed limits). A complete analysis of the thrust load must be conducted. As mentioned earlier, compressors with back-toback rotors reduce this load greatly on thrust bearings. Figure 11-24 shows a number of thrust bearing types. Plain, grooved thrust washers are rarely used with any continuous load, and their use tends to be confined to cases where the thrust load is of very short duration or possibly occurs at standstill or low speed only. Occasionally, this type of bearing is used for light loads (less than 50 lb/in), and in these circumstances, the operation is probably hydrodynamic due to small distortions present in the nominally flat bearing surface. When significant continuous loads have to be taken on a thrust washer, it is necessary to machine into the bearing surface of a profile to generate a fluid film. This profile can be either a tapered wedge, or occasionally, a small step.

| 448 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | The tapered-land thrust bearing, when properly designed, can take and support a load equal to a tilting-pad thrust bearing. With perfect alignment, it can match the load of even a self-equalizing tilting-pad thrust bearing that pivots on the back of the pad along a radial line For variablespeed operation, tilting pad thrust bearings, as shown in Figure 11-25, are advantageous when compared to conventional taper-land bearings. The

BEARING TYPE

LOAD CAPACITY

SUITABLE TOLERANCE OF TOLERANCE SPACE DIRECTION OF CHANGING OF ROTATION LOAD/SPEED MISALIGNMENT REQUIREMENT

PLAIN WASHER

POOR

GOOD

MODERATE

COMPACT

MODERATE

POOR

POOR

COMPACT

GOOD

POOR

POOR

COMPACT

GOOD

GOOD

GOOD

GREATER

GOOD

GOOD

TAPER LAND BI-DIRECTIONAL

UNI-DIRECTIONAL TILTING PAD BI-DIRECTIONAL

UNI-DIRECTIONAL

GOOD

GREATER

Figure 11-24: Comparison of thrust bearing types.

Figure 11-25: Various types of thrust bearings.

| 449 |

| Centrifugal Compressors | pads are free to pivot to form a proper angle for lubrication over a long speed range. The self-leveling feature equalizes individual pad loadings and reduces the sensitivity to shaft misalignments, which may occur during service. The major drawback of this bearing type is that standard designs require more axial space than a non-equalizing thrust bearing.

FACTORS AFFECTING THRUST BEARING DESIGN

TEMPERATURE °F

The principal function of a thrust bearing is to resist the thrust unbalance developed within the working elements of a turbomachine and to maintain the rotor position within tolerable limits. After an accurate analysis has been made of the thrust load, the thrust bearing should be sized to support this load in the most efficient method possible. Many tests have proven that thrust bearings are limited in load

Figure 11-26: Thrust bearing temperature characteristics.

| 450 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | capacity by the strength of the babbitt surface in the high load and temperature zone of the bearing. In normal steel-backed babbitted tiltingpad thrust bearings, this capacity is limited to between 250 psi (17.24 Bar) and 500 psi (34.5 Bar) average pressure. It is the temperature accumulation at the surface and pad crowning that cause this limit. The thrust-carrying capacity can be greatly improved by maintaining pad flatness and by removing heat from the loaded zone. By the use of high thermal conductivity backing materials with proper thickness and proper support, the maximum continuous thrust limit can be increased to 1000 psi or more. This new limit can be used to increase the factor of safety and improve the surge capacity of a given size bearing or reduce the thrust bearing size and consequently the losses generated for a given load. Since the higher thermal conductivity material (copper or bronze) is a much better bearing material than the conventional steel backing, it is possible to reduce the babbitt thickness to .010–.030 of an inch. Embedded thermocouples and RTDs will signal distress in the bearing if properly positioned. Temperature monitoring systems have been found to be more accurate than axial position indicators, which tend to have linearity problems at high temperatures. In a change from steel backing to copper-backing, a different set of temperature limiting criteria should be used. Figure 11-26 shows a typical set of curves for the two backing materials. This chart also shows that drain oil temperature is a poor indicator of bearing operating conditions because there is very little change in drain oil temperature from low load to failure load. THRUST BEARING POWER LOSS The power consumed by various thrust bearing types is an important consideration in any system. Power losses must be accurately predicted so that turbine efficiency can be computed and the oil supply system properly designed. Figure 11-27 shows the typical power consumption in thrust bearings as a function of unit speed. The total power loss is usually about 0.8-1.0% of the total rated power of the unit. New vectored lube bearings that are being tested show preliminary figures of reducing the horsepower loss by as much as 30%.

| 451 |

| Centrifugal Compressors |

Figure 11-27: Difference in total power loss data–test minus catalog frictional losses versus shaft speed for 6 x 6 pad doubleelement thrust bearings.

SEALS Seals are very important and often critical components in turbomachinery, especially on high-pressure and high-speed equipment. This chapter covers the principal sealing systems used between the rotor and stator elements of turbomachinery. They fall into two categories (1) non-contacting seals and (2) face seals. Since these seals are an integral part of the rotor system, they affect the dynamic operating characteristics of the machine; for instance, both the stiffness and the damping factors will be changed by seal geometry and pressures. Hence, these effects must be carefully evaluated and factored in during the design of the seal system.

Noncontacting Seals These seals are used extensively in high-speed turbomachinery and have good mechanical reliability. They are not positive sealing. There are two types of noncontacting seals (or clearance seals): labyrinth seals and ring seals.

| 452 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | Labyrinth Seals The labyrinth is one of the simplest of sealing devices. It consists of a series of circumferential strips of metal extending from the shaft or from the bore of the shaft housing to form a cascade of annular orifices. Labyrinth seal leakage is greater than that of clearance bushings, contact seals, or film-riding seals. Consequently, labyrinth seals are utilized when a small loss in efficiency can be tolerated. They are sometimes a valuable adjunct to the primary seal. In large gas turbines, labyrinth seals are used in static as well as dynamic applications. The essentially static function occurs where the casing parts must remain unjoined to allow for differences in thermal expansion. At this junction location, the labyrinth minimizes leakage. Dynamic labyrinth applications for both turbines and compressors are interstage seals, shroud seals, balance pistons, and end seals. The major advantages of labyrinth seals are their simplicity, reliability, tolerance to dirt, system adaptability, very low shaft power consumption, material selection flexibility, minimal effect on rotor dynamics, back diffusion reduction, integration of pressure, lack of pressure limitations, and tolerance to gross thermal variations. The major disadvantages are the high leakage, loss of machine efficiency, increased buffering costs, tolerance to ingestion of particulates with resulting damage to other critical items such as bearings, the possibility of the cavity clogging due to low gas velocities or back diffusion, and the inability to provide a simple seal system that meets OSHA or EPA standards. Because of some of the foregoing disadvantages, many machines are being converted to other types of seals. Labyrinth seals are simple to manufacture and can be made from conventional materials. Early designs of labyrinth seals used knife-edge seals and relatively large chambers or pockets between the knives. These relatively long knives are easily subject to damage. The modern, more reliable labyrinth seals consist of sturdy closely spaced lands. Some labyrinth seals are shown in Figure 11-28. Figure 11-28a is the simplest form of the seal. Figure 11-28b shows a grooved seal that is more difficult to manufacture but produces a tighter seal. Figure 11-28c and 11-28d are rotating labyrinth-type seals. Figure 11-28e shows a simple labyrinth seal with a buffered gas for which pressure must be maintained above the process gas pressure and the outlet pressure (which can be greater than or less than the atmospheric pressure). The buffered gas produces a fluid barrier to the process gas. The eductor sucks gas from the vent near the atmospheric end. Figure 11-28f shows a buffered, stepped labyrinth. This step labyrinth gives a tighter seal. The matching stationary seal is usually manufactured from soft materials such as babbitt or bronze, while the

| 453 |

| Centrifugal Compressors |

Stationary

Stationary sleeve

Stationary

Stationary sleeve

Figure 11-28: Various configurations of labyrinth seals.

| 454 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | stationary or rotating labyrinth lands are made from steel. This composition enables the seal to be assembled with minimal clearance. The lands can, therefore, cut into softer materials to provide the necessary running clearances for adjusting to the dynamic excursions of the rotor. To maintain maximum sealing efficiency, it is essential that the labyrinth lands maintain sharp edges in the direction of the flow. This requirement is similar to that in orifice plates. A sharp edge provides for maximum vena contracta effect, as seen in Figure 11-29, and hence maximum restriction for the leakage flows. High fluid velocities are generated at the throats of the constrictions, and the kinetic energy is dissipated by turbulence in the chamber beyond each throat. Thus, the labyrinth is a device wherein there is a multiple loss of velocity head. With a straight labyrinth, there is some velocity carryover that results in a loss of effectiveness, especially if the throats are closely spaced. To maximize the aerodynamic blockage effect of this carryover, the diameters can be stepped or staggered to cause impingement of the expanding orifice jet on a solid transverse surface. The leakage is approximately inversely proportional to the square root of the number of labyrinth lands. Thus, if leakage is to be cut in half at a four-point labyrinth, the

Figure 11-29: Theory behind the knife-edge arrangement.

| 455 |

| Centrifugal Compressors | number of lands would have to be increased to 16. The leakage formula can be modified and written as:  g   V ( P0 − Pn )   m l = 0.9 A  0  n + Pn  ln  P0  

1/ 2

(11-12)

For staggered labyrinths, the equation can be written as:  g   V ( P0 − Pn )   ml = 0.75 A 0 Pn   n + ln  P0  

1/ 2

(11-13)

where ml = leakage, lb/sec A = leakage area of single throttling, sq ft P0 = absolute pressure before the labyrinth, cu ft/lb V0 = Specific volume before the labyrinth, lb (m)/sq ft Pn = absolute pressure after the labyrinth, lb (f)/sq ft n = number of lands The leakage of a labyrinth seal can be kept to a minimum by providing (1) minimum clearance between the seal land and the seal sleeve, (2) sharp edges on the lands to to reduce the flow discharge coefficient, and (3) grooves or steps in the flow path for reducing dynamic head carryover from stage to stage. Operational tests should:

| 456 |

•

Detect and correct all leaks

•

Determine relief pressures and check for proper operation of each relief valve

•

Accomplish a filter cooler changeover without causing startup of the standby pump

•

Demonstrate the control valves have suitable capacity, response, and stability

•

Demonstrate the oil pressure control valve can control oil pressure.

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | Mechanical Seals A typical mechanical contact shaft seal has two major elements, as seen in Figure 11-30. They are the oil-pressure-gas seal and breakdown bushing. This seal will normally have buffering via a single ported labyrinth located inboard of the seal and a positive shutdown device which will attempt to maintain gas pressure in the casting when the compressor pressure is at rest and the seal oil is not being applied. For shutdown, the carbon ring is kept tightly sandwiched between the rotating seal ring and stationary sleeve with gas pressure to prevent gas from leaking out when no oil pressure is available. In operation, seal oil pressure is held at a differential of 35–40 psid over the process gas pressure which the seal is sealing against. This high pressure oil can be seen entering in the top in Figure 11-30 and completely fills the seal cavity. Some of the oil (a relatively small percentage, ranging from 2 to 8 gpd per seal depending on machine size) is forced across the carbon ring seal faces which are sandwiched between the rotating seal ring (rotating at shaft velocity) and the stationary sleeve (non-rotating and

Figure 11-30: Mechanical contact shaft seal.

| 457 |

| Centrifugal Compressors | forced against the carbon ring by a series of peripheral springs). Therefore, the actual rotative speed of the carbon ring can be anywhere between zero rpm and full rotational speed. Oil crossing these seal faces contacts the process gas and is “contaminated oil.” The majority of the oil flows out of the uncontaminated seal oil drain after taking a pressure drop from design seal oil pressure to atmospheric pressure across the breakdown bushing. An orifice is placed in parallel with the breakdown bushing to meter the proper amount of oil flow for cooling. The contaminated oil leaves through the drain to a degasifier for purification. The bearing oil drain can either be combined with the uncontaminated seal oil drain or kept separate; however, a separate system will increase bearing span and lower critical speeds. MECHANICAL SEAL S E L E C T I O N A N D A P P L I C AT I O N The following is a list of factors that have proven helpful in seal system design and selection: •

Product

•

Seal environment

•

Seal arrangement

•

Equipment

•

Secondary packing

•

Seal-face combinations

•

Seal gland plate

•

Main seal body

Product The physical and chemical properties of the liquid being sealed will place constraints upon the type of seal arrangement, the materials of construction, and the seal design that can be used.

Pressure The relative pressures of the material to be sealed affect the decision of whether to use a balanced or unbalanced seal design. Pressure also affects the choice of face material because of the seal-face loading.

| 458 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | If the service happens to be below atmospheric pressure, then special considerations are required to seal the material effectively. Most unbalanced seal designs are applicable up to 100-psig stuffing-box pressure. At more than 100 psig, balanced seals should be used. Seal manufacturers base their seal-face combinations of designs on PV ratings. These are the multiple of the (P) and the sliding velocity (V) of the faces. The maximum PV rating for an unbalanced seal is about 2,250,000 for a balanced seal.

Temperature The temperature of the liquid being pumped is important because it affects the seal-face material selection as well as face wear life. This is primarily a result of changes in lubricity of the fluid with changes in temperature. Common seal designs may handle fluid temperatures in the 0ºF to 200ºF range, special metal bellows seals may be used up to the 650ºF range. Low temperature (-100ºF) also requires special arrangements, since most hydrocarbons have little lubricity in this range. The most important consideration concerning temperature is to avoid operating close to a temperature that allows flashing of the liquid. Mechanical seals work well on many liquids; they work poorly on most gases.

Lubricity In any mechanical seal design, there is a rubbing motion between the dynamic seal faces. This rubbing motion is most often lubricated by the fluid being pumped. Therefore, the lubricity of the pumped liquid at the given operating temperature must be considered to determine if the chosen seal design and face combination will perform satisfactorily. Most seal manufacturers limit the speed of their seals to 90 fps with good lubrication of the faces. This is primarily due to the centrifugal forces acting on the seal, which tend to restrict its axial flexibility.

Abrasion When evaluating the possibility of installing a seal in a liquid, which has entrained solids, several factors must be considered. Is the seal constructed in such a way that the dynamic motion of the seal will be restricted by fouling of the seal parts? The seal arrangement that is usually preferred when abrasives are present is a flushed single type with

| 459 |

| Centrifugal Compressors | a face combination of very hard material. However, factors such as toxicity or corrosiveness of the material may dictate that other arrangements be used.

Corrosion When considering the corrosiveness of the material being pumped, (that is, whether the binder or the carbon or tungsten carbide will be attacked, or whether the base metal of the plated seal-face will be attacked), and what type of elastomer or gasket material can be used. The corrosion rate will affect the decision of whether to use a single or multiple-spring design because the spring can tolerate a greater amount of corrosion without weakening it appreciably.

Toxicity This factor is becoming an increasingly important consideration in the design of mechanical seals. Since the rubbing seal faces require liquid penetration to cool and lubricate them, it is reasonable to expect that there will be some vapor passing across the faces. This is, in fact, the case. A normal seal can be expected to “leak” from a few ppm to 10 cc/min. It is also generally accepted that the seal leakage rate will increase with speed.

Additional Product Considerations.

| 460 |

•

Is the product thermosensitive? The heat generated by the seal faces may cause polymerization.

•

Is the product shear sensitive, i.e., will it harden due to turbulence?

•

If the product is highly flammable, be aware of possible ignition sources.

•

In hazardous services, plan for personnel protection in the event of seal leakage.

•

Products with dissolved gas must be properly vented. In most, vent the stuffing box back to pump action.

•

Seals in cold services are extremely sensitive to moisture. There must be a way to “dry out the system” after repair.

•

Consideration must be given to the pressure and temperature that the seal will see during normal operation, start-up, shutdown, and upset conditions.

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | •

Vapor pressure of the product must be known in order to prevent vaporization in the stuffing box.

SEAL ENVIRONMENT Once an adequate definition of the product is made, the design of the seal environment can be selected. The following are general parameters that an environmental system may regulate or change: •

Pressure control

•

Temperature control

•

Fluid replacement

•

Atmospheric air elimination

The most common environmental control systems include flushing, barrier fluids, quenching, and heating/cooling systems. Each has its own use in regulating the parameters mentioned previously. S E A L A R R A N G E M E N T C O N S I D E R AT I O N There are four seal arrangement considerations: 1.

Double seals have been the standard with toxic and lethal products, but maintenance problems and the seal design contribute to poor reliability. The double face-to-face seal should be looked at more closely.

2.

Do not use a double seal in dirty service—the inside seal will hang up.

3.

The API standard is a good guide to the use of balanced and unbalanced seals.

4.

Application of a balanced seal at too low a pressure may encourage face lift-off.

The number of arrangements and auxiliary features are more than 100. Regardless of the seal vendor, the arrangement will generally determine success.

| 461 |

| Centrifugal Compressors | Equipment Too few people consider the equipment with the seal selection. In most cases, poor equipment will give poor seal performance, regardless of the deal or arrangement chosen. Also, beware that different pumps with the same shaft diameter and Total Differential Head (TDH) may present different sealing problems. (Note: these same considerations may be used for troubleshooting.) Secondary packing combinations have come a long way in the last 8–10 years. Stellite is being phased out in petroleum and petrochemical seal applications. Better grades of ceramic are being offered as the standard material. The cost of tungsten carbide has decreased considerably.

Dry Gas Seals The use of dry gas seals in process gas centrifugal compressors has increased over the last thirty years, replacing traditional oil film seals in most applications. More than 85% of centrifugal gas compressors manufactured today are equipped with dry gas seals. Dry gas seals are basically mechanical face seals consisting of a mating ring, which rotates, and a primary ring, which is stationary. A cross-sectional view of a dry gas seal is shown in Figure 11-31. The rotating assembly consists of the mating ring (with spiral grooves) mounted on a shaft sleeve held in place axially with a clamp sleeve and a locknut. It is typically pin driven. The mating ring with spiral grooves and the primary ring are held within the retainer assembly. The stationary assembly consists of the primary ring mounted in a retainer assembly held stationary within the compressor housing. Under static conditions, the primary and mating rings are held in contact due to the spring load on the primary ring. The spiral groove pattern, for a clockwise rotation, on the mating ring is shown in Figure 11-32. The operating principle of the spiral grooved gas seal is that of a hydrostatic and hydrodynamic force balance. As gas enters the grooves, it is sheared towards the center. The sealing dam acts as a restriction to the gas outflow, thereby raising the pressure upstream of the dam. This increased pressure causes the flexibly mounted, primary ring to separate from the mating ring. During normal operation, the running gap is approximately 3 microns. Under pressurization, the forces exerted on the seal are hydrostatic and are present whether the mating ring is stationary or rotating. Hydrodynamic forces are generated only upon

| 462 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Figure 11-31: Single dry gas seal.

Figure 11-32: Spiral grooved mating ring (courtesy Proceedings Seventeenth Turbomachinery Symposium, “Dry Gas Compressor Seals” by Piyush Shah, John Crane, Inc. )

| 463 |

| Centrifugal Compressors | rotation. The mating ring consisting of the logarithmic spiral grooves is the key to generating these hydrodynamic forces. During operation, the grooves in the mating ring generate a hydrodynamic force that causes the primary ring to separate from the mating ring creating a “running gap” between the two rings, which effectively seals against the process gas. During normal operation, the running gap is approximately 3 microns. A sealing gas is injected into the seal to provide the working fluid, which establishes the running gap.

Operating Range of Dry Gas Seals Gases ranging from inert gases such as nitrogen to highly toxic gaseous mixtures of natural gas and hydrogen sulfide, can be sealed utilizing the optimum seal arrangements. The operating range of the spiral grooved dry gas seals is as follows: •

Sealed Pressure: 2400 psi (165 bar)

•

Temperature: 500°F (260ºC)

•

Surface Speed: 500 ft./sec. (152 m/sec)

•

M.W.: 2–60

Dry Gas Seal Materials The gas composition, contaminants in the gas stream, operating temperatures, and process conditions dictate the choice of materials. The most common materials of construction are as follows: •

Mating Ring: Tungsten Carbide, Silicon Carbide

•

Primary Ring: Carbon, Silicon Carbide

•

O-Rings: Elastomers (Viton, Kalrez)

•

Hardware: 300 or 400 series ss (sleeves, discs, retainer rings)

•

Coil Springs: 316 ss, Hastelloy

Dry Gas Seal Systems The use of dry gas seals requires a system designed to supply sealing gas to the seal as a working fluid for the running gap. These gas seal systems are normally supplied by the compressor OEM mounted on the compressor base plate. There are two basic types of gas seal systems, differential pressure (∆P) control and flow control. Differential control systems control the supply of seal gas to the seal by regulating the seal gas pressure to a predetermined value typically 15 psi (1 bar) above the

| 464 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | sealing pressure. This is accomplished through the use of a differential pressure control valve. Flow control systems control the supply of seal gas to the seal by regulating the seal gas flow through an orifice upstream of the seal. This is accomplished through the use of a differential pressure control valve monitoring pressures on either side of the orifice.

Dry Gas Seal Degradation Contamination of the seal by foreign objects leads to seal failures. The running gap between the primary and mating gas seal rings is typically around 3 microns. Injection of any type of solids or liquids into this very narrow seal running gap can cause degradation of seal performance. This would create excessive gas leakage to the vent and eventual failure of the seal. Since the typical operating gaps between the two sealing surfaces range from 0.0001 in. to 0.0003 in., the resultant leakage is very small in magnitude. Under conditions of static pressurization beyond 50–75 psi (3.4–5.17 bar), the seal leaks a very small amount. This leakage increases with increasing pressure and reduces with increasing temperature. Increased viscosity of gases at higher temperatures reduces the amount of seal leakage. For example, a 4 in. (101.6 mm) shaft seal on a natural gas compressor statically pressurized to 1000 psi (69 bar) will leak about 1 scfm (0.03 scmm). Under dynamic condition, due to the pumping effect of the spiral grooves, the leakage increases as well. The power loss can also be increased with seal contamination. The seal surfaces being non-contacting under dynamic conditions, the power loss associated with dry gas seals is very small. The power loss for a 10 in. (254 mm) seal operating at 1000 psi (69 Bar) and 10,000 rpm is about 12–14 kW. With damaged seal surfaces, these losses can be increased by 20%–30%. Foreign material within the seal results in increased shearing forces between the primary and mating rings, causing overheating of the seal components, leading to o-ring extrusion or some other mechanical form of seal failure. The major areas from which gas seals contamination occurs are: •

Process gas leakages from the inboard or high-pressure side of the seal

•

Bearing lubrication oil from the outboard or low- pressure side of the seal

•

The seal gas injected into the seal being contaminated upstream of the injection

| 465 |

| Centrifugal Compressors | Contamination from Process Gas Contamination from process gas can occur when there is an insufficient supply of sealing gas into the seal, allowing process gas to come into direct contact with the seal ring faces. Contaminants existing within the process gas can then damage the seal.

Contamination from Bearing Lubrication Oil A barrier seal is required on the outboard side of the dry gas seal, between the gas seal and the compressor bearing. The primary function of the barrier seal, typically buffered with air or nitrogen, is to prohibit the flow of bearing lubrication oil into the gas seal. Contamination of the dry gas seal from lube oil can occur when the barrier seal fails to function as intended.

Contamination from Seal Gas Supply Contamination from the seal gas supply occurs when the sealing gas is not properly treated upstream of the dry gas seal. Gas seal manufacturers have stringent requirements for seal gas quality. Typically, the sealing gas must be dry and filtered of particles 3 micron and larger. Filters are normally provided in the gas seal system to meet this requirement. Dry gas seals operate under extremely tight tolerances, which demand that special care be taken in the design of the gas seal environment and in the operation of the compressor and gas seal system. While the threat of seal degradation and reduced seal life due to outside influences is real, the detrimental effects of these factors can be minimized. The replacement of mechanical seals by dry gas seals must be closely examined. There have been many cases where the replacement has caused the compressor to operate in an unstable manner. This is due to the fact that removal of the mechanical seal causes a change in the damping of the rotor and can cause the rotor to operate closer to its critical speed.

GEARS Gearing is one of the most important components between prime movers and driven units. If gearing is not selected properly, it can cause many problems. Gearing transmits great power at high rotational speeds. Recent advances in turbomachinery technology, especially in turbines, compressors, couplings, and bearings, have required gearing to withstand

| 466 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | high external forces. To design problem-free equipment, it is important to consider the effect of the external system on gearing. Thus, all the factors that influence design, application, and operation of gear drives, should be considered in the design phase. Since problems encountered with gears are complex, it is unfair to blame them on the gear manufacturers alone. The gear supplier is much less informed about the package than any other group. Problems should be handled as a team effort between manufacturers and users. One factor causing problems is that the system is not timed in terms of spring constants and masses. The gear is usually the only item required to operate with metal parts in such close contact with other components. This setup can result in early failure. Gearing is also subjected to cyclic loading varying from 0 to 55,000 cycles per minute. With current materials and heat-treating techniques, the use of high hardness gearing with tooth loads of 1500–2000 pounds per inch of face at pitchline velocities of 20,000–30,000 feet per minute is not at all uncommon. In turbine-driven test equipment gear drives have been built with pitchline velocities as high as 55,000 feet per minute and rotational speeds approaching 100,000 rpm. The magnitude of internal forces and material stresses coupled with the high speeds has resulted in gear drives, which are dynamically complicated and sensitive to influences from other components in the system. The system characteristics of the entire train must be known so that the selection of the gear will be proper. The major points affecting the system are: (1) couplings, (2) vibration, (3) operation conditions, (4) thrust loads, and (5) mounting type. Couplings are a constant source of unbalance vibration, and critical speed changes can be attributed to spacer shift and wear. Coupling lockeys can also cause severe housing vibration, while shaft vibration can remain low. Therefore, it is important to monitor vibration with accelerometers in addition to proximity probes. Gear failure from high vibration is common when the gear and pinion teeth operate within a few hundred microinches of each other. Accelerometers can also monitor gear mesh frequencies and thus act as early warning devices. Operating conditions must be known in detail. In many cases the gear manufacturer is provided with only the design horsepower of the machine. Actual transmitted loads can be much higher due to proximity of torsional or lateral critical speeds. Surge in centrifugal compressors can cause severe overload and result in failures.

| 467 |

| Centrifugal Compressors | External thrust loads are another major problem, and in many cases they lead to double-helical gear selection. The gear housing and the mounting type of the gear train are very important considerations in the overall life of the unit, since improper mounting and expansion of the gear housing can lead to misalignment problems. A substantial structure to support the gear drive weight, thrust, and torque reactions with minimum load deflections must be provided. At least two dowels for locating each gear housing are required, and it is necessary to minimize housing vibration from whatever source. Ideally, the structures should be steel-reinforced concrete filled with grout. The inclusion of oil reservoirs in the structure supporting major train components should be avoided, since unavoidable thermal changes will have adverse effects on alignment. If a reinforced concrete or a filled structure cannot be provided, resonance due to train component mass and structure stiffness at system rotational frequencies or harmonies should be avoided. GEAR HOUSINGS Gear housings are made from materials such as cast iron, steel, or aluminum. Before final matching, the gear housing must be stress relieved for dimensional stability. Housing should also be rigid enough to resist misalignment. A sufficient clearance should be provided around the gears to prevent oil choking. To prevent thermal distortion, the design should be able to maintain uniform case temperatures. Gear housings can cause alignment problems from thermal distortion. Oil in gear cases are used both for lubrication and cooling. Oil should be supplied in the temperature and pressure range specified by the manufacturer. Up to a pitch-line speed of approximately 15,000 feet per minute, the oil should be sprayed into the outmesh. Spraying allows maximum cooling time for the gear blanks and applies the oil at the highest temperature area of the gears. Also, a negative pressure is formed when the teeth come out of the mesh, pulling the oil into the tooth spaces. At more than approximately 15,000 feet per minute, 90% of the oil should be sprayed into the outmesh and 10% into the inmesh. This procedure is a safety precaution to assure the amount of oil required for lubrication is available at the mesh. When the speed ranges from 25,000 to 40,000 feet per minute, oil should be sprayed on the sides and gap area (on double-helical) of the gears to minimize thermal distortion.

| 468 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | COUPLINGS In most turbomachines, couplings attach the driver to the driven piece of machinery. High-performance flexible couplings used in turbomachines must perform three major functions: (1) efficiently transmit mechanical power directly from one shaft to another with constant velocity, (2) compensate for misalignment without inducing high stress and with minimum power loss, and (3) allow for axial movement of either shaft without creating excessive thrust on the other. There are three basic types of flexible couplings that satisfy these requirements. The first type is the mechanical-joint coupling. In this coupling, flexibility is accomplished by a sliding and rolling action. Mechanical-joint couplings include gear tooth couplings, chain and sprocket couplings, and slider or Oldham couplings. The second type is the resilient-material coupling. In resilient-material couplings, flexibility is a function of flexing of material. Resilient-material couplings include those that use elastomer in compression (pin and bushing, block, spider, and elastomer-annulus, metal-insert types); elastomer in shear (sandwich type, tire type); steel springs (radial leaf, peripheral coil types); steel-disc and diaphragm couplings. The third type is the combined mechanical and material couplings where flexibility is provided by sliding or rolling and flexing. Combination couplings include continuous and interrupted metallic-spring grid couplings, nonmetallic gear couplings, nonmetallic chain couplings, and slider couplings that have nonmetallic sliding elements. In choosing a coupling, the loading and speed must be known. Figure 11-33 shows the relation between coupling type, peripheral velocity coupling size, and speed. The loadings in these high-performance flexible couplings are as follows: •

Centrifugal force. Varies in importance, depending on the system speed.

•

Steady transmitted torque. Smooth nonfluctuating torque in electric motors, turbines, and a variety of smooth torqueabsorbing load (driven) machines.

•

Cyclically transmitted torque. Pulsating or cyclic torque in reciprocating prime movers and load machines such as reciprocating compressors, pumps, and marine propellers.

| 469 |

| Centrifugal Compressors | •

Additional cyclic torque. Caused by machining imperfections of drive components (particularly gearing) and imbalance of rotating drive components.

•

Peak torque, (transience). Caused by starting conditions, momentary shock, or overload.

•

Impact torque. A function of system looseness or backlash. Generally, mechanical joint flexible couplings have inherent backlash.

•

Misalignment loads. All flexible couplings generate cyclic or steady moments within themselves when misaligned.

•

Sliding velocity. A factor in mechanical-joint couplings only.

•

Resonant vibration. Any of the forced vibration loads, such as cyclic or misalignment loads, may have a frequency which coincides with a natural frequency of the rotating shaft system, or any component of the complete power plant and its foundation, and may thus excite vibration resonance.

The gas turbine is a high-speed, high-torque drive and requires that its coupling have the following characteristics. •

Low-weight, low-overhung moment

•

High-speed, capacity-acceptable centrifugal stresses

•

High balancing potential

•

Misalignment capability

Gear couplings, disc couplings, and diaphragm-type couplings are best suited for this type of service. Table 11-4 shows some of the major characteristics of these types of couplings. GEAR COUPLINGS A gear coupling consists of two sets of meshing gears. Each mesh has an internal and external gear with the same number of teeth. There are two major types of gear couplings that are used in turbomachinery. The first type of gear coupling has the male teeth integral with the hub as seen in Figure 11-34. In this coupling type, the heat generated at the teeth flows in a different way into the shaft than it does through the sleeve to the surrounding air. The sleeve will therefore heat up and expand more than

| 470 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Figure 11-33: Flexible coupling operating spectrum.

Speed capacity Power-to-weight ratios Lubrication required Misalignment capacity at high speed Inherent balance Overall diameter Normal failure mode Overhung moment on machine shafts Generated moment, misaligned with torque Axial movement capacity Resistance to axial movement Suddenly applied Gradually applied

Disc

Diaphragm

Gear

High Moderate No Moderate Good Low Abrupt (fatigue) Moderate Moderate

High Moderate No High Very good High Abrupt (fatigue) Moderate Low

High High Yes Moderate Good Low Progressive (wear) Very low Moderate

Low

Moderate

High

High High

Moderate Moderate

High Low

Table 11-4: Disc, diaphragm, and gear couplings. This table is intended as a rough guide only.

| 471 |

| Centrifugal Compressors |

Figure 11-34: Gear coupling (male teeth integral with the hub.)

Figure 11-35: Gear coupling (male teeth integral with the spool).

the hub. This expansion, plus the centrifugal force acting on the sleeve, will cause it to grow rapidly as much as 3–4 mils more than the hub, causing an eccentricity, which can lead to a large, unbalanced force. Thus, this coupling type is more useful in low power units. The second type of coupling, shown in Figure 11-35, has the male teeth integral with the spool. In this coupling type, the same amount of heat is produced, but the hollow-bored spool will accept heat in a manner similar to the sleeve so that no differential growth occurs. Gear couplings have a pilot incorporated into the male tooth form to support the loose member of the coupling in a concentric manner at speed, as shown in Figure 11-36.

| 472 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Figure 11-36: Schematic of gear used in coupling applications.

The sliding friction coefficient is another area of evaluation in gear couplings. It produces a resistance to the necessary axial movement as rotors heat and expand. This relative sliding motion between the coupling elements takes care of the misalignment problem in gear couplings. Relative motion between the meshing gears is oscillatory in the axial direction and has a low amplitude and a relatively high frequency. Some of the major advantages of the gear couplings are: •

They can transmit more power per pound of steel, or per inch of diameter, than any other coupling.

•

They are forgiving; they accept errors in installation and mistreatment more readily than other types of couplings.

•

They are reliable and safe; they do not throw around pieces of metal or rubber even when they fail, and they can work longer in corrosive conditions than many other couplings.

A major disadvantage in gear couplings is the misalignment problem. Tooth sliding velocity is directly proportional to the tooth-mesh misalignment angle and the rotational speed. Therefore, misalignment of highspeed drives must be kept to a minimum to limit sliding velocity to an acceptable value.

| 473 |

| Centrifugal Compressors | The coupling must be able to accommodate misalignment caused by cold startup. The physical misalignment capability of a gear-type coupling should never be considered an acceptable running condition for highspeed applications. The limits of misalignment versus operating speed are best stated on the basis of a constant, relative sliding velocity between the gear teeth. Figure 11-37 gives recommended limits of misalignment with the system at operating temperature. The graph is based on a maximum constant sliding velocity of 1.3 inches per second and includes coupling size, speed, and the axial distance between gear meshes. Gear couplings can be more tolerant of axial growth than other coupling types. In the disc-type couplings, the axial growth is limited by the disc deflection range, so the equipment must be adjusted with more axial accuracy than with gear couplings. High-speed couplings must be balanced very carefully, and, with a low overhung moment, the effect of coupling overhung moment is felt not only in machine bearing load but also in shaft vibration. The advantage of a reduction in overhung moment is not only to reduce bearing loads but also to minimize shaft deflection, which results in a reduction of the vibration amplitude. The reduction of the coupling

Figure 11-37: Recommended limits of misalignment vs. operating speed.

| 474 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | overhung moment produces an upward shift in shaft critical speeds. This change in natural frequencies results in an increase in the spread between natural frequencies. For many applications, reduced overhung moment is an absolute necessity to enable the system to operate satisfactorily at the required operating speed. The high-speed couplings have five components; usually two hubs, two sleeves, and a spacer. To obtain a proper balance, each hub should be balanced separately, then the spacer should be balanced, and finally the full coupling should be assembly balanced. The couplings should be carefully match-marked before removal from the balancing mandrels. Lubrication problems are a major consideration in the use of gear couplings. Relative sliding between the teeth of the hub and the sleeve requires proper lubrication to assure long component life. This sliding motion is alternative and is characterized by small amplitudes and relatively high frequencies. Gear couplings can be either packed with lubricant or continuously lubricated. Each system has advantages and disadvantages, and the choice depends on the conditions under which the coupling works. OIL-FILLED COUPLINGS Very few high-performance couplings use this system because it requires large-volume couplings. However, it is the best method of lubrication and, incidentally, the first used. Its major disadvantage is that it may leak lubricant from defective flange gaskets, etc. G R E A S E - PA C K E D C O U P L I N G S Besides enabling the user to select a good lubricant, grease packing has the advantage of sealing the coupling from the environment. The highperformance coupling works under very small misalignment and usually generates very little heat. In most cases, the couplings receive more heat from the shafts than they generate. Very few types of grease can work in temperatures of more than 250°F, and, for this reason, grease-packed couplings cannot be installed within an enclosure that prevents the heat from dissipating. Greases also separate under large centrifugal forces. In many high-speed couplings, forces exceed 8000 g1S.5. New lubricants are appearing on the market that do not separate under high loadings.

| 475 |

| Centrifugal Compressors | A second disadvantage of grease lubrication is the maintenance requirement. Coupling manufacturers generally recommend relubrication every six months. There are known cases, however, where grease-packed couplings were found to be in excellent condition after two years of maintenance-free service. C O N T I N U O U S LY L U B R I C AT E D C O U P L I N G S Lubrication by continuous oil flow can represent an ideal method if there is: •

Freedom to select the type of oil

•

Independent lube circuit

From the user’s point of view, neither condition is acceptable, not only because of the added cost of an independent lube circuit but because it is almost impossible to prevent mixing of the oil from this circuit with the lube system for the rest of the equipment. In practice, continuously lubricated couplings are supplied with oil from the main lube system. The oil is not the best type for couplings, and also brings to the coupling a large quantity of impurities. The accumulated sludge shortens coupling life. Sludge accumulates within a coupling for two reasons: (1) because the lubricant is not pure and (2) because the coupling centrifuges and retains the impurities. Very little can be done to prevent the coupling from retaining the impurities. The G-forces in a coupling are very high, and the oil dam built in the sleeve configuration prevents the impurities from going over it. Some manufacturers now offer couplings without a dam, or with sleeves provided with radial holes. Experience has shown that such couplings accumulate no sludge. The dam has, however, two useful purposes: 1.

It maintains an oil level high enough to submerge the teeth completely.

2.

It retains a quantity of oil within the coupling even if the lube system fails.

Removing the oil dam defeats both these features. To maintain the same performance for a damless coupling, the oil flow to the coupling should be reevaluated. Nothing can be done, however, to retain oil in the

| 476 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | damless coupling, and some users will not accept them for this reason. A proper decision can be made only by weighing a possible coupling failure because of sludge accumulation against an accidental failure of the lube system. M E TA L D I A P H R A G M C O U P L I N G S The metal diaphragm coupling is relatively new in turbomachinery applications. Although the first recorded use of such a coupling dates back to 1922 on a condensing steam turbine locomotive, the contoured diaphragm did not come into wide use until the late 1950s. Diaphragm couplings accommodate system misalignment through flexing. Fatigue resistance is the main performance criterion. The life expectancy of a diaphragm coupling that operates within its design limits is theoretically infinite. Figure 11-38 shows a section through a diaphragm coupling. The coupling has only five parts: two rigid hubs, one spool piece, and two alignment rings. These five parts are solidly bolted together, and misalignment is accommodated through flexing of the two diaphragms of the spool. The spool piece is made up of three separate parts: two diaphragms and a spacer tube. These parts are welded together by an electron beam.

Figure 11-38: Schematic of a typical diaphragm coupling (courtesy of Koppers Company, Inc. ).

| 477 |

| Centrifugal Compressors | The heart of these couplings is the flexing disc; it is manufactured from vacuum-degassed alloy steel, forged with a radial grain orientation, and has a contoured profile machined on high-precision equipment. The contoured profile is shown in Figure 11-39. The diaphragm undergoes axial deflection. The forces acting on the disc that are generating the stresses are caused by the torque effects, centrifugal forces, and axial deflection. Standard methods for calculating centrifugal forces in a rotating disc show that both tangential and radial stresses increase rapidly with a decrease in the radius.

Figure 11-39: Axial deflection in a disc.

The stresses imposed by axial deflection are much greater at the hub than at the rim, as seen in Figure 11-40. Therefore, to maintain uniform stresses in the diaphragm when all the various forces acting on the diaphragm are at their maximum, the diaphragm must be used to connect the contoured profile at both the hub and the rim to reduce stresses. Diaphragm couplings are more susceptible to axial movement problems than gear couplings, since the diaphragm has a maximum deflection that cannot be exceeded. Theoretically, a diaphragm coupling will have no problems or failures as long as it is operated within “design limits.” The diaphragm fails from excessive torque. Two distinct modes of failure can be found: one at zero axial displacement and the other at large axial displacement. Zero axial displacement is characterized by a circular crack line that goes through the thinnest portion of the diaphragm. The crack is relatively smooth, and there is no buckling of the disc. The large axial movement and angular misalignment, which lead to disc failures are characterized by a crack line

| 478 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears |

Figure 11-40: Stress distribution under axial deflection.

that follows a random path from the thinnest to the thickest portion of the disc. The crack line is very irregular, and there is severe buckling of the unfailed part of the disc. Failure in this mode shows that the crack line propogates some 270º before disc buckling takes place, indicating that the torque load makes only a small contribution to the total stresses in the disc. Metal diaphragm couplings can also have problems due to corrosive action on the diaphragms. Thus, care must be taken to apply coating to protect against damage from a harsh environment. M E TA L D I S C C O U P L I N G S The main difference between the metal diaphragm coupling and the typical metal-flexing disc coupling is that a number of discs replace the single diaphragm between the hubs and the spacer. Figure 11-41 shows a schematic of this type of coupling. A typical metal-flexing disc coupling consists of two hubs rigidly attached by interference fit or flange bolting to the driving and driven shaft of the connected equipment. Laminated disc sets are attached to each hub to compensate for the misalignment. A spacer spans the gap between the shafts and is attached to the flexing elements at each end. The functional requirements and characteristics of the flexing elements are to transmit rated torque as well as any system overloads without buckling or permanent deformation. In other words, they must possess torsional rigidity. However, under conditions of parallel, angular, and axial misalignment, the flexing element must have sufficient flexibility

| 479 |

| Centrifugal Compressors | to accommodate these conditions without imposing excessive forces and moments on equipment shafts and bearings. Both of the previous requirements must be met while maintaining stress levels, which are safely within the fatigue limit of the flexing material. Metal-flexing couplings have been known to exhibit occasional large amplitude vibrations in the axial direction when excited at the natural frequency of the coupling. The amount of damping present in a metal-flexing coupling is thought to be relatively small, although it is known to be greater for the laminated disc-type construction than for a coupling consisting of a single piece membrane. The reason for the greater damping in the laminate disc configuration is that, under conditions of axial movement, a microscopic amount of motion takes place between adjacent lamina, as shown in Figure 11-42. Since the element is clamped together under a bolt preload, there is a frictional force, which resists sliding.

Figure 11-41: Typical metal-flexing disc coupling.

Figure 11-42: Frictional damping in a metal-disc coupling.

| 480 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | Field experience by manufacturers and users of turbomachinery has shown that resonant axial vibration of a metal-flexing coupling can at times cause problems, which are reflected through the entire drive train. With laminated disc couplings, the problem occurs only when an external forcing function exists. This condition could be a result of aerodynamic or hydraulic fluctuations in the machine train, out-of-square thrust collars, gearing inaccuracies, or electrical excitations of motor-driven equipment. It is usually possible to avoid operating the couplings at or near resonance if the condition is anticipated during the system design stage. However, such problems do not always occur until after a machine is in service. More information is needed on the nature and magnitude of external excitations. L U B R I C AT I O N O I L S Y S T E M A typical lubrication oil system is shown in Figure 11-43. Oil is stored in a reservoir to feed the pumps and is then cooled, filtered, distributed to the end users, and returned to the reservoir. The reservoir can be heated for startup purposes and is provided with a local temperature indicator, a high-temperature alarm and high/low level alarm in the control room, a sight glass, and a controlled dry nitrogen purge blanket to minimize moisture intake. The reservoir shown in Figure 11-44 should be separate from the equipment base plate and sealed against the entrance of dirt and water. The bottom should be sloped to the low drain point, and the return oil lines should enter the reservoir away from the oil pump suction to avoid disturbances of the pump suction. The working capacity should be at least five minutes, based on normal flow. Reservoir retention time should be 10 minutes, based on normal flow and total volume below minimum operating level. Heating for the oil should also be provided. If thermostatically controlled, electrical immersion heating is provided, the maximum watt density should be 15 watts per square inch. When steam heating is used, the heating element should be external to the reservoir. The rundown level, which is the highest level the oil in the reservoir may reach during system idleness, is computed by considering the oil contained in all components, bearing and seal housings, control elements, and furnished piping that drains back to the reservoir. The rundown capacity should also include a 10% minimum allowance for the interconnecting piping. The capacity between the minimum and the maximum operating levels in an oil system that discharges seal oil from the unit should be

| 481 |

| Centrifugal Compressors | enough for a minimum operation of three days with no oil being added to the reservoir. The free surface should be a minimum of .25 sq ft/gpm of normal flow. The reservoir interior should be smooth to avoid pockets and provide an unbroken finish for any interior protection. Reservoir wall-to-top junctions may be welded from the outside by utilizing full penetration welds. Each reservoir compartment should be provided with two 3⁄4 in. (19mm) minimum size, plugged connections above the rundown oil level. These connections may be used for such services as purge gas, makeup, oil supply, and clarifier return. One connection should be strategically located to ensure an effective sweep of purge gas toward the vents.

Figure 11-43: A typical lube oil system.

| 482 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | The oil system should be equipped with a main oil pump, a standby and, for critical machines, an emergency pump. Each pump must have its own driver, and check valves must be installed on each pump discharge to prevent reverse flow through idle pumps. The pump capacity of the main and standby pumps should be 10–15% greater than maximum system usage. The pumps should be provided with different prime movers. The main pumps are usually steam turbine driven with an electric motor driven backup pump. A small mechanical-drive turbine is highly reliable as long as it is running, but it is undependable for starting automatically after long idle periods. A motor is thus the preferred backup pump driver. A “ready-to-run” status light is usually provided for the motor in the control room to give visible evidence that the electrical circuit is viable. Starting of the backup pump is initiated by multiple and redundant

Figure 11-44: Lube oil reservoir.

| 483 |

| Centrifugal Compressors | sources. The turbine drivers should be maintained for failure by either low-speed or low-steam chest pressure or both. Low oil pressure switches are provided on the pumps and discharge header ahead of the coolers and filters, sometimes after the cooler and filters, and always at the end of the line where the reduced oil pressure feeds the various users. A signal from any of these should start the motor-driven pump, and all alarms should be activated in the control room. The emergency oil pump can be driven with an AC motor but from a power source that is different from the standby pump. When DC power is available, DC electric motors can also be used. Process gas or air-driven turbines and quick-start steam turbines are often used to drive the emergency pumps. The pump capacities for lube and control oil systems should be based upon the particular system’s maximum usage (including transients) plus a minimum of 15%. The pump capacity for a seal oil system should be based upon the system’s maximum usage plus either 10 gpm or 20%, whichever is greater. Maximum system usage should include allowance for normal wear. Check valves should be provided on each pump discharge to prevent reverse flow through the idle pump. The pumps can be either centrifugal or positive displacement types. The centrifugal pumps should have a head curve continuously rising toward the shut-off point. The standby pump should be piped into the system in a manner that permits checking of the pump while the main pump is in operation. To achieve this, a restriction orifice is required with a test bleeder valve piped to the return oil line or the reservoir. Twin oil coolers, Figure 11-45, should be provided and piped in parallel using a single multiport transflow valve to direct the oil flow to the coolers. The water should be on the tube side, and the oil should be on the shell side. The oil-side pressure should be greater than the waterside pressure. This ratio is no assurance that water will not enter the system in the event of a tube leak, but it does reduce the risk. The oil system should be equipped with twin full flow oil filters located downstream from the oil coolers. Since the filters are located downstream from the oil coolers, only one multiport transflow valve is required to direct the oil flow to the cooler filter combinations. Do not pipe the filters and coolers with separate inlet and outlet block valves. Separate block valves can cause loss of oil flow due to the possible human error of flow blockage during a filter switching operation. Filtration should be 10 microns nominal. For hydrocarbon and synthetic oils, the pressure drop for clean filters should not exceed 5 psi at 100°F operating temperature at normal flow. Cartridges will have a minimum collapsing differential pressure of 50 psi (3.4 Bar).

| 484 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | The system should have an accumulator to maintain sufficient oil pressure while the standby pump accelerates from an idle condition. An accumulator becomes a must if a steam turbine drives the standby pump. Overhead tanks are specified by many users to assure flow to critical machinery components. The sizing of the tanks varies depending on the application. In some gas turbine applications, the bearings reach maximum temperature as long as 20 minutes after shutdown. The oil coolers and filters are controlled by a local temperature control loop with remote control room indication and high/low alarm. The coolers and filters also have a differential pressure alarm. These usually feed into a common high alarm to prewarn a need for switching and filter element replacement. To ensure the required constant pressure, a local pressure control loop is provided on each system for turbine lube oil, compressor lube oil, and control oil. Each oil pressure system should be recorded in the control room to provide troubleshooting information. The success of the oil

Figure 11-45: Cooler-filter arrangements

| 485 |

| Centrifugal Compressors | system depends upon not only the instrumentation but upon proper instrument location. The minimum alarms and trips recommended for each major driver and driven machine should be a low oil pressure alarm, a low oil pressure trip (at some point lower than the alarm point), a low oil level alarm (reservoir), a high oil filter differential pressure alarm, a high bearing metal temperature alarm, and a metal chip detector. Each pressure and temperature-sensing switch should be in separate housings. The switch type should be single-pole, double-throw, furnished as “open” (de-energized) to alarm, and “close” (energize) to trip. The pressure switches for alarms should be installed with a “T” connection pressure gauge and bleeder valve for testing the alarm. Thermometers should be mounted in the oil piping to measure the oil at the outlet of each radial and thrust bearing and into and out of the coolers. It is also advisable to measure bearing metal temperatures. Pressure gauges should be provided at the discharge of the pumps, the bearing header, the control oil line, and the seal oil system. Each atmospheric oil drain line should be equipped with steel, nonrestrictive, bull’seye-type flow indicators positioned for viewing through the side. View ports in oil lines can be very useful in providing a visual check for oil contamination. In the piping arrangement and layout, it is very important to eliminate air pockets and trash collectors. Before starting a new or modified oil system, every foot of the entire system right up to the final connection at the machine should be methodically cleaned, flushed, drained, refilled, and all instruments thoroughly checked. LUBRICANT SELECTION Good turbomachinery oil must have rust and oxidation inhibitor, good demulsibility and the correct viscosity, and be both non-sludging and form-resistant. Besides lubrication, the oil has to cool bearings and gears, prevent excessive metal-to-metal contact during starts, transmit pressure in control systems, carry away foreign materials, reduce corrosion, and rust, degradation. Gas turbines, especially the more advanced turbines, should use a synthetic oil, since synthetic oils have a high flash point. Mineral oils can be used for steam turbines. It is not uncommon to have two types of oil in a plant. Mineral oil costs much less than the synthetic oil.

| 486 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | The selection of the correct lubricant must begin with the manufacturer. Refer to the operator’s instruction manual for the oil required and the recommended viscosity range. The local environmental conditions should be seriously considered, including exposure to outside element conditions, acid gas, or steam leaks. As a general rule, most turbomachines are lubricated with premium-quality, turbine-grade oil. However, under certain environmental conditions, it may be advantageous to consider another oil. For example, if a machine is subject to exposure to low concentrations of chlorine or anhydrous hydrochloric acid gases, it may be better to select another oil that will outperform the premium turbine oil. Good results have been recorded using oil containing alkaline additives. Certain automotive or diesel engine oils contain the optimum amount and type of alkaline additives to protect the base oil from reaction with chlorine and HCL. In services where the attack on the lubricant by the gas is unknown, laboratory tests are suggested.

| 487 |

| Centrifugal Compressors | REFERENCES Alford, J. S. “Protecting Turbomachinery from Self-Excited Rotor Whirl.” Journal of Engineering for Power, ASME Transactions, October, 1965, pp. 333-344. Boyce, M. P., Morgan, E., and White, G. “Simulation of Rotor Dynamics of High- Speed Rotating Machinery.” Proceedings of the First International Conference in Centrifugal Compressor Technology. Madras, India, 1978, pp. 6-32. Ehrich, F. F. “Identification and Avoidance of Instabilities and SelfExcited Vibrations in Rotating Machinery.” Adopted from ASME Paper 72De-21. General Electric Co., Aircraft Engine Group, Group Engineering Division, May 11, 1972. Gunter, E. J., Jr. “Rotor Bearing Stability.” Proceedings of the 1st Turbomachinery Symposium, Texas A&M University, October, 1972, pp. 119-141. Lund, J. W. “Stability and Damped Critical Speeds of a Flexible Rotor in Fluid-Film Bearings.” ASME Paper No. 73-DET-103. Newkirk, B. L. “Shaft Whipping.” General Electric Review, Vol. 27, (1924), p. 169. Prohl, M.A. “General Method of Calculating Critical Speeds of Flexible Rotors.” Trans. ASME, J. Appl. Mech., Vol. 12, No. 3, Sept. (1945), pp. A142A148. Reiger, D. “The Whirling of Shafts.” Engineer, London, Vol. 158, (1934), pp. 216-228. Thomson, W. T. Mechanical Vibrations, 2nd edition. Prentice-Hall, Inc., Englewood Cliffs, N.J., 1961.

Bearings Abramovitz, S. “Fluid Film Bearings, Fundamentals and Design Criteria and Pitfalls.” Proceedings of the 6th Turbomachinery Symposium, December, 1977, pp. 189-204. Herbage, B. “High Efficiency Fluid Film Thrust Bearings for Turbomachinery.” 6th Proceedings of the Turbomachinery Symposium, Texas A&M University, December, 1977, pp. 33-38. Herbage, B.S. “High Speed Journal and Thrust Bearing Design.” Proceedings of the First Turbomachinery Symposium, Texas A&M University, October, 1972, pp. 56-61. King, T. L. and Capitao, J. W. “Impact on Recent Tilting Pad Thrust Bearing Tests on Steam Turbine Design and Performance.” Proceedings of

| 488 |

| Rotor Dynamics, Bearings, Lubrication Couplings, and Gears | the 4th Turbomachinery Symposium, Texas A&M University, October, 1975, pp. 1-8. Leopard, A. J. “Principles of Fluid Film Bearing Design and Application.” Proceedings of the 6th Turbomachinery Symposium, Texas A&M University, December, 1977, pp. 207-230. Reynolds, O. “Theory of Lubrication, Part I,” Trans. Royal Society. London, 1886. Shapiro, W. and Colsher, R. “Dynamic Characteristics of Fluid Film Bearings.” Proceedings of the 6th Turbomachinery Symposium, Texas A&M University, December, 1977, pp. 39-53.

Seals Egli, “The Leakage of Steam through Labyrinth Seals.” Trans. ASME, 1935, pp. 115-122. Shah, Piyush. “Dry Gas Compressor Seals.” Proceedings Seventeenth Turbomachinery Symposium, September 1990

Gears AGMA 211.02. Surface Durability (Pitting) of Helical and Herringbone Gear Teeth. Washington, D.C. 1969. AGMA 390.03, Gear Handbook, Vol. 1, Gear Classification, Materials, and Measuring Methods of Unassernbled Gears. Washington, D.C. 1973. API 613. Special Purpose Gear Units for Refinery Services, 2nd Edition. Washington, D.C. 1977. Partridge, J. R. “High-Speed Gears-Design and Application.” Proceedings of the 6th Turbomachinery Symposium, Texas A&M Univ., Dec. 1977, pp. 133-142. Phinney, J. M., “Selection and Application of High-Speed Gear Drives.” Proceedings of the First Turbomachinery Symposium, Texas A&M Univ., Oct. 1972, pp. 62-66.

Couplings Bloch, H. P. “Less Costly Turboequipment Uprates Through Optimized Coupling Section.” Proceedings of the 4th Turbomachinery Symposium. Texas A&M Univ., 1975, pp. 149-152. Calistrat, M. M. “Metal Diaphragm Coupling Performance.” Proceedings of the 5th Turbomachinery Symposium, Texas A&M Univ., Oct. 1976, pp. 117-123.

| 489 |

| Centrifugal Compressors | Calistrat, M. M. “Gear Coupling Lubrication.” American Society of Lubrication Engineers. 1974. Calistrat, M. M. “Grease Separation Under Centrifugal Forces.” ASME. Pub. 75-PTG-3. Calistrat, M. M., and Webb, S. G. “Sludge Accumulation in Continuously Lubricated Couplings.” ASME. 1972. Calistrat, M. M. and Leaseburge, G. G. “Torsional Stiffness of Interference Fit Connections.” ASME. Pub. 72-PTG-37. “Contoured Diaphragm Couplings,” Technical Bulletin, Bendix Fluid Power Corp. Kramer, K. “New Coupling Applications or Applications of New Coupling Designs.” Proceedings of the 2nd Turbomachinery Symposium. Texas A&M Univ., Oct. 1973, pp. 103-115. Timoshenko, S. Strength of Materials: Advanced Theory & Problems. 3rd ed. Van Nostrand Reinhold Pub. 1956. Webb, S. G. and Calistrat, M. M. “Flexible Couplings.” 2nd Symposium on Compressor Train Reliability, Manufacturing Chemists Assn., April 1972. Wright, J., “Which Shaft Coupling is Best-Lubricated or NonLubricated?” Hydrocarbon Processing. April 1975, pp. 191-196. Wright, J. “Which Flexible Coupling?” Power Transmission & Bearing Handbook. Industrial Publishing Co. 1971. Wright, J. “A Practical Solution to Transient Torsional Vibration in Synchronous Motor Drive Systems.” ASME. Pub. 75-DE-15.

Lubrication API Std 614. Lubrication, Shaft-Sealing, and Control-Oil Systems and Auxiliaries for Petroleum, Chemical and Gas Industry Services, 4th Edition, April 1999 Clapp, A. M. “Fundamentals of Lubricating Relating to Operating and Maintenance of Turbomachinery.” Proceedings of the lst. Turbomachinery Symposium. Texas A&M Univ. 1972. Fuller, D. D. Theory & Practice of Lubrication for Engineers. Wiley Inter-science. 1956.

| 490 |

12

Instrumentation Controls

The traditional concept of maintenance in the petrochemical industry has been undergoing a major change to ensure that equipment not only has the best availability but also is operating at its maximum efficiency. Throughout the world in the petrochemical industry, there is a consistent trend to improve maintenance strategy from fix-as-fail to total performancebased planned maintenance. In practice, this calls for on-line monitoring and condition management of all major equipment in the plant. To reach the Utopian goal of just-in-time maintenance with minor disruption in the operation of the plant requires a very close understanding of the thermodynamic and mechanical aspects of plant equipment. The introduction of the total maintenance condition monitoring system means the use of composite condition monitoring systems, which combine both mechanical and performance based analysis. Numerous case studies have shown that many turbomachinery operational problems can only be diagnosed and resolved by correlating the representative performance parameters with mechanical parameters. Major petrochemical complexes contain various types of large machinery. Examples include gas and steam turbines, pumps, and compressors, but also their effect on heat exchangers, distillation towers, and other major plant equipment. Thus, the logical trend in condition monitoring is to multi-machine train monitoring. To accomplish this goal, an extensive database that contains data from all machine trains, along

| 491 |

| Centrifugal Compressors | with many composite multi-machine analysis algorithms, are implemented in a systematic and modular form in a central system. Implementation of advanced performance degradation models necessitate the inclusion of advanced instrumentation and sensors, such as pyrometers for monitoring hot section components and dynamic pressure transducers for detection of surge and other flow instabilities, such as combustion stability in the new, Low NOx combustors. Using expert systems is a necessity to determine faults and life cycles of various components. To fully round out a condition monitoring system, the benefits of total performance based planned maintenance not only ensure the best and lowest cost maintenance program but also that the plant is operated at its most efficient point. An important supplementary effect is that the plant will be operating consistently within its environmental constraints. The new purchasing mantra for chemical-processing plants is “Life Cycle Cost,” and, to properly ensure that this is achieved, a “Total Performance Condition Monitoring” strategy is unsurpassed. CONTROL SYSTEMS Control systems can be an open loop or closed-loop system. The openloop system positions the manipulated variable either manually or on a programmed basis, without using any process measurements. A closed loop

Figure 12-1: Forward and feedback control loop.

| 492 |

| Instrumentation Controls | control system is one that receives one or more measured process variables and then uses it to move the manipulated variable to control a device. Most combined cycle power plants have a closed-loop control system. Closed-loop systems include either a feedback or feed-forward control loop, or both, to control the plant. In a feedback control loop, the controlled variable is compared to a set point. The difference between the controlled variable to the set point is the deviation for the controller to act on to minimize the deviation. A feed-forward control system uses the measured load or set point to position the manipulated variable in such a manner to minimize any resulting deviation. In many cases, the feed-forward control is usually combined with a feedback system to eliminate any offset resulting from inaccurate measurements and calculations. The feedback controller can either bias or multiply the feed-forward calculation. A controller has tuning parameters related to proportional, integrated, derivative, lag, dead time, and sampling functions. A negative control loop will oscillate if the controller gain is too high, but if it is too low, it will be ineffective. The controller must be properly related to the process parameters to ensure closed-loop stability while still providing effective control. This is accomplished first by the proper selection of control modes to satisfy the requirements of the process and second by the appropriate tuning of those modes. Figure 12-1 shows a typical block diagram for feed forward and feedback control. Computers have been used in the new systems to replace analog Proportional Integraf Derivative (PID) controllers, either by setting set points (or lower level set points in supervisory control) or by driving valves in direct digital control. Single station digital controllers perform PID control in one or two loops, including computing functions such as mathematical operations, with digital logic and alarms. D-CS provides all the functions, with the digital proc shared among many control loops. A high level computer may be introduced to provide condition monitoring, optimization, and maintenance scheduling. S TA R T- U P P R O C E D U R E S OF CENTRIFUGAL COMPRESSORS A start-up is the series of events leading up to and including the very first time a new centrifugal compressor is started in a new process. After the very first start, some experience will have been gained, and the procedure, which worked as well as the various valve settings used, would

| 493 |

| Centrifugal Compressors | need only to be refined. Judgment and engineering guidance are needed prior to start-up. The first start-up may last from several days to many weeks, depending on the preparation and complexity of the system. Normally the start-up team will initially write up the procedures of the start-up and test procedures as per ASME PTC-10. Make at least the following checks or inspections prior to rolling the equipment for the first time: 1.

Check that all the instrumentation has been fully calibrated as per the ASME Codes.

2.

Check the lube system and flush it with oil. This usually includes circulating the oil through the lube and seal supply lines and returning the oil to the reservoir by way of the drain lines. For this operation, the bearings and seals are either bypassed or removed.

3.

When the lines are clean, clean the oil filters and replace the cartridges; drain the reservoir, clean it, and fill it with the proper oil.

4.

Clean, inspect, and reinstall the bearings and seals.

5.

Commission the lubrication system and adjust pressures and temperatures to proper levels.

6.

Adjust and check all lubrication and seal oil protection equipment for both proper operating point, as well as being capable of carrying out proper action. This includes alarms, auxiliary cut-in, and trips.

7.

Check the axial shaft float and place the unit in the center of the axial float. Most of these floats are about 15–20 mils.

8.

Check and adjust all other protection equipment, such as high temperature alarms, low voltage, vibration sensors, and lubrication pumps.

9.

If the drive for the compressor is a steam turbine drive, blow down the steam line, and check the inlet pressure and temperature gauges.

10. If it is a motor drive, check the starter for proper sequence, timing, and voltage; torsional characteristics of the motor must be fully known.

| 494 |

| Instrumentation Controls | 11. If the drive is a gas turbine, check out the fuel pressures and the oil system for the gas turbine, which is usually separate from the driven compressor. Also check the gearbox lubrication system. 12. The driver should be run uncoupled from the compressor to check out overspeed trip (steam or gas turbine), rotation, and general operation. This also provides training for starting up part of the system. 13. Install inlet screens (either temporary or permanent) on all compressor inlets, unless piping can be completely inspected. It has not been uncommon during start-ups to find much debris such as welding slag, rods, and other construction debris. After completing item 12 above, alignment should be checked prior to coupling up compressor. 14. When conditions permit, process centrifugal compressors are normally operated on air prior to operation on process gas. Under this condition both discharge temperature and/or seal operation may be the limiting factor and should be checked prior to attempting an air run. The manufacturer may recommend unit buffering on machines having oil seals. When starting a centrifugal compressor on process, several different methods are available. The method used is a function of the system as well as the driver. Figure 12-2 shows a centrifugal compressor driven by a steam turbine. Therefore, in order to cover specifics, let us go through the start-up of a steam turbine-driven centrifugal compressor system, such as

Figure 12-2: Steam turbine driven centrifugal compressor system.

| 495 |

| Centrifugal Compressors | shown in Figure 12-3. For starting, this system, coolant would be circulated through the cooler, the bypass valve opened, and the unit brought up to operating speed. The normal procedure for bringing a steam turbine unit up to speed is usually governed by the manufacturer’s recommended starting procedure for the steam turbine, and this should be followed. Typically, this involves introducing steam into the turbine until it “breaks loose” and a speed of approximately 500 RPM is established. After warming up, the unit’s speed is gradually increased. Care must be exercised when approaching criticals, and they should be passed through rapidly.

Figure 12-3: Steam turbine driven centrifugal compressor.

Figure 12-4: Gas turbine driven centrifugal compressor.

| 496 |

| Instrumentation Controls | The next system to be examined is the gas turbine driven centrifugal compressor as seen in Figure 12-4. The gas turbine control system is complex and has a number of safety interlocks to ensure the safe startup of the turbine. The start-up speed and temperature acceleration curves, as shown in Figure 12-5, are one such safety measure. If the temperature or speed are not reached in a certain time span from ignition the turbine will be shut down. During start-up, the turbine is on an auxiliary drive, which speeds up the turbine to about 1200–1600 RPM. At that point, the turbine is ignited, and the turbine speed and temperature rise very rapidly. The turbine is usually declutched at speeds around 2400 RPM. If the turbine is a two or three shaft turbine, as is the case with aero-derivative turbines, the power turbine shaft will “break loose” at a speed of about 60% of the rated speed of the turbine. The driven compressor is loaded, as shown in Figure 2; this is a typical pressure ratio characteristic compressor curve. The ability to gradually increase speed is truly a positive feature of the steam turbine and gas turbine, because it provides the operator with time to make checks and adjustments as the speed is increased. The short time in which the motor reaches full speed can place very high torsional stresses on the shaft, and the inability to gradually increase to full speed makes checks and adjustments after start-up impossible.

Figure 12-5: Start-up characteristics of a gas turbine.

| 497 |

| Centrifugal Compressors |

Figure 12-6: Compressor characteristics showing the start-up characteristics.

Figure 12-6 shows the typical pressure ratio characteristics as might be experienced as the steam turbine or gas turbine is brought up to speed. The start-up characteristic curve “A” is the line that has the least resistance for start-up purposes and ensures that the torque during the start-up is minimal. Several comments can be made relative to this starting method.

| 498 |

•

By providing a large bypass or recirculation system around the compressor, the pressure ratio across the unit can be kept low. Generally, the pressure ratio should not be so low as to allow a pressure drop to occur across any stage.

•

Low-pressure ratio, and the corresponding high flow, mean very little chance for surge.

•

Actual stonewall flow (shown as dashed lines) is very hard to predict because stonewall is not only a function of the

| Instrumentation Controls | compressor’s design but also a function of the gas being compressed. Therefore, the amount of flow at low-pressure ratios with a given unit may vary considerably from the values shown. •

The driven compressor after the steam or gas turbine has reached full speed can be loaded, and the loading goes up to the design pressure along the constant speed line.

The power required by the compressor lies below the turbine torque curves, as shown in Figure 12-7, and curve up to approximately 96% speed. Thus, instead of reaching the 100% speed characteristic at Point A, the unit reaches only 96% speed. This is no problem for our simple system because, as we close the bypass valve, the discharge pressure will increase and the power required will drop off toward Point D, thus, allowing the speed to reach 100% before Point D is reached. At some point prior to reaching Point D, the check valve will open allowing flow into the process. Further closing of the bypass will reduce the flow through the compressor until Point D is reached. When the bypass valve is completely closed, the full compressor flow is going to process.

Figure 12-7: A turbine torque characteristic curve.

| 499 |

| Centrifugal Compressors |

Figure 12-8: Typical simple motor-driven centrifugal compressor unit.

Figure 12-9: Schematic of electric turbine-driven centrifugal compressor system.

| 500 |

| Instrumentation Controls | For the starting sequence discussed, it was assumed that the compressor inlet pressure temperature and molecular weight were at design. If this is not the case, these changes should be factored into both the volume vs. pressure ratio as well as speed vs. torque curves used in preparing the start-up plan. Another type of system commonly encountered is the motor driven centrifugal compressor unit. Unlike the steam turbine, or gas turbine unit, the motor driven unit usually has lower starting torque capabilities due to the need for limiting current inrush and heat build up. When a motor is starting, it draws four to seven times its full load current. This depresses the voltage, particularly for large motors. Thus, depressing the starting torque by the square of the voltage drop. If the voltage drops to 80%, the torque drops to 64% (80%2=64%). Normal speed up time for a motor is limited to less than 20 seconds. Figure 12-8 shows a typical, simple motordriven centrifugal compressor unit. Figure 12-9 is a schematic of the motor driven turbine system. Like the steam turbine-driven unit, we would start coolant circulating in the gas cooler and open the bypass valve prior to starting. Here the similarity ends. In order to minimize current inrush, the inlet butterfly would be almost closed, and the compressor characteristic is changed, as seen in Figure 12-10. If we now push the start button, the machine should come to 100% speed within approximately l5 seconds. With the compressor now operating at 100% speed, the butterfly valve can be slowly opened to increase the compressor flow. Then, the bypass valve can be slowly closed to increase the discharge pressure. Normally, the closing of the bypass valve and opening of the inlet butterfly valve can be shown as a series of oscillations back and forth on the 100% speed line at about Point D. The discharge pressure will continue to increase until the check valve opens and gas starts flowing to the process. The bypass valve can then be closed and compressor load regulated by the inlet valve. In Figure 12-11, one can see the large difference in torque between that required by the compressor and that developed by the motor. This difference in torque provides for the acceleration of the system. Thus, the smaller the difference, the longer the starting time. With the two approaches to starting a centrifugal compressor having been discussed at some length, one should be ready for more complex systems. While we will not go into them here, several things should be noted:

| 501 |

| Centrifugal Compressors |

Figure 12-10: A typical compressor speed-torque characteristic curve.

Figure 12-11: A typical motor vs. driven compressor speed-torque characteristic curve.

| 502 |

| Instrumentation Controls | •

Probably the most important thing to define, or attempt to define, is the condition that prevails at the end of the speed increasing cycle of Points A and B. Once this is defined, along with the various pressures and temperatures, the approach to starting will take form.

•

Below 30% speed, very little head or pressure ratio is developed across the compressor. The flow is also pretty low. Thus, when warming up the steam turbine at less than 30% speed, flow control or cooling of the compressor flow is seldom required.

•

On refrigeration units, the discharge flow normally goes to a condenser where the gas is condensed into liquid at a fixed pressure and temperature. In order to get the machine on stream, three things have to happen: –

Inlet pressure must be kept low to provide low flash temperature.

–

At the same time, suction temperatures must be kept low so that pressure ratios can be realized.

–

Discharge pressure must be reached so condensing starts.

•

Thus, for a typical refrigeration system, once 30% speed is reached, the speed should be increased rapidly, and, at the same time, the suction drums should be quenched with liquid in order to provide cooling. At the same time, gas may be blown off to keep suction pressures near design. As speed increases and inlet temperature falls, the blow-off can be closed and the gas diverted to the condenser. Bypasses must be controlled to maintain flows between surge and overload so that pressure ratios can be realized as design speed and temperature are approached.

•

On a compressor with side loads, or one that is inter-cooled, it is often a good idea to establish flows and pressures across compressor sections starting at the suction and working toward the discharge.

| 503 |

| Centrifugal Compressors | FACTORS AFFECTING THE CONTROL OF CENTRIFUGAL COMPRESSORS The type of control for centrifugal compressor usually depends on the type of driver is used. Turbine driven compressors usually control the compressor by varying the speed. This type of control permits a wide range of operation in an efficient manner. Speed control is more efficient than throttling the flow at a constant compressor speed. This is due to the fact that throttling creates artificial resistant, and an unrecoverable loss in power results. In motor-driven compressors, the control is much more complex, since, in most cases, the motors are constant speed. Thus, the following controls are needed for electric motor drives: Butterfly valve at the compressor inlet or discharge. Throttling at the inlet is preferred, because it reduces the gas density. Throttling at the exit does not take advantage of the reduction of the head, nor the density reduction obtained by inlet throttling, which results in lowering the power requirements. Adjustable inlet guide vane. This adjusts the characteristics of a centrifugal compressor. Adjustable guide vanes change the aerodynamic design by creating a pre-whirl. Effects of pre-whirl are fully explained in chapter 4.

MODES OF OPERATION (STARTUP, ON-STREAM, SHUTDOWN) The complete antisurge control system must effectively prevent surge during startup and shutdown, as well as during on-stream operation. Discrete signals are used to override the analog control system and manipulate valves during startup and shutdown. As when designing the analog system, it is equally important to properly evaluate the compressor and process requirements when designing the startup and shutdown logic system. S TA R T U P C O N T R O L A C T I O N For startup, it has been common practice to maintain the antisurge control valve fully open during most of the acceleration period. Typically,

| 504 |

| Instrumentation Controls | the logic override signal is removed at 80 or 85% speed to transfer control to the system for on-stream operation. This introduces a step control change. The immediate action by the antisurge controller is an attempt to drive the valve fully closed. The valve, moving toward its closed position, brings the compressor operation to the point where the antisurge controller must take over control and modulate the valve in a partially open position for stable compressor operation. If this recovery action is not fast enough, the valve will close too far, and the compressor will surge. If this is a problem, the system must include a feature to smooth the transfer action and to give the controller sufficient time to recover control. Instead of activating the antisurge controller at 80 or 85% speed, it is better to have it in control during the entire acceleration period whenever possible. To facilitate this, the surge control line must never intersect the surge line. The control line is always on the right side of the surge line. The control system should be in control throughout the startup, then a transfer at 80 or 85% speed would not be necessary. In most cases, a startup override to open the valve at low speed would be necessary. Another factor that must be considered is the fact that the stages of a multistage compressor become mismatched when operated at less than normal speed. Under some conditions, the compressor could surge even though the anti-surge valve is fully open. Usually, surge is initiated by a stage near the discharge when the machine is operating at rated speed. At lower speeds, a stage closer to the inlet, frequently the first stage, initiates surge. The reason for this is that at lower speeds, the gas volume is not reduced by each stage as much as it is reduced at rated speed. Consequently, at reduced speed, the inlet volumetric flow rate through the final stages is greater than their rated flow, and the volume entering stages near the inlet tends to become less than their rated flow. At high flow rates, the gas velocity can reach the speed of sound at some location in the compressor. Since the maximum velocity of any gas is limited by its sonic velocity, this imposes a maximum flow limit on the compressor. This condition seldom occurs while operating at rated speed, however, when the speed of a multi-stage compressor is reduced, a speed finally is reached where it is no longer possible to keep the compressor out of surge by recycling or venting gas from the final discharge. The reason is that some stage near the discharge has reached its choked (sonic) flow limit, which backs up the flow in the earlier stages to the point where one stage reaches its surge point. To further explain what happens, assume a six-stage compressor with a rated inlet volume flow of 10,000 acfm. Also assume that, at rated speed, each stage makes a

| 505 |

| Centrifugal Compressors | compression ratio of 2.0, and surge occurs at approximately 7700 acfm at 100% speed. Disregarding the effect of increased temperature, the discharge volume from Stage 1 therefore is 5000, which is the inlet volume to Stage 2. If we carry this on through, we find the inlet volume flows to Stages 3 through 6 to be 2500, 1250, 625, and 313, respectively, as shown in Figure 12-12. Now let us assume that, at 70% speed, each stage makes a compression ratio of only 1.5 and that Stage 6 reaches its choke limit at a volume flow of 600. Also assume that each stage will surge at 60% of its rated capacity when operating at 70% speed. The 1.5 volume reduction per stage means that each stage will have an inlet volume flow 1.5 times greater than its discharge volume. If we start with the 600 maximum (choked) limit of Stage 6 and multiply back through all stages, we determine that the volume flows into Stages 5, 4, 3, 2 and 1 are 900, 1350, 2025, 3038 and 4556, respectively, as shown in Figure 12-13. The 4556 volume flow at the inlet to Stage 1 is

Figure 12-12: Volume reduction per stage, compression ratio = 1.5.

| 506 |

| Instrumentation Controls | considerably less than its 6000 minimum surge limit, or, in other words, choked flow in Stage 6 has forced Stage 1 to operate below its surge limit. This condition might occur at a speed in the vicinity of 80% to 90% on some compressors and at much lower speeds on others. In order to accommodate this condition, a recycle valve at some intermediate stage must be open while the compressor is operating at reduced speed during startup or shutdown. The type of gas handled is a factor in the condition just described. For a heavy gas, e.g., carbon dioxide, several valves located at intermediate stages may be required because its sonic velocity is higher. Also, a change in speed makes a greater difference in the volume reduction produced by each stage. Stage mismatching also can alter the surge line shape. The shape can change with speed, as well as gas variations. The surge line for a lighter gas is slightly curved, but it curves in the reverse direction from the surge

Figure 12-13: Volume reduction per stage, compression ratio = 2.0.

| 507 |

| Centrifugal Compressors | line for a heavy gas. For the same compression ratio, an axial flow compressor has more stages than a centrifugal, so the surge initiating stage changes more times. This accounts for the reverse curvature typical of axial flow compressors. For a constant speed, axial flow machine the effect is less than for a variable speed axial, because only the stages with variable angle stators are affected. Typically, approximately 40% of the stages have variable stator vanes.

SHUTDOWN CONTROL ACTION A very high potential for surge exists during an emergency shutdown, or any time the compressor is tripped while operating at normal conditions. Deceleration is very rapid, typically with speeds having decreased to 80% rpm approximately 1 to 11⁄2 seconds after trip-out. The compressor’s ability to produce pressure rise decreases approximately proportionally to speed squared. Thus, the need exists to lower the discharge pressure very quickly after trip-out; otherwise, severe surging can result. During deceleration, the recycle valve must handle all of the compressor flow, plus any excess gas accumulation in the discharge system. If there is an excessive quantity of gas compressed in the discharge system, the recycle valve must be capable of passing a very high flow rate during the first several seconds of the deceleration period in order to lower the pressure fast enough to prevent surge. The most important factor in preventing surge under these conditions is to have the least possible volume at the discharge. Sometimes this can be accomplished by locating a check valve as close as possible to the compressor discharge. However, in many systems the recycle takeoff point must be located downstream of a cooler and separator; this adds a considerable volume at the compressor discharge. In some cases it is possible to accommodate this by a fast opening (1 second or less), generously sized (125 to 200% of compressor rated flow) recycle valve. In other cases, it may be necessary to use an additional valve and check valve very close to the compressor discharge. If possible, this valve can vent gas to atmosphere. If atmospheric venting is not feasible, then the extra shutdown valve could dump hot gas back to the compressor inlet. The temperature rise from inlet to discharge decreases rapidly as speed decreases, so recycling hot gas back to inlet normally is not a problem during the relatively short coast-down time. It should also be noted that the high temperature effect of recycling hot discharge gas is often

| 508 |

| Instrumentation Controls | mitigated by the temperature drop (Joule-Thompson effect) resulting from the expansion across the recycle valve. This effect is significant, especially with natural gas at high pressures. If a temporary high inlet temperature is a problem for a particular shutdown characteristic, then the hot gas recycle valve could be closed approximately 10 to 15 seconds after tripout. For the rest of the coast-down period, the deceleration rate normally is slow enough that the norma1 recycle valve is able to prevent surge.

TOTAL CONDITION MONITORING SYSTEM A total condition monitoring system must be designed to provide the operators and rotating equipment engineers with clear insight into machinery performance problems: •

Enhance predictive maintenance capability by diagnostic tools

•

Plant operation with minimum degradation based on optimal washing of the train

•

Integrated condition monitoring utilizing field-proven hardware and software

•

Provide voltage free contacts at the monitors for machine safeguarding

•

Data link to and from the plant DCS system so as to upgrade performance curves in the DCS system, which control plant processes

A total condition monitoring system for a major petrochemical installation requires a fully integrated system with an extensive database to ensure that it can achieve the following goals: •

High machinery availability

•

Maintaining peak efficiency and limiting performance degradation of machine trains

•

Extending time between inspections and overhauls

•

Optimizing the cycle configuration

•

Estimating availability

•

Evaluating scenarios by means of “What If” analysis

| 509 |

| Centrifugal Compressors | •

Estimating maintenance requirements and life of hot section components

•

Fault identification by expert system modules

A condition monitoring system designed to meet these needs must consist of hardware and software designed by engineers with experience in machinery and energy system design, operation, and maintenance. Each system needs to be carefully tailored to individual plant and machinery requirements. The systems must obtain real time data from the plant DCS, and, if required, from the gas and steam turbine control systems. Dynamic vibration data is taken in from the existing vibration analysis system into a data acquisition system. The system can consist of several high-performance, networked computers—depending on plant size and layout. The data must be presented using a graphic user interface (GUI) and include the following: Aerothermal Analysis. This pertains to a detailed thermodynamic analysis of the full power plant and individual components. Models are created of individual components including the gas turbine, steam turbine heat exchangers, and distillation towers. Both the algorithmic and statistical approaches are used. Data is presented in a variety of performance maps, bar charts, summary charts, and baseline plots. Combustion Analysis. This includes the use of pyrometers to detect metal temperatures of both stationary and rotating components such as turbine blades. The use of dynamic pressure transducers to detect flame instabilities in the combustor especially in low NOx applications Vibration Analysis. This includes an on-line analysis of the vibration signals, FFT spectral analysis, transient analysis, and diagnostics. A wide variety of displays are available including orbits, cascades, Bode and Nyquist Plots, and transient plots. Mechanical Analysis. This includes detailed analysis of the bearing temperatures, lube, and seal oil systems, and other mechanical subsystems. Diagnosis. This includes several levels of machinery diagnosis assistance available via expert systems. These systems must integrate both mechanical and aerothermal diagnostics. Trending and Prognosis. This includes sophisticated trending and prognostic software. These programs must provide users information to clearly understand underlying causes of operating problems. “What-If” Analysis. This program should allow the user to do various studies of plant operating scenarios to ascertain the expected

| 510 |

| Instrumentation Controls | performance level of the plant due to environmental and other operational conditions.

MONITORING SOFTWARE The monitoring software for every system will be different. However, all software is there to achieve one goal, that is, it must gather data, ensure that it is correct, and then analyze and diagnose the data. Presentations must be in a convenient form and should be easily understood by plant operational personnel. All priorities must be given to the data collection process. This process must not in any manner be hampered since it is the cornerstone of the whole system. A convenient framework within which to categorize the software could be as follows: Graphic User Interface (GUI). This consists of screens that would enable the operator to easily interrogate the system and to visually see where the instruments are installed and their values at any point of time. By carefully designed screens, the operator will be able to view at a glance the relative positions of all values and, thus, fully understand the operation of the machinery. Alarm/System Logs. To fully understand a machine, we have to have various types of alarms. The following are some of the suggested types of alarms: •

Instrument Alarms: These alarms are based on the instrumentation range.

•

Value Range Alarm: These alarms are based on operating values of individual points—both measured and calculated points. These alarms should be variable in that they would change with operating conditions.

•

Rate of Change Alarm: These alarms must be based on any rapid change in values in a given time range. This type of alarm is very useful to detect bearing problems, surge problems, and other instabilities.

•

Prognostic Alarms: These alarms must be based on trends and the prognostics based on those trends. It is advisable not to have

| 511 |

| Centrifugal Compressors | prognostics that project in time more than the collected hours of data that is trended. Performance Maps. These are performance maps based on design or initial tests (base lines) of the various machinery parameters. These maps, for example, present how power output varies with ambient conditions, or with properties of the fuel, the condition of the filtration system, or how close to the surge line a compressor is operating. On these maps, the present value is displayed, thus allowing the operator to determine the degradation in performance occurring in the units. Analysis Programs. These include aerothermal and mechanical analysis programs with diagnostics and optimization programs. Aerothermal Analysis. Typical aerothermal performance calculations involve evaluation of component unit power, polytropic and adiabatic head, pressure ratio, temperature ratio, polytropic and adiabatic efficiencies, temperature profiles, and a host of other machine specific conditions under steady state as well as during transients—start ups and shut downs. This program must be tailored to individual machinery and to the instrumentation available. Data must be corrected to a base condition so that it can be compared and trended. The base condition can vary from ISO ambient conditions to design conditions of a compressor or pump, if those conditions are very different from ISO ambient conditions. To analyze off-design operation, it is necessary to transpose values from the operating points back to the design point for comparison of unit degradation. Mechanical Analysis. This program must be tailored to the mechanical properties of the machine train under consideration. It should include bearing analysis, seal analysis, lubrication analysis, rotor dynamics, and vibration analysis. This includes the evaluation and correlation of bearing metal temperatures, shaft orbits, vibration velocity, spectrum snapshots, waterfall plots, stress analysis, and material properties. Diagnostic Analysis. This program can be part of an expert system or consist of an operational matrix that can point to various problems. The program must include comparison of both performance and mechanical health parameters to a machine specific fault matrix to identify if a fault exists. Expert analysis modules can, in many cases, aid to faster fault identification but are usually more difficult to integrate into the system. Optimization Analysis. Optimization programs take into account many variables, such as, deterioration rate, overhaul costs, interest, and

| 512 |

| Instrumentation Controls | utilization rates. These programs may also be dependent on more than one machine train if the process is interrelated between various trains. Life Cycle Analysis. The determination of the effect of the material, the temperature excursions, the number of start-ups and shutdowns, and the type of fuel all relate to the life of hot section components. Historical Data Management. This includes the data acquisition and storage capabilities. Present day prices of storage mediums have been dropping rapidly, and systems with 14–16 gigabyte hard disks are available. These disks could store a minimum of two years of one-minute data for most plants. One-minute data is adequate for most steady state operation, while start-ups and shutdowns or other non-steady state operation should be monitored and stored at an interval of one second. To achieve these time rates, data for steady state operation can be obtained from most plant-wide DCS systems, and, for unsteady state conditions, data can be obtained from control systems. I M P L E M E N TAT I O N O F A CONDITION MONITORING SYSTEM The implementation of a condition monitoring system in a major petrochemical plant requires a great deal of forethought. A major petrochemical plant will have a number of varied, large, rotating equipment. This will consist usually of various types of prime movers such as large electrical motors, gas turbines, steam turbines, hot gas expanders, and diesel or gas reciprocating engines. These prime movers will be driving large compressors. The following are some of the major steps, which need to be taken to ensure a successful system installation. First, decide on what equipment should be monitored on line and what systems should be monitored off-line. This requires an assessment of the equipment in terms of both first cost and operating costs, redundancy, reliability, efficiency, and criticality. Obtain all pertinent data of the equipment to be monitored. This would include details of the mechanical design and the performance design. Some of this information may be difficult to obtain from the manufacturer and will have to be calculated from data being obtained in the field or after installation during commissioning tests in a new installation. Obtaining baseline data is critical in the installation of any condition monitoring system. In most systems, it is the rate of change of parameters that is being trended not the absolute values of these points. It is also

| 513 |

| Centrifugal Compressors | important to decide what type of alarms will be attached to the various points. Rate of change alarms must be for bearing metal temperatures, especially for thrust bearings where temperature changes are critical. Prognostic alarms should be applied to critical points. Alarms randomly applied tend to slow down the system and do not provide added protection. The following are some of the basic data that would be necessary in setting up a system: Type of gasses and fluids used in the various processes. The Equation of State and other thermodynamic relationships that govern these gases and fluids. Type of fuel used in the prime movers. If the fuel analysis is available, include the fuel composition and the heating values of the fuel. Materials used in various hot sections. Includes combustor liners, turbine nozzles, and blades. This includes stress and strain properties as well as Larson-Miller parameters. Performance maps of various critical parameters. These include power and heat consumption as a function of ambient conditions, pressure drop in filters, and the effect of backpressure, compressor surge, efficiency, and head maps.

Location of Instrumentation Determine the instrumentation that exists and the location. Location of the instrumentation from the inlet or exit of the machinery is important so that proper and effective compensation may be provided for the various measured parameters. In some cases, additional instrumentation will be needed. Experience indicates that older plants require 10 to 20% more instrumentation, depending upon the age of the plant.

Typical Instrumentation Location and types of sensors depend on the type of machine under consideration. The following are some of the major instrumentation for a typical gas turbine-centrifugal compressor system: •

| 514 |

Accelerometer –

At machine inlet bearing case, vertical

–

At the machine discharge bearing case, vertical

–

At machine inlet bearing case, axial

| Instrumentation Controls | •

•

•

•

Process pressure –

Pressure drop across filter

–

Pressure at compressor and turbine inlet

–

Pressure at compressor and turbine discharge

Process temperature –

Temperature at compressor and turbine inlet

–

Temperature at compressor and turbine discharge

Machine speed –

Machine speed of all shafts

–

Thrust bearing temperature

–

Thermocouples or resistance temperature elements embedded in front and rear thrust bearing

Desirable Instrumentation (Optional) –

–

Non-contacting eddy-current vibration displacement probe adjacent to: °

Inlet bearing, vertical

°

Inlet bearing, horizontal

°

Discharge bearing, vertical

°

Discharge bearing, horizontal

Non-contacting eddy-current gap-sensing probe adjacent to: °

Forward face of thrust bearing collar

°

Rear face of thrust bearing collar (Note: the noncontacting sensor in its role of measurement of gap DC voltage is sensitive to probe and driver temperature variations. Careful evaluation must be conducted of sensor type, its mounting, and location for this measurement.)

°

Process flow measurement at inlet or discharge of machine

°

Radial bearing temperature thermocouple or resistance temperature element embedded in each bearing, or temperature at lube oil discharge of each bearing.

| 515 |

| Centrifugal Compressors |

Figure 12-14: Instrumentation for monitoring and diagnostics on an industrial type gas turbine driver.

Figure 12-15: Instrumentation for monitoring and diagnostics on a centrifugal compressor.

| 516 |

| Instrumentation Controls | –

Lube oil pressure, temperature, and corrosion probe

–

Dynamic pressure transducer at compressor discharge for indication of flow instability

–

Fuel system (water capacitance probe, corrosion probe, and Btu detector)

–

Exhaust gas analysis

–

Torque measurement

Figures 12-14 and 12-15 show possible instrument locations for an industrial gas turbine and centrifugal compressor. Once the data points have been decided, limits and alarm must be set. This is a long and challenging task, as the limits on many points are not given in the operation manuals. In some cases, the criticality of the equipment may necessitate that the alarm threshold on certain points be lowered to give early warning of any deterioration of the system. It should be noted that, since this is a condition monitoring system, early alarm warnings are desirable in most cases. Types of reports and summary charts should be planned to optimize the data and to present it in the most useful manner to the plant operations and maintenance personnel. Consider the types of DCS and the control systems available in the plant, along with the protocol of these systems and their relationships to the condition monitoring system. The slave or master relationship is important in setting up the protocols. Diagnostics for the system requires that someone note any unusual characteristics of the machinery, especially in older plants, which have a history of operation inspections and overhauls. Costs of operations such as fuel costs, labor costs, down time costs, overhaul hours, and interest rates are necessary in computing parameters such as time of major inspections, off–line cleaning, and overhauls. T Y P I C A L C O N D I T I O N M O N I T O R I N G R E S U LT S Examining a typical petrochemical plant with a gas turbine centrifugal compressor train, a hot gas expander compressor train, a steam turbine compressor train, and large motor-driven pumps, one realizes that we have all the elements of a complex and varied plant machinery system.

| 517 |

| Centrifugal Compressors | GAS TURBINE COMPRESSOR TRAIN The gas turbine driven compressor train is a very common application and is the most complicated configuration to monitor. The gas turbine is very sensitive to temperature excursions and the type of fuel being burnt. It is, therefore, very important to measure the exhaust temperature profile. If there were only one parameter that could be measured, then the exhaust temperature profile would be the one to be measured. The temperature spread between adjacent thermocouples and the maximum spread measured are very important parameters to monitor. The temperature spread, as shown in Figure 12-16, is indicative of the overall health of the turbine.

Figure 12-16: Exhaust temperature profile for a typical gas turbine.

| 518 |

| Instrumentation Controls | The firing temperature of the gas turbine is another very important parameter to monitor. The value of this temperature must be computed since thermocouples are not placed in that region, as thermocouple life would be greatly shortened at those high temperatures. The firing temperature must be calculated using two techniques, 1) based on the fuel energy input and, 2) based on a heat balance between the gasifier part of the turbine and the power turbine. The second method is hard to do in a compressor train because there is no easy way to measure the turbine output power. The most accurate technique would be to insert a torque meter between the compressor and the gas turbine. This, however, is a very costly process, and, in installations, that are already operating, there may be no space to do so. Problems with drift in torque meters have been minimized in the new designs. Computing the power absorbed by the compressor requires sophisticated techniques and an in-depth understanding of the process gas properties and behavior. This understanding is necessary to not only determine the compressor power absorbed but also to study its own characteristics such as the head produced and the surge characteristics of the compressor. Figure 12-17 shows that proper

Figure 12-17: Comparison of computed gas turbine firing gas temperature.

| 519 |

Polytropic Head (ft / lbf / lbm)

| Centrifugal Compressors |

Corrected Horsepower (HP)

Figure 12-18: A typical compressor characteristic curve showing a relationship between the corrected compressor flow vs. the corrected polytropic head.

Figure 12-19: A typical compressor characteristic curve showing a relationship between the corrected compressor flow vs. the corrected power.

| 520 |

Corected Temperature (°F)

| Instrumentation Controls |

Corrected Pressure (psia)

Figure 12-20: A typical compressor characteristic curve showing a relationship between the corrected compressor flow vs. the corrected exit temperature.

Figure 12-21: A typical compressor characteristic curve showing a relationship between the corrected compressor flow vs. the corrected discharge pressure.

| 521 |

| Centrifugal Compressors | computing of these temperatures can give a close relationship between these two techniques for calculating the turbine firing temperature. The driven compressor in the train could be either a centrifugal compressor or an axial flow compressor. Over eighty percent of the installations are centrifugal compressor trains. Centrifugal compressors have a greater operating range than axial flow compressors. The gas turbine’s compressor is usually an axial flow compressor. This, therefore, necessitates an understanding of both types of compressors. Figures 12-18 through Figures 12-21 show typical compressor maps with operating points indicating where the unit was being operated and its relationship with the design point. The surge line shown is where the compressor is operating in a very unstable region. Surge is the reversal of flow inside the flow passages of the compressor. Further analysis could be performed by analyzing the flow in each individual centrifugal impeller. This can be done by noting the impeller geometry, which includes the impeller eye and exit diameters and the inlet and exit blade angles. In this manner, the flow characteristics can be described, and the impeller in the compressor that is close to surge can be identified. The reduction in circulation flow, due to the operation of the unit in a stable but close to surge position, can save in some applications 7 to 15% of the total power, which, in a unit of about 15,000 hp, can translate to a savings of nearly half a million dollars. Monitoring the turbine inlet air filter pressure drop and the pressure ratio across the turbine compressor can maximize the turbine operation. On-line water washes can be very effective for some time but need to have an off line wash to achieve its full potential. The off-line wash is very expensive and, in some cases, a near impossibility as the cost of a shutdown could be prohibitive since the process, once shutdown, could take a couple of days to get back on-line at full power. S P E C T R U M A N A LY S I S Spectrum analysis transforms a displacement/time chart into an amplitude/frequency chart known as a spectrum. This analysis consists of decomposing a time-varying signal into its component pure tones. Pure tones are sinusoidal waveforms of constant frequency and amplitude. This decomposition is done digitally upon a signal by a minicomputer using the Fourier transformation or by filtering the signal. Signals generated by high-speed machinery are very complex in nature and are generated by several forces with a net effect that masks the

| 522 |

| Instrumentation Controls | pure tones. The random portion of the signal, which is blended with the pure tones, is called noise. The ratio of the total amplitude (area under spectrum) to that of the noise is called the signal-to-noise (S/N) ratio. Sometimes this ratio is expressed in decibels or db. The Fourier analyzer is a digital device based on the conversion of time-domain data to a frequency domain by the use of the fast Fourier transform. The fast Fourier transform (FFT) analyzers employ a minicomputer to solve a set of simultaneous equations by matrix methods. Time domains and frequency domains are related through Fourier series and Fourier transforms. By Fourier analysis, a variable expressed as a function of time may be decomposed into a series of oscillatory functions (each with a characteristic frequency), which, when superpositioned or summed at each time, will equal the original expression of the variable. This process is shown graphically in Figure 12-22. Since each of the oscillatory signals has a characteristic frequency, the frequency domain reflects the amplitude of the oscillatory function at that corresponding frequency.

Figure 12-22: Decomposition of a time signal into a sum of oscillatory functions from which a spectrum is obtained.

| 523 |

| Centrifugal Compressors | V I B R AT I O N M E A S U R E M E N T Successful measurement of machine vibration requires more than a transducer (randomly selected and installed), and a piece of wire to carry the signal to the analyzer. When the decision to monitor vibration is made, three choices of measurement are available: (1) displacement, (2) velocity, and (3) acceleration. These three measurement types emphasize different parts of the spectrum. DISPLACEMENT TRANSDUCERS Eddy-current proximity probes are primarily used as displacement transducers. Eddy probes generate an eddy-current field, which is absorbed by a conducting material at a rate proportional to the distance between the probe and the surface. They are often used to sense shaft motion relative to a bearing (by mounting them within the bearing itself) or to measure thrust motions. They are generally indifferent to hostile environments, including temperatures up to 250ºF, and are not expensive. One drawback is that shaft surface conditions and electrical runout can result in false signals. Also, the smallest displacement that can be successfully measured is limited by the S/N ratio of the system. In practice, it is difficult to measure values less than 0.0001 of an inch. If shaft displacement is being measured, the shaft runout (measured with the same pickup) should be less than the smallest measurable value. To achieve the proper shaft runout, it is necessary that the shaft be precision ground, polished, and demagnetized. VELOCITY TRANSDUCERS Velocity transducers are made up of an armature mounted in a magnet. The motion of the armature in the magnet creates a voltage output proportional to the velocity of the armature. Usually, the forces being measured must be relatively great to cause a signal output. However, the signal is quite strong when mounted on the machine bearings, and amplification is usually not needed. They are very rugged but are also large and cost roughly 10 times as much as a proximity probe. Because of damping, transfer function characteristics of the armaturemagnet construction generally limit the low-frequency response to approximately 10 Hz. At the high end of the frequency range, the resonant peak of the pickup itself is the limiting factor. Thus, the useful linear bandwidth is limited. The main advantage of the velocity pickup is that it

| 524 |

| Instrumentation Controls | is a high-output/low-impedance device, and hence, it provides an excellent S/N ratio, even under less than ideal conditions. The major disadvantage of the velocity pickup is its sensitivity to placement. The probe is directional so that, if the same force is applied horizontally or vertically, the probe will give different readings. A C C E L E R AT I O N T R A N S D U C E R S Most accelerometers consist of some small mass mounted on a piezoelectric crystal. A voltage is produced when accelerations acting on the mass create a force acting on the crystal. Accelerometers have a wide frequency response and are not excessively costly. They also are temperature resistant. Accelerometers have two main limitations. First, they are extremely low-output/high-impedance devices requiring loading impedances of at least 1 MW. Such requirements rule out the use of long cables. One solution has been to have an amplifier built into the pickup to provide a low-impedance/amplified signal. A power supply is required, and the weight is increased. The second limitation of this pickup is illustrated by an example. Acceleration of one g at 0.5 Hz represents a displacement of 100 inches. It is obvious that in spite of its wide-band response (sometimes 0.1 Hz–15 kHz), it is severely limited at the low end by a poor S/N ratio. The transducer type used should be matched to the machine being analyzed. Knowledge of the types of problems normally encountered will benefit this selection. For instance, the non-contacting shaft displacement probe helps to correct misalignment and balancing problems but is inappropriate in analyzing gear mesh problems and blade passage frequencies. Also, if signal integration or double integration is to be carried out, the lowpass filters used to attenuate high-frequency spectra also have a highpass filter, which effectively creates a lower frequency limit (often as high as 5 Hz). As mentioned before, one main criterion in deciding which transducer to use is the frequency range to be analyzed. Figure 12-23 shows the frequency limitations placed on the three types of transducers discussed previously. Charts are available to convert from one type of measurement to another, as shown in Figure 12-24. Many of these charts also show approximate vibration limits. The charts demonstrate the independence of velocity measurements relative to frequency, except at very low and very high frequencies where the amplitude limits are constant throughout the operating speed range. These limits are approximate and depend on the type of machinery, casing, foundation, and bearings to determine the final vibration limits.

| 525 |

| Centrifugal Compressors |

Figure 12-23: Limitations on machinery vibration analysis system and transducers.

Figure 12-24: Vibration nomograph and severity charts (courtesy of IRD Mechanalysis, Inc.).

| 526 |

| Instrumentation Controls | LIFE CYCLE COSTS The life cycle costs of any machinery are dependent on the life expectancy of the various components and the efficiency of its operation through out its life. Figure 12-25 shows the cost distribution by the three major categories, initial costs, maintenance costs, and operating or energy costs of a gas turbine centrifugal compressor train over a 20-year life cycle of the train. This figure indicates that the new costs are about 10–15% of the life cycle costs, while maintenance costs are approximately 15–20% of the life cycle costs, and operating costs, which essentially consist of energy costs, make up the remainder between 65-75% of the life cycle costs of most major machinery in a petrochemical plant. It is, therefore, clear why the new purchasing mantra for a petrochemical plant, or for that matter for any major plant operating large machinery, is “Life Cycle Cost”. This brings performance monitoring to the forefront as an essential tool in any type of plant condition monitoring system. The major costs in a life cycle are the cost of energy. Thus, operating the plant as close to its design conditions guarantees that the plant will reduce its operating costs. This can be achieved by ensuring that the turbine compressor is kept clean and that the driven compressor is operating close to its maximum efficiency, which in many cases is close to the surge line. Thus, knowing where the compressor is operating with respect to its surge line is a very critical component in plant operating efficiency. The life expectancy of most hot section parts depends on various parameters and is usually measured in terms of equivalent engine hours. The following are some of the major parameters that affect the equivalent engine hours in most machinery especially gas turbines: •

Type of fuel

•

Firing temperature

•

Materials stress and strain properties

•

Effectiveness of cooling systems

•

Number of starts

•

Number of trips

| 527 |

| Centrifugal Compressors |

Figure 12-25: Life cycle costs of a gas turbine-centrifugal compressor train.

STEAM TURBINE COMPRESSOR TRAIN The steam turbine in a steam turbine driven compressor train could be either a backpressure steam turbine or a condensing steam turbine. The backpressure steam turbines are usually smaller in size than the condensing steam turbines. The backpressure steam turbines are easier to monitor than the condensing steam turbines, which requires the system to monitor the condenser also. The backpressure steam turbine conditions at both the inlet and the exit can be measured accurately and the power computed. The problems encountered in monitoring the condensing steam turbine is the fact that the last stages in the turbine are operating inside the saturation line in a gaseous liquid medium. Thus, knowing the steam temperature does not inform us of the steam properties, as the quality of the steam is still unknown. It is, therefore, necessary to know the cooling water flow to the condenser and the temperature difference of the cooling water flow. The accuracy of these measurements is essential in calculating the steam quality. The steam quality is not only important in determining the performance of a condensing steam turbine but is also important in determining the life of the last blade rows of the condensing turbine, which are affected due to erosion of the blade leading edge by the water droplets in the steam flow. H O T G A S E X PA N D E R S Hot gas expanders are used widely in large petrochemical complexes. Hot gas expanders are both axial flow and radial inflow design. Most of these expanders are operating on high-pressure and high-temperature waste gases. Most applications use radial inflow design impellers. They are more

| 528 |

| Instrumentation Controls | robust and the waste gases they use are the source of energy efficiency as well as reliability and ease of maintenance. The downside is that waste gases usually have a lot of particulate suspended in the gas stream and, though the gas usually enters through a cyclone, particles still enter the expander. The gas path in most of these expanders is heavily pitted.

CASE STUDIES Condition monitoring is no longer considered as a “nice” tool to have but as a necessary tool to operate the plant in the 21st century. Total condition monitoring consists of a combined performance and mechanical based system. On-line performance-based monitoring introduced by this author to the petrochemical industry in the late seventies has proven to be a very important tool in ensuring that plants are operated at their maximum efficiency. A number of systems installed by the author on various offshore platforms in the Norwegian North Sea led to large savings of fuel by closely monitoring the platform-based machinery. A typical platform installation consisted of between 15–25 trains. Most of the large trains were driven by gas turbines. These trains consisted of aeroderivative gas turbines driving centrifugal compressors. By closely monitoring the gas turbine and scheduling on-line and off-line water washes, a savings of over $45,000 per train was achieved. With the advent of the CO2 tax, which in many cases doubled the cost of fuel, the savings were even larger. The goal in these systems was to prevent the approximate 6% decrease in efficiency per year to about 2–3% per year. These goals were achieved with careful monitoring. The primary parameters that contributed to this were the monitoring of the following parameters: •

Pressure drop across the gas turbine inlet filter

•

Pressure ratio of the gas turbine compressor

•

Efficiency of the gas turbine compressor

•

Turbine firing temperature

•

Turbine exhaust temperature profile

•

Overall turbine thermal efficiency

•

Driven compressor head delivered

•

Efficiency of the various compressor sections

•

Impeller performance

| 529 |

| Centrifugal Compressors | In the same complex, by carefully monitoring the gas turbine, interval times between major overhauls were increased by more than 50% and, in one case, nearly 100%. The schedule overhaul times were 8000 hours for major overhauls. This resulted in savings of over $300,000 per unit. An even more important fact is, that studies have shown, that more than 30% of failures in the field occur within 90 days of a major overhaul. Another U.S. petrochemical complex saved over $1,000,000 by closely monitoring the centrifugal compressor performance in a gas turbine driven train and operating the unit close to the surge line, which reduced the flow throughput by over 12% percent. The centrifugal compressor in this case was re-circulating over 15% of the flow so as to operate in a “safe zone.” Each wheel was carefully monitored by studying the design closely and then inputting the impeller dimensions. The flow in each impeller at its surge point was monitored and fed back to the controller resulting in a reduction of the re-circulation flow to less than 3%. In another petrochemical complex, the gas turbine driven axial flow compressor train suffered a major failure due to major hurricane activity. The failure was caused by the compressor going into surge and then having an axial movement which resulted in the loss of all blades. Further, the heat dissipated during this process, which resulted in the compressor casing being out of round by nearly 15mm. Thus, when the new rotor was placed, the casing could not be closed so the blades had to be reduced by the out of roundness of the casing. This resulted in the blades having large clearances in some areas. It was decided to install an on-line performance system, which could monitor the compressor performance. Dynamic pressure transducers were placed to monitor the compressor flow and to study the flow instabilities. This monitoring allowed the operator to operate the unit for nearly two years until a new casing was available. This being the major cat cracker compressor, the plant would have had to be shut down without this performance monitoring system in place. Performance condition monitoring played a major role in another petrochemical plant where there were several gas turbines driven refrigerant compressors and a large gas turbine driven axial flow compressor. By monitoring the plant performance closely, a major overhaul was postponed for several thousand hours so that it could be performed at a convenient time tied around major sale cycles. Performance monitoring also allowed the turbine to be fired at higher exhaust temperatures, with the agreement of the turbine manufacturer, by proving to them that the actual calculated firing temperatures were lower than those originally corresponding to the pseudo exhaust temperatures. This was due to the

| 530 |

| Instrumentation Controls | fact that the power turbine was operating at about 75% of the design power turbine speed. Thrust bearing problems on several refrigerant compressors was solved using the rate of change alarms on the bearing metal temperatures. The use of pyrometers on turbine blades have been responsible for operating the turbine at higher power ratings by ensuring that the blade metal temperatures were below those that could cause accelerated hot corrosion problems. A particular blade in another installation was diagnosed to have a higher temperature by more than 100º than the other adjacent blades. Inspection of that particular blade pointed to clogged blade cooling air passages. This resulted in the manufacturer changing inspection parameters so that blades with partially closed cooling passages could be easily located.

CONCLUSIONS More and more, maintenance practices are being combined with operational practices to ensure that plants have the highest reliability with maximum efficiency. This has led to the importance of performance condition monitoring as a major tool in the operation and maintenance of a plant. Life cycle costs, rightly so, now drive the entire purchasing cycle and, thus, the operation of the plant. Life cycle costs, based on a 25-year life, indicate that the following are the major cost parameters: •

Initial purchase cost of equipment is 7-10 % of the overall life cycle cost.

•

Maintenance costs are about 15-20% of the overall life cycle cost.

•

Energy costs are about 70-80% of the life cycle costs.

This distribution in life cycle costs indicates that component efficiency throughout the life period of the plant is the most important factor affecting the cost of a particular machine train. Thus, monitoring the efficiency of the train and ensuring that degradation rates are slowed down ensures that the predicted life cycle costs are achieved. Performance monitoring of the entire train is a must for plants operating on life cycle cost strategies. Performance monitoring also plays a major role in extending life, diagnosing problems, and increasing time between overhauls. Online performance monitoring requires an in-depth understanding of the equipment

| 531 |

| Centrifugal Compressors | being measured. Most trains are very complex in nature and thus require very careful planning in installation of these types of systems. The development of algorithms for a complex train needs careful planning and understanding of the machinery and process characteristics. In most cases, help from the manufacturer of the machinery would be a great asset. For new equipment, this requirement can be part of the bid requirements. For plants with equipment already installed, a plant audit to determine the plant machinery status is the first step. To sum up, total performance condition monitoring systems will help the plant engineers to achieve their goals of: •

Maintaining high availability of their machinery

•

Minimizing degradation and maintaining operation near design efficiencies

•

Diagnosing problems, and avoiding operating in regions that could lead to serious malfunctions

•

Extending time between inspections and overhauls

•

Reduce life cycle costs

REFERENCES Boyce, M. P. and Herrera, G. “Health Evaluation of Turbine Engines Undergoing Automated FAA Type Cyclic Testing.” Presented at the SAE International Ameritech ’93. Costa Mesa, CA, 27-30 Sept. 1993. SAE Paper No. 932633. Boyce, M. P. “Total Performance Condition Monitoring Of Major Petrochemical Facilities” Presented at the Materials, Selection and Application in the Chemical Process Industry Seminar, London, UK, Sponsored by The Institute of Materials, and co-sponsored by Institute of Chemical Engineers, and the Institute of Corrosion, November 1998. Boyce, M. P. and Cox, William M. “Condition Monitoring ManagementStrategy.” Presented at The Intelligent Software Systems in Inspection and Life Management of Power and Process Plants in Paris, France, August 1997. Boyce, M. P., Gabriles, G. A., Meher-Homji, C. B., Lakshminarasimha, A. N. and Meher-Homji, F. J. “Case Studies in Turbomachinery Operation

| 532 |

| Instrumentation Controls | and Maintenance using Condition Monitoring.” Proc. of the 22nd Turbomachinery Symposium. Dallas, TX. 14-16 Sept. 1993. 101-12. Boyce, M. P. “Principles of Operation and Performance Estimation of Centrifugal Compressors.” Prod. of the 22nd Turbomachinery Symposium. Dallas, TX. 14-16 Sept. 1993. 161-78. Boyce, M. P., Lakshminarasimha, A. N., and Meher-Homji, C. B. “Design and Implementation of an advanced Multi-Machine Gas Turbine Train Condition Monitoring and Control system.” Presented at the Third Congress of Turbomachinery. Queretaro, Mexico. Dec. 1992. Lakshminarasimha, A. N., Boyce, M. P. and Meher-Homji, C. B. “Modeling and Analysis of Gas Turbine Performance Deterioration.” Presented at the 37th ASME International Gas Turbine and Aeroengine Congress and Exposition. Cologne, Germany. 1-4 June1992. ASME Paper No. 93-GT-395. Boyce, M. P. and Echegaray, L. E. “Aerothermal Analysis Improves Efficiency and Availability of TurboMachinery.” Plant Services Handbook of Computerized Maintenance Management and Predictive Maintenance. Oct. 1989, pp. 102-104. Barnes, G., Boyce, M. P., Wooldridge, B. and Meher-Homji, C. B. “OnLine Condition Monitoring of Two Westinghouse CE 352 Gas Turbine Compressor Sets.” Proc. of the 8th NRCC Symposium on Industrial Application of Gas Turbines. Ottawa, Canada, 24-27 Sept. 1989. 85-98. Boyce, M. P., Bhargava, R. K. and Chinoy, R. “On-Line Condition Monitoring of TurboMachinery.” Proc. of the 1st International Machinery Monitoring and Diagnostic Conference. Las Vegas, NV. 11-14 Sept. 1989. 419-27. Boyce, M. P., Meher-Homji, C. B. and Wooldridge, B. “Condition Monitoring of Aeroderivative Gas Turbines.” Presented at the Gas Turbine and Aeroengine Congress and Exposition. Toronto, Canada. 4-8 June 1989. ASME Paper No. 89-GT-36. Unsal, R. and Boyce, M. P. “Experience with On-Line Condition Monitoring System.” Proc. of the 17th TurboMachinery Symposium. Dallas, TX. 8-10 Nov. 1988. pp. 57-66. Boyce, M. P. “Rerating of Centrifugal Compressors–Part I.” Diesel and Gas Turbine Worldwide. Oct. 1988. 46-50. Boyce, M. P. “How to Identify and Correct Efficiency Losses through Modeling Plant Thermodynamics” Proceedings of the CCGT Generation Power Conference, March, 1999, London U.K.

| 533 |

13

Compressor Performance Testing

An accurate performance evaluation of a dynamic rotary compressor is of paramount interest to the builder because it defines the ability of a machine to perform a specific job and the related requirement for energy expenditure. Since the compression process is continuous rather than batch (as in reciprocating compressors), the physical dimensions of the compressor will not permit accurate performance evaluation. Instead, the characteristics of the machine must be determined by calculation from work done on the gas, as indicated by measurement of observable gas conditions. While a compressor builder might be quite satisfied with relative values for gauging development work, absolute determinations must be made for effective comparison of a machine with its peers. To facilitate such comparisons, a code of practice has been established by the ASME, Power Test Code 10 (PTC 10). It sets guidelines for conducting and reporting tests on a compressor under certain conditions. A compressor’s thermodynamic performance on a specified gas of known properties may be determined during the compression process where, under specified conditions, no condensation or evaporation occurs and there is no injection of liquids. The code defines conditions and provides methods by which a compressor may be tested on a suitable test gas, and the results can be converted into anticipated performance of the same compressor when pumping the specific gas at design conditions. Also, guidelines are

| 535 |

| Centrifugal Compressors | provided for testing compressors with interstage sidestream inlets or outlets, internally cooled compressors, and uncooled tandem-driven compressors with externally piped intercoolers. TEST PLANNING The need to field test centrifugal compressors (in accordance with PTC 10) usually results for one of two reasons. Most commonly, the user has indicated that an equipment train is not performing to expectations. The user knows the compressor facility is not doing the job but not necessarily the reason why. The facility’s energy consumption may be exceeding its design allowance. It may be an apparent source of plant steam balance upset (that is, overloading of the plant cooling water system) or an apparent bottleneck to achieving desired production rates. These factors, and possibly others, can dictate that a test be conducted to define completely the operating characteristics of a compressor. The other reason for field testing in accordance with the code is to satisfy contract requirements that spell out a code test prior to acceptance of a machine from the builder. Regardless of the reason for a test, planning is an essential element for testing that should be considered early in the conceptual state of compressor installation. In addition, PTC 10 can be a helpful guide to monitoring equipment performance. However, in this type of testing the results are rarely reported in detail and are only of interest to the user as a relative indication of a machine’s condition. The real requirement in this test is consistency in test procedure so that successive test results are readily comparable with previous results. Historically, most field test situations preclude application of the test code. Once it is recognized that strict adherence to the code is not possible, it then becomes mandatory that all interested parties (i.e., user, purchaser, contractor, and all equipment suppliers) are in agreement in advance as to the intent of a test, how it is to be accomplished, and what machine performance will constitute satisfactory accomplishment of a test. To field test a compressor in accordance with PTC 10, the initial test planning needs to be done while the compressor installation is in its planning stage. While the incremental cost of providing a proper number of adequate instrumentation facilities is relatively small at the time of

| 536 |

| Compressor Performance Testing | construction, the lost production and incorporation costs for an operating machine generally prohibit their addition. As will be evident from review of ASME test codes and supplements, accuracy of measurement is of prime importance in successful machinery testing. However, too many times the installation is complete and the machine is running before consideration is given to providing for the required accuracy. The descriptions of required piping and instrumentation are provided in later sections. Some of the more common handicaps that frequently cast doubt on the validity of a field test are also discussed later. C L A S S I F I C AT I O N S O F T E S T S Three distinct classes of tests are prescribed by the code, as described in the following sections. Class 1. This class includes all tests made on the specified gas (whether treated as perfect or real) at the speed, inlet pressure, inlet temperature, and cooling conditions for which the compressor is designed and intended to operate. The specified operating conditions and the fluctuations of the test readings should be closely controlled within the limits shown in Table 13-1 and Table 13-2. With these limits, the adjustments that have to be made to the results are kept to a minimum, and the accuracy of the results are at a maximum. Class I tests should be made wherever feasible, since they usually yield the most accurate results. Class 2. These tests are intended for use where the compressor cannot be reliably tested on the specified gas at specified operating conditions. Methods for predicting the performance at specified conditions from a test made at different operating conditions and/or with other gases are discussed later. The reliability of these methods is controlled by an accurate knowledge of the gas properties, the choice of methods used for computing and converting test results, and the extent of deviation from the fundamental design parameters of volume ratio, volume-speed ratio, Mach number, and Reynolds number. Limits for these deviations are shown in Table 13-3 and are mandatory. Class 2 tests differ from Class 3 only in the methods of computation. If the thermodynamic properties of either test gas or the specified gas depart from the perfect gas laws beyond the limits, as shown in Table 13-4, computation methods specified for Class 3 tests should be used. Otherwise, the computation methods for Class 2 can be used.

| 537 |

| Centrifugal Compressors | Test gas is one of the first deviations from the code that frequently requires consideration. While the compressor builder can usually make his shop test with nearly any desired gas, in the field the gas will be whatever is in the process or pipeline. In some instances it will be difficult to supply the same gas for the entire test or even long enough to obtain an accurate data point. This sort of problem can frequently be accommodated to some degree by obtaining frequent gas samples in the course of the test, but in some instances the gas will change slightly by reaction with the sample bomb, thus destroying the relevance of the laboratory analysis. PIPING ARRANGEMENTS The location of pressure, temperature, and flow-measuring devices should have a specific relation to the compressor inlet and outlet openings, as described and illustrated in this section. Minimum lengths of straight pipe are required for flow-measuring devices and for certain pressure measurements. Flow straighteners and/or equalizers should be used in the vicinity of throttle valves and elbows, as shown in Figure 13-1.

Variable

Symbol

Inlet pressure Inlet temperature Specific gravity of gas Speed Capacity Cooling temperature difference Cooling water flow rate

Pi Ti G N qi

Unit

Departure (%)

(1)

psia R ratio rpm cfm °F gpm

5 8 2 2 4 5 3

(2) (2) (2) (3)

Table 13-1: Allowable Departure from Specified Operating Conditions for Class 1 Test. (1) (2) (3)

| 538 |

Departures are based on the specified value where pressures and temperatures are absolute. The combined effect of variables a, b, and c shall not produce more than 8% departure in inlet gas density. Difference is defined as inlet gas temperature minus inlet cooling water temperature. (Source: Power Test Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1965.)

| Compressor Performance Testing | Measurement Inlet Pressure Inlet Temperature Discharge Pressure Nozzle Differential Pressure Nozzle Temperature Speed Torque Electric Motor input Specific Gravity test gas Cooling water inlet temperature Cooling water flow rate Line Voltage

Symbol Pi Ti Pd P 1 - P2 T1 N τ G

Unit psia R psia psi R rpm lb-ft kW ratio °F gpm volts

Fluctuation (1) 2.0% 0.5%(2) 2.0% 2.0% 0.5% 0.5% 1.0% 1.0% 0.25% 3°F(2) 2.0% 2.0%

Table 13-2: Allowable Fluctuation of Test Readings During a Test Run for All Test Classes 1, 2, and 3. (1) Pressure and temperature fluctuation for the gas expressed at percent of average absolute values. (2) Temperature fluctuation for the water is deviation from average in degrees F. (3) Values do not apply for power measurements by heat balance or heat exchanger methods. (Source: Power Test Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1965.)

Range of Test Value Limits Design Value % Variable Symbol Minimum Maximum Volume ratio q1/qd 95 105 Capacity–speed ratio q1/N 96 104 Machine Mach number 0–0.8 Mm 50 105 Above 0.8 Mm 95 05 Machine Reynolds number Re Where the design value is: Below 200,000 centrifugal 90 105 Above 200,000 centrifugal 10(1) 200 Below 100,000 axial compressor 90 105 Above 100,000 axial compressor 10(2) 200 Mechanical losses shall not exceed 10% of the total shaft power input at test conditions. Table 13-3: Allowable Departure from Specified Design Parameters for Class 1 and Class 3 tests. (1) Minimum allowable test machine Reynolds number is 180,000 (2) Minimum allowable test machine Reynolds number is 90,000 (Source: Power Test Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1965.)

| 539 |

| Centrifugal Compressors |

Maximum Ratio Pressure Ratio 1.4 2.0 4.0 8.0 6.0 32.0

γ max* γ min

1.12 1.10 1.09 1.08 1.07 1.06

Maximum Compressibility Functions X Y 0.279 0.167 0.071 0.050 0.033 0.028

1.071 1.034 1.017 1.011 1.008 1.006

Minimum Compressibility Functions X Y -0.344 -0.175 -0.073 -0.041 -0.031 -0.025

0.925 0.964 0.982 0.988 0.991 0.993

Table 13-4: Departure of Gas Properties from Perfect Gas Laws of Test and Specified Gas Permissible for Class 2 Tests. When these limits are exceeded by either the test gas or the specified gas at any state point along the compression path, the methods described for computing Class 3 tests shall be used. *Maximum and minimum values of γ over either test of specified range of conditions:

X=

T  δV    −1 V  δT  P

Y=

P  δV    V  δP  T

(Source: Power Test Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1965.)

Piping configuration is most often not subject to change in the field test situation. The tortuous path often taken by piping as it approaches or leaves the compressor can determine how the compressor will perform, since the piping associated with the compressor will also contain the flow measurement device. This situation may cause the flow meter to be located in such a manner that the meter primary element is not in accordance with ASME codes. Frequently, the only flow meter associated with an installed compressor is for a surge control system that is calibrated empirically and not intended to have absolute accuracy. Even a process flow meter will ultimately have greater utility as a relative indicator than it will have as an absolute meter. In any event, inspection of the primary element of a flow meter prior to test is mandatory. Valves in the piping system also require consideration. Care must be exercised to ensure that any branch connections in the system are properly positioned to account totally for the gas flow through the compressor. Thought must be given to proving that a closed valve is not providing significant flow as a result of leakage. This factor can be particularly important if sidestreams are involved because some systems will not have provision for sidestream flow measurement. Any test will be judged a complete failure if it is determined after the fact that there was unaccounted gas involved in the test that did or did not pass through the compressor.

| 540 |

| Compressor Performance Testing |

Figure 13-1: Flow equalizers and straighteners (Power Test Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1997).

INLET PIPING In shop tests, compressors may be operated without an inlet pipe if the air at the prevailing atmospheric pressures and temperatures will satisfy the conditions required for the test. However, the inlet should be protected with a screen, and a suitably designed bellmouth should be used to minimize the entrance losses. In this case, the inlet stagnation pressure is measured by the barometer, and the temperature should be measured at the screen.

| 541 |

| Centrifugal Compressors |

Figure 13-2: Inlet and discharge configuration (Power Test Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1997).

DISCHARGE TEMPERATURE 4 – MEASURING STATIONS SPACED 90 DEG. INLET TEMPERATURE 4 – MEASURING STATIONS SPACED 90 DEG.

INLET VELOCITY PRESSURE 2 – TAPS, SPACED 90 DEG.

INLET STATIC PRESSURE 4 – TAPS, SPACED 90 DEG. BAFFLE

• (USED ONLY WHEN VELOCITY PRESSURE IS GREATER THAN 5% OF TOTAL PRESSURE.)

DISCHARGE STATIC PRESSURE 4 – TAPS, SPACED 90 DEG.

DISCHARGE VELOCITY PRESSURE • 2 – TAPS, SPACED 90 DEG. CAPACITY BY NOZZLE ARRANGEMENT

THROTTLE VALVE

Figure 13-3: Short inlet pipe (Power Test Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1965).

| 542 |

| Compressor Performance Testing | Figure 13-2 gives the dimensions required for typical inlet and outlet piping. A flow-straightening device should be installed upstream of the pressure-measuring station because, under some conditions in an axial inlet, the impeller produces a vortex which will cause substantial error in the measurement of inlet pressure. Compressors that do not have an axial inlet may be tested with a “short inlet pipe” not less than three pipe diameters in length, as shown in Figure 13-3. Location of the pressure, temperature, and flow-measuring devices are shown in Figure 13-3. The previous conditions described may be impossible to achieve in field testing. Thus, a pitot traverse may have to be used to determine the velocity profile and calculate the flow rate. DISCHARGE PIPING For compressors operating as exhausters, discharge piping is not necessary when the velocity pressure is not more than 50% of the pressure rise. The barometer measures the static discharge pressure, and the discharge gas temperature should be measured at the outlet of the compressor. Compressors with a volute-type diffuser should be tested with a “long outlet pipe” not less than 10 pipe diameters in length, as shown in Figure 13-4. Flow straighteners and vanes are necessary, because volute-type diffusers may be tested with the “short pipe” arrangement, as shown in Figure 13-3.

Figure 13-4: Diffusing volute discharge with nonsymmetric flow (Power Text Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1997).

| 543 |

| Centrifugal Compressors |

Figure 13-5: Typical test cell setup for centrifugal compressors (courtesy of DresserRand Corporation).

| 544 |

| Compressor Performance Testing | CLOSED-LOOP PIPING This type of testing is usually limited to a manufacturer’s shop facilities shown in Figure 13-5. The closed arrangement provides an economical method for testing with many gases other than air and under precisely controlled conditions of pressure and temperature. The essential elements of this arrangement, as shown in Figure 13-6, are the heat exchanger, the throttle valve, the primary element for flow measurement, the inlet pipe, and the discharge pipe A modification of the closed-loop system for testing compressors with side-streams is shown in Figure 13-7. For testing with a condensable gas, the piping arrangement is as shown in Figure 13-8. A combination of these arrangements may be used as needed by the circumstances, provided the piping includes stations for measuring flow, temperature, and pressure for each sidestream independently and in accordance with the requirements in this section.

Figure 13-6: Closed loop test arrangement 1 (Power Text Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1997).

| 545 |

| Centrifugal Compressors | Provisions should be made to remove liquids from the test loop. The test loop should not be charged with air or any oxidizing gases when the compressor shaft seals are lubricated with combustible fluids. Precautions should be taken against overpressures, overtemperatures, loss of cooling water, and other unsafe malfunctions.

Figure 13-7: Closed loop test arrangement 2 (Power Text Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1997).

| 546 |

| Compressor Performance Testing | D ATA A C Q U I S I T I O N The major objective of any compressor test is to obtain good data. However, a proper test setup can be an expensive, time-consuming job. The best check is a rough field calculation that permits plotting the results. With the advent of programmable hand calculators, this task is greatly simplified. Bad data will invariably result in an irregular curve or

Figure 13-8: Closed loop for condensable gases (Power Text Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1965).

| 547 |

| Centrifugal Compressors | reveal calculated values that are impossible. Embarrassment and excessive cost become plentiful when, long after the test setup has been dismantled, the reduced data reveals a machine to have an efficiency greater than 100%. Field testing requires one final admonition. There is no way that a single head-flow point will verify the performance characteristics of a machine. The motor-driven compressor in process service that will not allow for a 20% variation in flow is essentially an un-testable machine. Confronted with a single-point test situation, greater-than-normal precautions must be taken to provide the best possible data if any meaningful information is to be obtained. Evaluation of compressor performance involves determination of the capacity, pressure ratio, horsepower consumed, and surge characteristics for the specific test conditions, including inlet temperature and pressure, discharge pressure, compressor speed, and gas properties. Several measurements of the following parameters are required: •

Inlet temperature

•

Inlet pressure

•

Discharge temperature

•

Discharge pressure

•

Barometric pressure

•

Compressor speed

•

Differential pressure across flow meter (or pitot traverse)

•

Temperature and pressure at flow meter

•

Gas properties

PRESSURE MEASUREMENT The following types of instruments are used to make pressure measurements:

| 548 |

•

Bourdon tube gauges

•

Dead-weight gauges (used for calibration purposes only)

•

Liquid manometers

•

Impact tubes

| Compressor Performance Testing | •

Pitot-static tubes

•

Pressure transmitters

•

Pressure transducers

•

Barometers

Good-quality Bourdon tube test gauges are highly suitable for pressure measurements of more than 20 psi. They should be calibrated against a dead-weight tester in their normal operating range. When selecting a pressure gauge, it is important to see that the measure value is above midpoint on the scale. Differential pressures and subatmospheric pressures should be measured by manometers with a fluid that is chemically stable when in contact with the test gas. Mercury traps should be used where necessary to prevent the manometer fluid from entering the process piping. Errors in these instruments should not exceed 0.25%. A common failure in pressure measurement is the uncertainty of the configuration of static pressure taps penetration through the pipe wall. This failure is another early planning concern, since proper taps are easy to provide prior to placing the machine in service, but inspection of the taps after operation has commenced is a luxury rarely afforded the test team. Another pitfall in pressure measurement, particularly important in flow measurement, is the potential for liquids in gauge lines. All too often, gauge lines coming from overhead pipes have no provision for maintaining a liquid free status, even though the flowing fluid may be condensable at gauge-line temperatures. Calibration of the pressure-measuring device presents another pitfall for test crews. All too often, a test is conducted through the field calculation step before bad data reveals that gauges, possibly with too large a minimum increment, were removed from the shipping carton and installed, relying on the vendor’s calibration. On-site calibration of all instruments is always good insurance against a bad test. Frequently, new machines are put into service with a “startup screen” in the compressor inlet piping to guard against the inevitable weld slag and construction debris that will remain in a new or rebuilt piping system after construction. Regardless of the age of the installation, care must be exercised to ensure that measurements defining suction or discharge conditions are not influenced by such devices. Inlet and discharge pressures are defined as the stagnation pressures at the inlet and discharge, which are the sum of static and velocity

| 549 |

| Centrifugal Compressors | pressures at the corresponding points. Static pressures should be measured at four stations in the same plane of the pipe, as shown in the piping arrangements. Velocity pressure, when less than 5% of the pressure rise, can be computed by the formula Pv =

(V av ) 2 ρ (V )2 ρ = av 2g c x 144 9266.1

(13-1)

where Vav is the ratio of measured volume flow rate to the cross-sectional area of the pipe. When the velocity pressure is more than 50/Q of the pressure rise, it should be determined by a pitot-tube traverse of two stations. For each station, the traverse consists of 10 readings at positions representing equal areas of the pipe cross section, as shown in Figure 13-9. The average velocity pressure Pv, is given by 3

Pv =

ρ ΣV P 288 gc nt Vav

(13-2)

where at each traverse point VP =

9266.1 pv ρ

and n equals the number of traverse points. Barometric pressure should be measured at the test site at 30-minute intervals during the test. T E M P E R AT U R E M E A S U R E M E N T Temperature may be measured by any of the following instruments: •

Mercury-in-glass thermometers

•

Thermocouples

•

Resistance thermometers

•

Thermometer wells

Thermocouples are the preferred type of instruments because of the simplicity in basic design and operation. They can attain a high level of accuracy, are suitable for remote reading, and are robust and relatively inexpensive.

| 550 |

| Compressor Performance Testing |

Figure 13-9: Traverse points in pipe (Power Text Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1965).

Regardless of the temperature measuring device to be used, on-site calibration of the entire measurement system is desirable. Usually, a twopoint check can be made by employing frozen and boiling water. At the very least, all devices can be checked at a common temperature, preferably in the midrange of expected temperatures, so that any deviant devices can be discarded. This check is particularly desirable for low-head machines where the temperature rise will be slight. Test plans frequently are prepared on the assumption that a laboratory thermometer can replace an operating instrument in an existing thermometer well. While this change may be satisfactory, the prudent tester needs to be aware that, because of the propensity of thermowells to break off and perhaps enter the machine or cause a hazardous leak, their design is compromised such that true gas temperature determination is impossible. The compromise may be to make the well short and/or to make it thickwalled. In either event, the mass of metal exposed to ambient temperature

| 551 |

| Centrifugal Compressors | may exceed that exposed to the gas, resulting in significant error if the gas temperature is much different from the ambient. High-pressure systems requiring thick-wall pipe are particularly susceptible to this fault. However, the use of a good heat transfer fluid can minimize the error. The best gas temperature reading is attained by a calibrated fine-wire thermocouple with the junction directly exposed to the gas near the center of the flow. As deviations from this ideal are made, the potential for error is increased. Inlet and discharge temperatures are the stagnation temperatures at the respective points and should be measured within an accuracy of l°F. When the velocity of the gas stream is more than 125 fps, the velocity effect should be included in the temperature measurement with a total temperature probe. This probe is a thermocouple with its hot junction provided with a shielded cup. The cup opening points upstream. A tradeoff has to be made in a field test situation where the gas is not clean. FLOW MEASUREMENT Gas flow through the compressor is measured by flow nozzles or other devices installed in the piping. Among the various devices are: •

Orifice plates. Either the concentric orifice, eccentric orifice, or segmented orifice-type. Choice depends on the quality of the fluid handled.

•

Venturi tubes. These consist of a well-rounded convergent section at the entrance, a throat of constant diameter, and a divergent section. Their accuracy is high; however, installation, unless planned for in advance, is very difficult in the field.

•

ASME flow nozzle. These nozzles provide for accurate measurements. Their use is limited because they are not easily placed in a process plant; however, they are excellent for shop tests. Venturi meters and nozzles can handle about 60% more flow than orifice plates with varied pressure losses.

•

Elbow flow meters. The principle of centrifugal force at the bend is used to obtain the difference in pressure at the inside and outside of the elbow, which is then related to the discharge pressure.

Other techniques for measuring flow through the compressor include:

| 552 |

| Compressor Performance Testing | •

Calibrated pressure drops from the inlet flange to the eye of the first-stage impeller in centrifugal compressors, when such data is available from the manufacturer.

•

A flow trace technique in which Freon is injected into the constream and flight time between two detection points is measured.

•

Velocity traverse techniques must be used when, due to the configuration in piping, nozzles or orifice plates cannot be used.

These techniques have been described previously in the pressure measurement section. Usually, one of the flow measuring devices and the required instrumentation is incorporated as a part of the plant piping. The choice of technique depends on the allowable pressure drop, flow type, accuracy required, and cost. Nozzle arrangements for various applications vary considerably. For subcritical flow measurement at the outlet end, where nozzle differential pressure p is less than the barometric pressure, flow should be measured with impact tubes and manometers, as shown in Figure 13-10.

Figure 13-10: Discharge nozzle on an open loop, subcritical flow (Power Text Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1997).

| 553 |

| Centrifugal Compressors | For critical measurement, where the drop in p is more than the barometric pressure, flow should be measured with static-pressure taps upstream from the nozzle, as illustrated in Figure 13-11. For compressors operating as exhausters, differential pressure is measured at two static taps located down-stream from the nozzle at the inlet, as shown in Figure 13-12. Nozzle arrangement to measure flow within a closed-loop system is shown in Figure 13-6.

Figure 13-11: Discharge nozzle on an open loop, critical flow (Power Text Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1997).

Figure 13-12: Inlet nozzle on an open loop (Power Text Code 10, Compressors and Exhausters, American Society of Mechanical Engineers, 1997).

| 554 |

| Compressor Performance Testing | POWER MEASUREMENT Power consumed by a compressor can be measured directly at the drive shaft or at the coupling. It can also be determined indirectly from measurement of electrical input to the driving motor, by a heat-balance method, or by the heat absorbed at the heat exchanger in a closed-loop arrangement. Typical impediments to good field tests are steam turbines without flow meters, engines without fuel meters, motors with unreliable potential, and current-measuring devices. The torque at the drive shaft can be measured by strain gauges or optical devices. Torque, together with compressor speed, will yield shaft power hpsh consumed hp sh =

τN 5252.1

(13-3)

When an electric motor is used to drive the compressor, hpsh =

net kW input x motor efficiency − gear losses 0.7457

(13-4)

By the heat-balance method, hp sh =

m ( hd − hi ) + Qr + Qm + Qsl 42.408

(13-5)

where: . m = mass flow rate, 1bm/min hd = enthalpy at discharge, Btu/1bm hi = inlet enthalpy, Btu/1bm Qr = external heat loss from casing, Btu/1bm Qm = total mechanical losses, Btu/1bm Qsl = external seal loss equivalent, Btu/lbm SPEED MEASUREMENT In most cases, speed can be measured with an electrical counter actuated by a magnetic pulse generator or a 60-tooth gear. The latter method is preferred, and the driver manufacturer usually installs a gear on the driver shaft for this purpose. Optical probes can be used by placing

| 555 |

| Centrifugal Compressors | reflective tape on a shaft; however, care should be taken not to place it in direct light. With a synchronous motor drive, the speed can be calculated with only a slight error from the number of poles and line frequency. TEST PROCEDURES The goal of any type of compressor testing is to obtain a compressor map, as shown in Figure 13-13. This map plots the corrected flow versus the pressure rise at various aerodynamic speeds. The aerodynamic speeds show the effect that ambient temperature has on the operating characteristics of the compressor. In selecting a particular test procedure, consideration should be given to plan operation and the kind of variables available. To obtain the characteristics of a compressor, data should be recorded for different values of Q/N, flow-to-speed ratio. The choice of techniques used for varying the Q/N ratio depends on the flexibility of the compressor or the variable speed prime mover. For fixed-speed drives, the compressor is run near its overload condition. By throttling either the inlet or the discharge valve, the flow is decreased in increments. Fine flow adjustments can be made using a bypass across the throttle valves. Data are recorded at each valve of the flow until a minimum flow operating point is reached. Also, taking points over a great range of ambient temperatures will provide data for the operating speed line. For variable-speed drives, the flow can be varied by changing the speed. Characteristics of the compressor conditions should be allowed to stabilize by holding the speed inlet conditions constant. All the measured parameters should be monitored, and, when they level-out, two sets of readings should be recorded. The procedure is repeated for all the operating points. Usually, a speed line is determined by a minimum of three points; however, five points are desirable. The surge line is determined also by a minimum of three speed lines and is usually parallel to the operating line. Care should be exercised to see that the compressor is not thrown into surge when operating near minimum flow conditions. Gas samples should be collected at regular intervals and analyzed for their properties. Samples collected at the inlet and exit of the compressor provide a meaningful insight to the changes in molecular weight and percentage of the component gases. These samples at regular intervals can also be used for statistical analysis.

| 556 |

| Compressor Performance Testing |

Figure 13-13: Typical compressor map.

T E S T C O M P U TAT I O N S The calculations of the various parameters are based on the assumption that the basic measurements followed the techniques outlined previously. The compressor is tested at various fractions, x, of design speed xN design =

xN test Θ

,Θ =

Tt1 520

(13-6)

The abscissa of the compressor performance characteristics is the corrected mass flow rate, given as: m c =

m Θ δ

(13-7)

. where m is the mass flow rate calculated based upon the type of flow measuring device used and

| 557 |

| Centrifugal Compressors |

δ=

(13-8)

Pt1 14.7

The adiabatic efficiency of the unit can be computed by the use of the following relationship: γ −l    Pt 2  γ   − 1 Ttl  P  t1    η ad =  ∆ Tact

(13-9)

Polytropic index of compression for the gas mixture can be calculated with the following relationship: ηp =

(γ −1) γ

(n −1)

(13-10)

n

This relationship leads to a polytropic efficiency, calculated as follows, for a perfect gas: ηp =

(γ −1) γ

(n −1) n

(13-11)

Polytropic efficiencies are the truncated values of the Taylor series and agree closely with the adiabatic efficiency at low-pressure ratios. At high-pressure ratios, they are higher, as seen in Figure 13-14. For real gases, the index of adiabatic compression k can be computed as follows: γ −1 R' = γ M mix c pmix

  ∂Z      Z + Trmix    ∂Trmix  Prmix 

where the expression   ∂Z     R ′  Z + Trmix    ∂Trmix  Prmix 

| 558 |

(13-12)

Isentropic Adiabatic Efficiency (%)

| Compressor Performance Testing |

Figure 13-14: Relationship between polytropic and adiabatic isentropic efficiency for a centrifugal processor.

has been represented as a function of Trmix and Prmix and the reduced temperature and pressure of the mixture, respectively. This function is shown in Figures 13-15 and 13-16. As an example, the following steps determine cpmix , Mmix , Trmix , and Prmix: The specific heat at constant pressure for the gas mixture is estimated as c pmix = Σx i c pi

(13-13)

where xi is the molecular fraction of the ith component gas. Consider natural gas compressed to a pressure of 210 psig and 120°F. Methane, ethane, and propane are the main components of the gas with properties, as shown in Table 13-5. Then 1. Mmix = (.84 X 16.0) + (.10 X 30.1) + (.06 X 44.1) =19.10 2. Tcrmix = (.84 X 344) + (.10 X 550) + (.06 X 66) °R = 384°R

| 559 |

| Centrifugal Compressors |

Figure 13-15: Isentropic Z function—low range (courtesy of Chemical Engineering Progress).

| 560 |

| Compressor Performance Testing | Hence Trmix =

120 + 460 = 1.510 384

(13-14)

3. Pcrmix = (.84 X 673) + (.10 X 708) + (.06 X 617) psia = 676.1 psia. Hence Trmix =

120 + 460 = 1.510 384

With the help of these quantities, the adiabatic index γ can be computed from Equation 13-12.

Figure 13-16: Isentropic Z function—high range (courtesy of Chemical Engineering Progress).

| 561 |

| Centrifugal Compressors | Gas Composition

Mol. Weight

Methane CH4 Ethane C2H6 Propane C3H8 Butane C4H10 Pentanes C5H12 Carbon Dioxide CO2 Nitrogen N2 Helium He Heating value Btu/ft3 (kJ/m3) Specific gravity Ref.: Air at 60°F (288°K)

16.043 30.068 44.094 58.123 72.15 44.01 28.0134 4.0026 0.586

Gas Composition Volume Range (%)

Critical Pressure

Critical Temperature

Low High 86.3 95.2 2.5 8.1 0.6 2.8 0.13 0.66 0 0.44 0 1.1 0.31 2.47 0.01 0.06 1024 1093 (38,150) (40,720) 0.641

psia 673 708 617 715

°R 334 550 666 765

1094 502 3.9

548 227 9.33

Table 13-5: Analysis of Natural Gas. Adapted from Gas Engineers Handbook, American Gas Association, Industrial Press, New York 1965. Ranges are the high and low value of annual averages reported by 13 utilities (1954 data).

These properties are obtained from chemical analysis, and the rest are the intensive properties of the respective gases. The polytropic head developed by the compressor is calculated as n −1    Pt 2  n  − 1 Z av RTtl     Ptl     Hp =  n −1     n 

(13-15)

where R is the universal gas constant. The average compressibility factor Zav, necessary to calculate the head developed by the compressor is computed as follows: Zav =

Zin + Z ex 2

Zin = Z°in + ωZ'in And

Zex = Z°ex + ωZ'ex

| 562 |

(13-16)

| Compressor Performance Testing | (Z°in, Z'in) and (Z°ex, Z'ex), corresponding to inlet and exit conditions of temperature and pressure, have been plotted as functions of Trmix and Prmix (Figures 13-17 and 13-18). The eccentric factor can be evaluated as   3 1n (Pcr in atm )  ω =  − 1 Tcr 7  −1    Tboil

(13-17)

Tboil is the normal boiling point of the component gas at 14.7 psia. For a mixture of gases, ω mix = Σ xi mi

(13-18)

Alternately, the compressibility factor can be evaluated by means of the various empirical equations of state, such as the Benedict-Webb-Rubin equation. This equation in its virial form is given as  C  2 P = ρ R ′T +  B0 R′T − A0 R ′T − A0 − 0  ρ T2   3 6 + (bR ′T − a ) ρ + aαρ +

Cρ 3 2 −γρ 2 2 (1 + γρ )e T

(13-19)

The empirical constants are B0, A0, b, a, ω, c, and γ. Rearranging the previous equation, the expression for the compressibility factor becomes A C Z = 1 +  B0 − 0 − 0  ρ R ′T R ′T   a  2 aα 5  + b − ρ ρ + R ′T  R ′T  2

+

2 Cρ (1 + γρ 2 )e −γρ 3 R ′T

(13-20)

This equation of state is widely used to evaluate the thermodynamic properties of a number of single-component, hydrocarbon systems. A system utilizing a combination of these hydrocarbons is handled with the help of the following set of relationships between the system- and component-constants:

| 563 |

| Centrifugal Compressors | (13-20)

Bomix = Σx1 Boi ; A01 / 2 mix = Σxi A0i 1 / 2 Comix = Σxi C0 i1 / 2 ; b1 / 3 mix = Σx i bi1 / 3 α

1/ 3

mix

= Σxiα

c 1 / 3 mix = Σxi ci

1/ 3

1/ 3

; and α

1/ 3

mix

= Σx iα i

; γ 1 / 2 mix = Σx iγ i

1/ 3

1/ 2

Z COMPRESSIBILITY FACTOR FOR SIMPLE FLUID

PR REDUCED PRESSURE

Figure 13-17: Generalized compressibility factor for simple fluid (adapted with permission from Journal of the American Chemical Society, Copyright 1955, American Chemical Society).

| 564 |

| Compressor Performance Testing | The relative variations of these constants, with regard to the constant Bo, for a number of hydrocarbons are shown in Figures 13-19a, 13-19b, 13-19c, 13-19d. These graphs and Equations, (13-18) and (13-19), enable the computation of the compressibility factor. However, it is important to note that the Benedict-Webb-Rubin equation of state fails to hold for systems with more than two component hydrocarbons. Finally, the horsepower required to drive the compressor is estimated by the following relationship: Hp =

m H p

(13-22)

ηp

Z COMPRESSIBILITY FACTOR CORRECTION FOR DEVIATION FROM SIMPLE FLOW

TR REDUCED TEMPERATURE

Figure 13-18: Generalized compressibility factor correction from deviation from simple fluid (adapted with permission from Journal of the American Chemical Society, Copyright 1955, American Chemical Society).

| 565 |

| Centrifugal Compressors |

Figure 13-19a: Constants of Benedict-Webb-Rubin equation.

Figure 13-19b: Constants of Benedict-Webb-Rubin equation.

| 566 |

| Compressor Performance Testing |

Figure 13-19c: Constants of Benedict-Webb-Rubin equation.

Figure 13-19d: Constants of Benedict-Webb-Rubin equation.

| 567 |

| Centrifugal Compressors | REFERENCES Boyce, M. P., Bayley, R. D., Sudhakar, V., and Elchuri, V. “Field Testing of Compressors.” Proceedings of the 5th Turbomachinery Symposium. Texas A&M Univ., 1976, pp. 149-160. ASME Power Test Code 10 (PTC-10), 1997. Edmister, W. C. Applied Hydrocarbon Dynamics, Vol. 1. Gulf Publishing Co. Houston, TX. 1961. pp. 1–3. Canjar, L. N. “There’s a Limit to Use of Equations of State.” Petroleum Refiner, February 1956, p. 113.

| 568 |

14

Maintenance Techniques

Maintenance, defined as the “upkeep of property,” is one of the most important operations in a plant. The manufacture and maintenance of turbomachinery are totally different. The first involves the shaping and assembly of various parts to required tolerances, while the second, maintenance, involves restoration of these tolerances through a series of intelligent compromises. The crux of maintenance technique is in keeping the compromises intelligent. Maintenance is not a glamorous procedure; however, its importance is second to none. Maintenance procedures are always controversial, since the definition of “upkeep” varies with the individual interpretation of each maintenance supervisor. The latitude of maintenance ranges from strict planning and execution, inspection, and overhaul, accompanied by complete reports and accounting of costs, to the operation of machinery until some failure occurs and then making the necessary repairs. Modern day turbomachinery is built to last between thirty to forty years. Thus, the keeping of basic maintenance records and critical data is imperative for a good maintenance program. Economic justification is always the controlling factor for any program, and maintenance practices are not different. Maintenance costs can be minimized by, and are directly related to, good operation; likewise, better operating results can be obtained when the equipment is under the control of a planned maintenance program. Improper operation of mechanical equipment can be as much or more the

| 569 |

| Centrifugal Compressors | cause of its deterioration and failure as is actual, normal mechanical wear. Thus, operation and maintenance go together. Combining the practice of preventive maintenance with total quality control and total employee involvement results in an innovative system for equipment maintenance that optimizes effectiveness, eliminates breakdowns, and promotes autonomous operator maintenance through day-today activities. This concept known as Total Productive Maintenance (TPM) was conceived by Seiichi Nakajima and is well documented in his book Introduction of TPM. It is highly recommended reading for all involved in the maintenance area. A new maintenance system is introduced based on the new mantra for the selection of all equipment “Life Cycle Cost.” This new system, especially for major power plants, is based on the combination of total condition monitoring, and the maintenance principles of total productive maintenance and is called the “Performance Based Total Productive Maintenance System.” The general maintenance system is fragmented and can be classified into many maintenance concepts. The following are five Ps of maintenance for major power plants, petro-chemical corporations, and other process type industries leading to the ultimate maintenance system: 1.

Panic maintenance based on breakdowns

2.

Preventive maintenance

3.

Performance based maintenance

4.

Performance productive maintenance

5.

Performance based total productive maintenance (PTPM)

Performance based total productive maintenance consists of the following elements:

| 570 |

•

Performance based total productive maintenance aims to maximize equipment efficiency and time between overhaul (overall performance effectiveness).

•

Performance based total productive maintenance aims to maximize equipment effectiveness (overall effectiveness).

•

Performance based total productive maintenance establishes a thorough system of PM for the equipment’s entire life span.

•

Performance based total productive maintenance is implemented by various departments (engineering, operations, maintenance).

| Maintenance Techniques | •

Performance based total productive maintenance involves every single employee, from top management to workers on the floor.

•

Performance based total productive maintenance is based on the promotions of PM through motivation management: autonomous small group activities.

The word “total” in “performance-based total productive maintenance” has the following meanings that describe the principal features of PTPM: •

Total overall performance effectiveness (referred to in the first bulleted point above) indicates PTPM’s pursuit of maximum plant efficiency and minimum downtime.

•

Total Overall performance effectiveness (referred to in the second point above) indicates PTPM’s pursuit of economic efficiency or profitability

•

Total maintenance system (the third point) includes maintenance prevention (MP) and maintainability improvement (MI), as well as preventive maintenance.

•

Total participation of all employees (the last three points) includes autonomous maintenance by operators through small group activities.

Table 14-1 shows the relationship between PTPM, productive maintenance, and preventive maintenance.

Performance Based Total Performance Performance Productive Productive Based Preventive Panic Maintenance Maintenance Maintenance Maintenance Maintenance Economic Efficiency Economic and Time Efficiency Total System Efficiency Autonomous Maintenance by Operators

Yes Yes

Yes Yes

Yes Yes

Yes No

No No

Yes

Yes

No

No

No

Yes

No

No

No

No

Table 14-1: Benefits of various maintenance systems maintenance.

| 571 |

| Centrifugal Compressors | Performance based total productive maintenance eliminates the following seven major losses: Downtime 1.

Loss of time due to unnecessary overhauls based only on time intervals

2.

Equipment failure from breakdowns

3.

Loss of time due to spare part unsuitability or insufficient spares

4.

Idling and minor stoppages-due to the abnormal operation of sensors, or other protective devices

5.

Reduced output—due to discrepancies between designed and actual operating conditions

Defects: 6.

Process defects due to improper process conditions that do not meet machinery design requirements

7.

Reduced yield from machine startup to stable production due to inability of machine to operate at proper design conditions

MAXIMIZATION OF EQUIPMENT EFFICIENCY AND EFFECTIVENESS Maintaining the health of the equipment can help to achieve high machine efficiency and availability. Total performance condition monitoring can play a major part here as it provides early warnings of potential failures and performance deterioration. Figure 14-1 shows the concept of a total performance condition monitoring system. Pure preventive maintenance alone cannot eliminate breakdowns. Breakdowns occur due to many factors, such as design and or manufacturing errors, operational errors, and wearing out of various components. Thus, changing out components at fixed intervals does not solve the problems and, in some cases, adds to the problem. A study at a major nuclear power station indicated that nearly 35% of the failures occurred within a month of a major turnaround. Figure 14-2 shows the life characteristics of a major piece of turbomachinery.

| 572 |

| Maintenance Techniques |

Figure 14-1: Total performance based condition monitoring system.

Figure 14-2: Machinery life cycle characteristics.

| 573 |

| Centrifugal Compressors | The goal of any good maintenance program is zero breakdown. The following five countermeasures help achieve this goal: 1.

Maintaining well regulated, basic conditions (cleaning, lubricating, and bolting)

2.

Adhering to proper operating procedures

3.

Total condition monitoring (performance, mechanical and diagnostic based)

4.

Improving weaknesses in design

5.

Improving operation and maintenance skills.

Figure 14-3: Breakdown countermeasures.

Figure 14-4: Responsibilities of the operations and maintenance departments.

| 574 |

| Maintenance Techniques | The interrelationship between these five items is shown in Figure 14-3. The division of labor between operations and maintenance is shown in Figure 14-4. It is the primary responsibility of the production department to establish and regulate basic operating conditions, and it is the primary responsibility of the maintenance department to improve defects in design. The other tasks are shared between the two departments. The successful implementation of total productive maintenance requires: •

Elimination of the six big losses to improve equipment effectiveness

•

An autonomous maintenance program with total condition monitoring.

•

A scheduled maintenance program for the maintenance department

•

Increased skills of operations and maintenance personnel

•

An initial equipment management program

ORGANIZATION STRUCTURES FOR A PERFORMANCE BASED TOTAL PRODUCTIVE MAINTENANCE PROGRAM Typically successful implementation of PTPM in a large plant takes three years. Implementation calls for: 1.

Changing people’s attitudes

2.

Increasing Motivation

3.

Increasing Competency

4.

Improving the work environment

The four major categories in developing a performance based total productive maintenance program are: 1.

Preparation for the PTPM Program

2.

Preliminary implementation

3.

PTPM implementation

4.

Stabilization of the program

| 575 |

| Centrifugal Compressors | I M P L E M E N TAT I O N O F A P E R F O R M A N C E B A S E D T O TA L P R O D U C T I V E M A I N T E N A N C E There are several steps involved in implementation of a PTPM program.

| 576 |

1.

Announcement of Decision to Implement PTPM. Top management must make a formal presentation by introducing the concepts, goals, and benefits of PTPM. Management commitment must be made clear to all levels of the organization.

2.

Educational Campaign. The training and promotion of PTPM philosophy is a must. This is useful to reduce the resistance to change. The education should cover how PTPM will be beneficial to both the corporation and the individuals.

3.

Creation of Organization to Promote PTPM. The PTPM promotional structure is based on an organizational matrix. Obviously, the optimal organizational structure would change from organization to organization. In large corporations, PTPM promotional headquarters must be formed and staffed. Thus, any questions can be addressed here on a corporate level.

4.

Establishment of Basic PTPM Goals. Establishing mottos and slogans can do this. All goals must be quantifiable and precise specifying: •

Target (what)

•

Quantity (how much)

•

Time frame (when)

5.

Master Plan Development for PTPM. A master plan must be created. Total condition monitoring equipment should be designed and equipment purchased

6.

Initiation of PTPM. This represents a “kickoff” stage. At this point, the whole staff must start to get involved.

7.

Improvement of Equipment Effectiveness. This should start with a detailed design review of the plant machinery. A performance analysis of the plant could point to a specific area known to have problems (i.e., section of plant). This must be selected and focused on, project teams should be formed, and

| Maintenance Techniques | assigned to each train. An analysis should be conducted that addresses the following:

8.

•

Define the problem. Examine the problem (loss) carefully; compare its symptoms, conditions, affected parts, and equipment with those of similar cases.

•

Do a physical analysis of the problem. A physical analysis clarifies ambiguous details and consequences. All losses can be explained by simple physical laws. For example, if scratches are frequently produced in a process, friction or contact between two objects should be suspected. (Of the two objects, scratches will appear in the object with the weaker resistance.) Thus, by examining the points of contact, specific problem areas and contributing factors are revealed.

•

Isolate every condition that might cause the problem. A physical analysis of breakdown phenomena reveals the principles that control their occurrence and uncovers the conditions that produce them. Explore all possible causes.

•

Evaluate equipment, material, and methods. Consider each condition identified in relation to the equipment, jigs and tools, material, and operating methods involved and draw up a list of factors that influence the conditions.

•

Plan the investigation. Carefully plan the scope and direction of investigation for each factor. Decide what to measure and how to measure it and select the datum plane.

•

Investigate malfunctions. All items planned in Step 5 must be thoroughly investigated. Keep in mind the optimal conditions to be achieved and the influence of slight defects. Avoid the traditional factor analysis approach; do not ignore malfunctions that might otherwise be considered harmless.

•

Formulate improvement plans. Define consultants who could re-design the given piece of equipment. Discuss with manufacturers your plans.

Establishment of Autonomous Maintenance Program for Operators. This is focused against the classic “Operations vs. Maintenance” battle. Operators here must be convinced that they should maintain their own equipment. For example, an attitude

| 577 |

| Centrifugal Compressors | has to be developed for operators to understand and act on the reports produced by the on-line performance condition monitoring systems. 9.

Setup of Scheduled Maintenance Program. Scheduled maintenance conducted by the maintenance department must be smoothly coordinated with autonomous maintenance done by the plant operators. This can be done by frequent meetings and plant audits. In most plants, an undeclared conflict exists between the operations and maintenance groups. This arises from the false perception that these two groups having conflicting goals. The PTPM philosophy will go a long way in bringing these groups together.

10. Training for Improvement of Operation and Maintenance Skills. This is a key part of PTPM. Ongoing training in advanced maintenance techniques, tools, and methods must be done. This could cover areas such as: •

Bearings and Seals

•

Alignment

•

Balancing

•

Vibration

•

Troubleshooting

•

Failure Analysis

•

Welding Procures

•

Inspection Procedures

•

NDT

11. Equipment Management Program. Startup problems, solutions, and design changes should be clearly documented and available for use in a good equipment management plan. All items that can reduce Life Cycle Costs (LCC) should be considered. These include:

| 578 |

•

Economic evaluation at the equipment investment stage

•

Consideration of MP or maintenance free design and economic LCC

•

Effective use of accumulated MP data

| Maintenance Techniques | •

Commissioning control activities

•

Thorough efforts to maximize reliability and maintainability

12. Final Implementation of PTPM. This stage involves the refinement of PTPM and the formulation of new goals that meet specific corporate needs.

MAINTENANCE DEPARTMENT REQUIREMENTS To ensure the success of the PTPM program, the maintenance department must be well equipped and trained. The following six basic categories are prerequisite to the proper functioning of the maintenance department under the PTPM: 1.

Training of personnel

2.

Tools and equipment

3.

Inspections

4.

Condition and life assessment

5.

Spare parts inventory

6.

Redesign for higher machinery reliability

7.

Maintenance scheduling

8.

Maintenance communication

TRAINING OF PERSONNEL Training must be the central theme. The days of the mechanic armed with a ball-peen hammer, screwdriver, and a crescent wrench are gone. More and more complicated maintenance tools must be placed in the hands of the mechanic, and he must be trained to utilize them. People must be trained, motivated, and directed so that they gain experience and develop, not into mechanics but into highly capable technicians. While good training is expensive, it yields great returns. Machinery has grown more complex, requiring more knowledge in many areas. The old, traditional craft lines must yield before complicated equipment maintenance needs. A joint effort by craftsmen is necessary to accomplish this.

| 579 |

| Centrifugal Compressors | TYPE OF PERSONNEL Maintenance Engineer In most plants, the maintenance engineer is a mechanical engineer with training in the turbomachinery area. His needs are to convert what he has learned in the classroom into actual hands-on solutions. He must be well versed in a number of areas such as performance analysis, rotor dynamics, metallurgy, lubrication systems, and general shop practices. His training must be well planned so that he can pick up these various areas in steps. His training must be a combination of a hands-on approach coupled with the proper theoretical background. He should be well versed in the various ASME Power Test Codes. Table 14-2 is a listing of some of the applicable codes for process compressors and their drivers. Attendance at various symposiums where users of machinery get together to discuss problems should be encouraged. It is not uncommon to find a solution to a problem at these types of roundtable discussions. 1.

ASME, Performance Test Code on Test Uncertainty: Instruments and Apparatus, American Society of Mechanical Engineers PTC 19.1, 1988

2.

ASME, Performance Test Code on Gas Turbines, ASME PTC 22 1997

3.

ASME, Performance Test Code on Steam Turbines, ASME PTC 6 1996

4.

ASME, Performance Test Code on Steam Condensing Apparatus, ASME PTC 12.2 1983

5.

ASME, Performance Test Code on Atmospheric Water Cooling Equipment, PTC 23, 1997

6.

ASME Gas Turbine Fuels, B 133.7M Published: 1985 (Reaffirmed year: 1992)

7.

ISO, Natural Gas—Calculation of Calorific Value, Density and Relative Density, ISO 6976-1983(E)

Table 14-2: Performance test codes.

Foremen and Lead Machinist These men are the key to a good maintenance program. They should be sent frequently to training schools to enhance their knowledge. Some

| 580 |

| Maintenance Techniques | plants have one foreman who is an “in-house serviceman;” he supervises no personnel but acts as an in-house consultant on maintenance jobs. Machinist/Millwright The machinist should be encouraged to operate most of the machinery in the plant maintenance shop. By rotating him among various jobs, his learning and development is accelerated. He should then become as familiar with a large compressor as a small pump. Encouragement should be given to the machinist to learn balancing operations and to participate in the solution of problems. Spreading around the hardest jobs develops more competent people—the basis of any PTPM program. Restricting a man to one type of work will probably make him an expert in that area, but his curiosity and initiative, prime motivators, will eventually fade. TYPES OF TRAINING Update Training This training is mandatory for all maintenance personnel, so that they may keep abreast of this high technology industry. Personnel must be sent to schools run by manufacturers (OEM). These schools, in turn, should be encouraged to cover some basic machinery principles as well as their own machinery. In-house seminars should be provided with in-house personnel and consultants at the plant. Engineers should be sent to various schools so that they may be exposed to the latest technology. An in-house website, cataloging experiences and special maintenance techniques should be updated and available for the entire corporation especially maintenance and operation personnel. These websites should be full of illustrations, short, and to the point. A small library should be adjacent to the shop floor, with field drawings, written histories of equipment, catalogs, API specifications, and other literature pertinent to the machine maintenance field. Drawings and manuals should be transferred to the electronic digital media as soon as possible. Access to the Internet on the maintenance and production area computers is a must, as many manufacturers post helpful operational and maintenance hints on their websites. API specifications, which govern mechanical machinery, are listed in Table 14-3.

| 581 |

| Centrifugal Compressors | ASME Basic Gas Turbines B 133.2 Published: 1977 (Reaffirmed year: 1997) ASME Gas Turbine Control And Protection Systems B133.4 Published: 1978 (Reaffirmed year: 1997) ASME Gas Turbine Installation Sound Emissions B133.8 Published: 1977 (Reaffirmed: 1989) ASME Measurement Of Exhaust Emissions From Stationary Gas Turbine Engines B133.9 Published: 1994 ASME Procurement Standard For Gas Turbine Electrical Equipment B133.5 Published: 1978 (Reaffirmed year: 1997) ASME Procurement Standard For Gas Turbine Auxiliary Equipment B133.3 Published: 1981 (Reaffirmed year: 1994) API Std 611, General Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services, 4th Edition, June 1997 API Std 613 Special Purpose Gear Units for Petroleum, Chemical and Gas Industry Services, 4th Edition, June 1995 API Std 614, Lubrication, Shaft-Sealing, and Control-Oil Systems and Auxiliaries for Petroleum, Chemical and Gas Industry Services, 4th Edition, April 1999 API Std 616, Gas Turbines for the Petroleum, Chemical and Gas Industry Services, 4th Edition, August 1998 API Std 617, Centrifugal Compressors for Petroleum, Chemical and Gas Industry Services, 6th Edition, February 1995 ANSI/API Std 670 Vibration, Axial-Position, and Bearing-Temperature Monitoring Systems, 3rd Edition, November 1993 API Std 671, Special Purpose Couplings for Petroleum Chemical and Gas Industry Services, 3rd Edition, October 1998 API Std 672, Packaged, Integrally Geared Centrifugal Air Compressors for Petroleum, Chemical, and Gas Industry Services, 3rd Edition, September 1996 API Std 677, General-Purpose Gear Units for Petroleum, Chemical and Gas Industry Services, 2nd Edition, July 1997, Reaffirmed March 2000 ISO 10436:1993 Petroleum and Natural Gas Industries—General purpose Steam Turbine for Refinery Service, 1st. Edition. Table 14-3: Mechanical specifications.

Manufacturers’ instruction books are often inadequate and need to be supplemented. The rewriting of maintenance manuals on such subjects as mechanical seals, vertical pumps, hot-tapping machines, and gas and steam turbines is not uncommon. The turbine overhaul manuals transferred on CDs could consist of (1) step-by-step overhaul procedures

| 582 |

| Maintenance Techniques | developed largely from the manufactures training school, (2) hundreds of photographs illustrating the step-by-step procedures on various types of gas and steam turbines, (3) an arrow diagram showing the sequences of the procedures, and (4) typical case histories. Detailed drawings on CDs are developed to aid in maintenance, such as a contact seal assembly, because the “typical” dimensionless drawing supplied by the OEM is not adequate to correctly assemble the compressor seals. Many other assembly drawings should be developed to facilitate the overall maintenance program. Videotaped programs are being developed on seals, bearings, and rotor dynamics, which will be a tremendous asset to most company maintenance programs. Practical Training The engineers in the maintenance group should be encouraged to gather pertinent vibration and aerothermal data and analyze the machinery. ASME performance specifications, which govern all types of plants and other critical equipment, are listed in Table 14-2. They should be encouraged to work closely at the various maintenance schedules and turnarounds so that they are familiar with the machinery. They should be sent to special training sessions where hands-on experience can be gained. After the completion of basic machinist training, the machinist should continue his training with on-the-job experiences. His skills should be tested, and he should be encouraged to take on different tasks. To develop the skills of in-house personnel, as much repair work as possible should utilize plant personnel. Encouraging the participation of the machinist in the solution of difficult problems often results in the machinist seeking information on his own. References to API and ASME specifications should not be uncommon on the shop floor. Today’s machinist and mechanic must be computer literate. Internet training must be provided with some basic training on word processing and spreadsheet programs. Basic Machinist Training Most of the basic training can be developed and conducted by in-plant personnel. This training can be highly detailed and tailored precisely to meet individual plant requirements. Training must be carefully planned and administered to fit the requirements of different machinery in the plant. Many plants have a full-time training program and personnel for conducting training at this basic level. Good maintenance practices should be inculcated into the young machinists from the beginning. They should

| 583 |

| Centrifugal Compressors | be taught that all clearances should be carefully checked and noted both before and after reassembly. They should learn the proper care in the handling of instrumentation and the care in placing and removing seals and bearings. A base course on the major turbomachinery principles is a must, so there is basic understanding of what these machines do and how they function. The young machinists should also be exposed to basic machinery related courses such as: 1) Reverse indicator alignment 2) Gas and Steam Turbine overhaul 3) Compressor overhaul 4) Mechanical seal maintenance 5) Bearing maintenance 6) Lubrication system maintenance 7) Single plane balancing

TOOLS AND SHOP EQUIPMENT A mechanic must be supplied with the proper tools to facilitate his jobs. Many special tools are required for different machines in order to ensure proper disassembly and reassembly. Torque wrenches should be an integral part of his tools as well as his vocabulary. The concepts of “finger tight” and “hand tight” can no longer be applied to high speed, high-pressure machinery. A recent major explosion at an oxygen plant, which resulted in a death, was traced back to gas leakage due to improper torquing. A good dial indicator and special jigs for taking reverse indicator dial readings is a must. The jigs must be specially made for the various compressor and turbine trains. Special gear and wheel pullers are usually necessary. Equipment for heating wheels in the field for assembly and disassembly are needed; specially designed gas rings are often used for this purpose. A maintenance shop should have the traditional horizontal and vertical lathes, mills, drill presses, slotters, bores, grinders, and a good balancing machine. A balancing machine can pay for itself in a very short time in providing a fast turnaround and accurate dynamic balance. Techniques to

| 584 |

| Maintenance Techniques | check balance of gear-type couplings for the large high-speed compressors and turbine drives as a unit should be developed. This leads to solving many vibration related problems. High-speed couplings should be routinely check-balanced. By dynamically balancing most parts, seal life and bearing life is greatly improved, even on smaller equipment. Dynamic balancing is needed on pump impellers, as the practice of static balance is woefully inadequate. Vertical pumps must be dynamically balanced; the long, slender shafts are highly susceptible to any unbalanced-induced vibration. This assembly and disassembly of rotors must be in a clean area. Horses or equivalents should be available to hold the rotor. The rotor should rest on the bearing journals, which must be protected by soft packing, or the equivalent, to avoid any marring of the journals. To accomplish uniform shrink fits, the area should have provisions for heating and/or cooling. A special rotor-testing fixture should be provided; this is very useful in checking for wheel wobbles, wheel roundness, and shaft trueness. Rotors in long-term storage should be stored in a vertical position in temperature-controlled warehouses. S PA R E PA RT S I N V E N T O RY The problem of spare parts is an inherent phase of the maintenance business. The high costs of replacement parts, delivery, and, in some instances, poor quality are problems faced daily by everyone in the maintenance field. The cost of spare parts for a petrochemical plant or a refinery runs into many millions of dollars. The inventory of these plants can run into over 20,000 items, including more than 100 complete rotor systems. The field of spare parts is changing rapidly and is much more complex than in the past. A group of plants have gotten together in a given region and formed part banks. Many pieces of equipment are made up of unitized components from several different vendors. The traditional attitude has been to look to the packaging vendor as the source of supply. Many vendors refuse to handle requests for replacement parts on equipment not directly manufactured by them. More and more specialty companies are entering the equipment parts business; some are supplying parts directly to OEM companies for resale as their “own” brand. Others supply parts directly to the end user. The end user must develop multiple sources of supply for as many parts as possible.

| 585 |

| Centrifugal Compressors | Gaskets, turbine carbon packing, and mechanical seal parts can be purchased from local sources. Shafts, sleeves, and cast parts can be purchased from local sources. Shafts, sleeves, and cast parts such as impellers, are becoming increasingly available from specialty vendors. All this competition is causing the OEMs to alter their spare parts system to improve service and reduce prices, which is definitely a bright spot in the picture. The quality control of both OEM and some specialty houses leaves much to be desired. In turn, this causes many plants to have an in-house quality control person checking all incoming parts, a concept highly recommended. INSPECTIONS As with any major equipment, a compressor train requires a planned program of planned inspections with repair or replacement of damaged components. Most plants follow the maintenance and inspection schedules suggested by the OEM. Condition and Life Assessment Condition and life assessment is significant for all types of plants. The most important aspect of a plant is high availability and reliability; in some cases, this is even more significant than higher efficiency. The availability of a piece of equipment is the percent of time the equipment is available to generate power in any given period. The reliability of the machinery is the percentage of time between planned overhauls. The availability of a machine is defined as A=

P− S − F P

(14-1)

where P = Period of time, hours, usually this is assumed as one year, which amounts to 8760 Hrs. S = Scheduled outage hours for planned maintenance F = Forced outage hours or unplanned outage due to repair. The reliability of a machine is defined as R=

| 586 |

P−F P

(14-2)

| Maintenance Techniques | Availability and reliability have a very major impact on the plant economy. Most of these compressor trains need to be available 24 hours per day. Planned outages are scheduled for non-peak periods and are planned years in advance. The centrifugal compressors usually do not contribute significantly in lowering the plant availability due to their robustness. The problems with centrifugal compressors, which are usually a series of impellers on a long shaft, are with the rotor assembly. Figure 14-5 shows the major components that create problems in a pump, such as problems with the rotor assembly, and the rotor impeller. The rotor assembly can suffer from problems due to the dynamic balance of the rotor, especially those that operate above the first critical, and also due to surge problems in the rotor impeller. Bearing failures are one of the major causes of failures in turbomachinery. The changing of various types of radial bearings from cylindrical and/or pressure dam babbitted sleeve bearings to tilting pad journal bearings is becoming common in the industry. In most cases, this gives better stability, eliminates oil whirl, and, under misalignment conditions, is more forgiving. Thrust bearing changes, from the simple, tapered land thrust bearings to tilting pad thrust bearings with leveling links (Kingsbury type), is another common area. These types of bearings absorb sudden load surges and liquid slugs. A major plant replaces the entire large journal and thrust bearings in their main machinery to tilting pad bearings in their plant as a matter of practice.

Figure 14-5: Contributions of various major components to pump down time.

| 587 |

| Centrifugal Compressors | Material changes of the babbitt are sometimes undertaken. Changing from the more common steel backed babbitted bearings to the copper alloys conducts surface heat away at a faster rate, thus increasing the load carrying capacity. In some instances, a 50-100% load carrying capacity improvement can be achieved. Some equipment manufacturers are offering bearing-upgrading kits for their machines in service. MAINTENANCE SCHEDULING The scheduling of maintenance inspections and overhauls is an essential part of the total maintenance philosophy. As the move from “breakdown” or “panic” maintenance towards a performance based total productive maintenance system takes place, total condition monitoring and diagnostics becomes an integral part of both operation and maintenance. Total condition monitoring and diagnostics examines both the mechanical and performance of the machinery and then carries out diagnostics. Condition monitoring systems, which are only mechanical systems without performance inputs, give less than half of the picture and can be very unreliable. Unscheduled maintenance is very costly and should be avoided. To properly schedule overhauls, both mechanical and performance data must be gathered and evaluated. As indicated earlier, repairs, must be conducted during a planned “turnaround” not “random” repairs which are frequently done on an “emergency” basis and where, due to time restraints, techniques are sometimes used which are questionable and should only be used in emergencies. To plan for a “turnaround,” one must be guided by the operating history of the given plant and, if it is the first “turnaround,” by conditions found in other plants utilizing the same or closely similar process and machinery. This is how the time between subsequent “turnarounds” has been extended to three years or more in many instances. By utilizing the operating history and inspection at previous “turnarounds” at this or similar installations, one can get a fair idea of what parts are most likely to be found deteriorated and, therefore, must be replaced and/or repaired and what other work should be done to the unit while it is down. It should be pointed out that, with modern turbomachinery, items such as bearings, seals, filters, and certain instrumentation, which are precision made, are seldom, if ever, repaired except in an emergency; such items are replaced with new parts. This means that parts must be ordered in advance for the “turnaround” and other work must be planned so that the whole operation may

| 588 |

| Maintenance Techniques | proceed smoothly and without holdups that could have been foreseen. This usually means close collaboration with the manufacturer or consultant and the OEM (or specialty service shop) so that handling facilities, servicemen, parts, cleaning facilities, inspection facilities, chrome plating and/or metalizing facilities, balancing facilities, and, in some cases, even heat treatment facilities are available and will be open for production at the proper time required. This is the planning, which must be done in detail before the shutdown with sufficient lead-time available in order to have replacement parts available at the job site. The old maxim “if it ain’t broke don’t fix it” is very applicable in today’s machinery. A study conducted at a major nuclear power facility found that 30% of the failures occurred after a major turnaround. This is why total condition monitoring is necessary in any performance based total productive maintenance system and why overhauls should be planned on proper data evaluation of the machinery rather than on a fixed interval. M A I N T E N A N C E C O M M U N I C AT I O N S It is not uncommon to hear the complaint that the maintenance department has “never been informed as to what is happening in the plant.” If this is a common complaint, the maintenance manager needs to examine the communications in his department. The following are six practical suggestions for improving communications: 1.

Operation and service manuals

2.

Continuous updating of drawing and print files

3.

Updating of training materials

4.

Pocket guides

5.

Written memos, inter office e-mails

6.

Seminars

7.

Website postings

Each of these items listed, if properly employed, can transmit knowledge to the person who must keep the plant’s machinery running. How well the information is transmitted depends entirely on the communication skills applied to the preparation of the materials.

| 589 |

| Centrifugal Compressors | Operation and Service Manuals To be of real value to the mechanic, the operation and service manual must be indexed to permit quick location of needed information. The manual must be written in simple, straightforward language, have illustrations, sketches, or exploded views adjacent to pertinent text, and have minimum references to another page or section. Major sections or chapters should be tabbed for quick location. Most often a mechanic or serviceman refers to a manual because of a problem. Problems seem to happen during a production run. It is essential, therefore, that he be able to find the needed information quickly. The mechanic should not be delayed by wordy, irrelevant text. The objective of any manual is to be an effective, immediate source of service information. The assignment of a non-technical person to write a manual is shortsighted and costlier in the long run. A well-written manual is continuously in use. Good manuals need not be complicated. In fact, the simpler the better. Manuals should be readable and understandable, whether they are compiled in-house or outside. Drawing and Print File A good print file is a vital tool for any maintenance organization. Reference files in a large or multi-plant company can be particularly burdensome for several reasons: •

Prints are bulky and difficult to store properly

•

Control of use is necessary

•

Files must be kept up to date

•

Handling and distribution of new or revised prints is usually expensive

A practical solution is to digitize the drawings and place them on CDs available to the maintenance and operation department. A good digital file reduces search time and helps the departments do a better job of keeping the machinery operating at their peak efficiency with minimal downtime. Training Materials Like any other written or audio-visual maintenance tool, training materials of all kinds are basically communication devices and, to be

| 590 |

| Maintenance Techniques | effective, should be presented in a simple straightforward, attractive, and professional manner. Once the need for specific maintenance training has been determined, a program must be developed. If the training need applies to a proprietary machine or one that is unique to a very few industries, it might be necessary to contact companies who specialize in custom digital programs on CDs, slide/tape, movie, videotape, or written training programs. The cost may shock the uninitiated, but, after shopping around, the company may find that it can recover far more than the initial cost in tangible benefits over a relatively short period. Training of an In-house Rotor Crew There are several advantages to having a well-trained rotor crew: •

Reduced cost of repairs because the owner can exercise more judgment factors than can be done at a distant factory. In addition, overhead costs are reduced.

•

Quicker repairs result in shorter outage times.

•

The rotor can be re-engineered to remove problem areas readily; for example, a redesigned thrust collar can be installed or an additional balance plane added.

•

In-plant skills are developed for emergency use.

Pocket Guide When a new maintenance form or procedure is introduced, a quick reference pocket guide can promote understanding and accuracy. The key to effectiveness is a deliberate design to provide maximum illustrations or examples in simple language. If it cannot be prepared in-house, outside help should be sought. Professionalism is essential to good communications. Written Memos One of the most effective devices for improving maintenance communications is a newsletter or internal memo. The memo’s success depends heavily on communicating formal tips and techniques in the mechanic’s language and using photos, sketches, and drawings generously to get the message across. Everyone in the maintenance department should be encouraged to contribute ideas on a better way to do a task or a solution to a nagging

| 591 |

| Centrifugal Compressors | problem related to the maintenance or operation of production equipment. Each contributor should be given credit by name and location for his or her effort. Very few workers can resist a bit of pride in seeing their names attached to an article that is seen by virtually everyone in the company. SEMINARS AND WORKSHOPS College or industry-sponsored seminars, continuing education courses, and workshops are means of upgrading or sharpening skills of maintenance people. Such an approach serves a two-fold purpose. First, it communicates the company’s good faith in the person’s ability to benefit from the experience, and, by acceptance, the worker shows willingness to improve his or her usefulness to the company. The seminars are very useful in disseminating knowledge. They also provide a forum for gripes and meaningful solutions. Discussion groups in these seminars and workshops are very important as participants share experiences and solutions to problems. The knowledge gained from these seminars is very useful.

COMPRESSOR PROBLEMS COMPRESSOR FOULING Centrifugal compressors, especially in the process gas applications, suffer greatly from fouling. Fouling is the deposit and the non-uniform accumulation of debris in the gas; this occurs due to the carryover of liquids and other debris from the suction knockout drums. This, in many cases, can also roughen the surface. Polymerization can occur also, due to changes in process conditions. In wet gas compressors, ethylene plant cracked gas compressors, and polyethylene recycle compressors the temperature of the gas must be kept below the critical temperature that would initiate the formation of polymers. The temperature varies with the various processes but generally should be kept below 235ºF to avoid the formation of these polymers. These problems can be further aggravated by deterioration of the stage seal clearances; this is especially true in compressors where there is a knockout drum in a series flow centrifugal compressor between the first and the second sections, as seen in

| 592 |

| Maintenance Techniques | Figure 14-6, since the gas from the first section leaks into the second section where the gas has been cleaned by going through a knockout drum but does suffer a pressure drop in the drum. The buildup will usually occur on the hub and the shroud with a larger buildup on the shroud at the elbow of the impeller on closed-faced impellers, as shown in Figure 14-7. There is also a build-up on the blades; the build-up is usually more on the pressure side than the suction side, and, in many cases, the build-up is the heaviest on the pressure side at the blade exit where there is also separation of the flow. Fouling in air compressors is due to the environment. This can be due to many reasons, such as from adjacent plants, highways, and salt in the atmosphere, especially in plants near the sea. The problems with sodium in the air can be very detrimental to the impellers.

Figure 14-6: Leakages in a labyrinth seal of a centrifugal compressor and contamination of gases.

| 593 |

| Centrifugal Compressors | TECHNIQUES TO PREVENT FOULING PROCESS GAS COMPRESSORS •

Condition monitoring of compressor aerodynamic and mechanical parameters. Regardless of the service, condition monitoring of molecular weight, pressure, and temperature will provide valuable information regarding process condition changes that

Figure 14-7: Areas of foulant build-up in a centrifugal impeller.

| 594 |

| Maintenance Techniques | can cause or increase fouling. Vibration monitoring could also alert the operator of fouling problems. •

Process control. Accurate control of process conditions can prevent fouling in applications where polymers can be formed. Control of temperature is usually the most important. The following applications can be affected by excessive process temperature: –

Ethylene cracked gas

–

Linear low density polyethylene

–

High density propylene

–

Fluid catalytic cracker off-gas (wet gas)

–

Thermal catalytic cracker off-gas (wet gas)

–

Coker gas

The temperature below which fouling can be prevented varies with each process, compressor, and application. Monitoring of process conditions is necessary to establish a threshold temperature in each case. In some cases, fouling cannot be prevented with the existing compressor. It may be necessary to modify the aerodynamic design and/or add additional cooling. O N L I N E S O LV E N T I N J E C T I O N Online solvent injection is very successful in various processes. The objective of this measure is to continuously inject a small amount of solvent to reduce the friction coefficient of the blade and impeller surface and thus prevent the fouling from accumulating on the surface. The injection should be done from the start, otherwise the foulant could be dislodged and moved downstream, creating a major problem. The downstream areas are much smaller, thus, foulant lodging there could create a major blockage problem. The object of solvent injection is often misunderstood. The idea is to prevent foulant accumulation, not clean the impeller or blade on-line. Non-continuous solvent injection will allow the impeller or blade surface to dry and thus promote fouling. Type and location of injection spray nozzles are critical factors in ensuring that the foulant does not form. Selection of the solvent is also critical in ensuring that the solvent works on the process gas. Knowledge

| 595 |

| Centrifugal Compressors | of solvent vapor pressure and internal compressor temperatures is necessary to determine if section or stage solvent injection is required. Spray nozzles are most effective when injecting directly into each stage of the compressor. These nozzles must have a very fine spray and should be at a pressure of about 20 psia above the pressure of the gas. This, in many cases, could mean an additional small compressor. Most solvents used are naphtha based. Many users have required flow meters at each injection point to determine the proper amount of injection. Typically, the total amount of solvent injection in one compressor case should be 1-2% of the mass flow. Excessive solvent injection could erode leading edge blade tips and create a problem near the shroud. This could cause impeller failure if not detected in time.

AIR COMPRESSORS I N L E T F I LT E R In air compressors and gas turbine applications, filter selection is an important factor in preventing fouling of the compressor blades as well as, in some cases, preventing problems in the gas turbine blades. This is due to the fact sodium is a major catalyst in hot corrosion problems in gas turbine blades. High efficiency air filters in most cases have a triple stage filtration system. Also, these filters often have rain shades to prevent water from entering the filters. Site conditions play a very important part in the selection of the filters. The first stage of the filters must prevent debris and excessive moisture from entering the air compressor. In platform applications and installations close to water, special designs must be used to prevent excessive chloride (salt) build-up. Compressor fouling in gas turbines can reduce output power by 3–5%, since the compressor consumes about 55–60% of the total power produced by the gas turbines. “Huff and Puff” filters that use a reverse air pulse to clean filter cartridges when filter pressure drop becomes excessive are used extensively throughout the industry.

| 596 |

| Maintenance Techniques | ONLINE COMPRESSOR WATER WASH Online compressor washing is used extensively in turbine operations and in some air compressors. The new high pressure compressors are very susceptible to dirt on the blades, which not only can lead to a reduction in performance but can also lead to compressor surge. If an on-line wash is to be used, it should be employed periodically, usually at least once a week, and begun when the rotor is in new (clean) condition. As described previously, the injected amount of wash fluid should not exceed 3% of mass flow. Process and compressor designers should be consulted regarding the exact procedure to be used. It is advisable that a radial vibration trip system be employed if on-line washing procedure is to be implemented. This action would prevent catastrophic compressor failure in the event of an incomplete on-line wash. Washing efficacy is site specific due to the different environmental conditions at each plant. There are many excellent techniques and systems for water washing. Operators must often determine the best approach for their gas turbines. This includes what solvents, if any, should be used and the frequencies of wash. Many operators have found that water wash without any solvent is as effective as with the use of solvents. This is a complex technical-economical problem also depending on the service that the gas turbines are in and the plant surroundings. However, the use of non-demineralized water could result in more harm than good. Water washing (with or without detergents) cleans by water impact and by removing the water-soluble salts. The effect of water cleaning is usually not very effective after the first few stages. It is most important that the manufacturer’s recommendations be followed with respect to water wash quality, detergent/water ratio, and other operating procedures. Water washing using a water and soap mixture is an efficient method of cleaning. This cleaning is most effective when carried out in several steps, which involve the application of a soap and water solution followed by several rinse cycles. Each rinse cycle involves the acceleration of the machine to approximately 50 percent of the starting speed, after which the machine is allowed to coast to a stop. A soaking period follows during which the soapy water solution may work on dissolving the salt. A fraction of airborne salt always passes through the filter. The method recommended for determining whether or not the foulants have a substantial salt base is to soap wash the turbine and collect the water from all drainage ports available. Dissolved salts in the water can then be analyzed.

| 597 |

| Centrifugal Compressors | On-line washing is being widely used as a means to control fouling by keeping the problem from developing. Techniques and wash systems have evolved to a point where this can be done effectively and safely. Washing can be accomplished by using water, water based solvents, petroleum based solvents, or surfactants. The solvents work by dissolving the contaminants while surfactants work by chemically reacting with the foulants. Water-based solvents are effective against salt but fare poorly against oily deposits. Petroleum-based solvents do not effectively remove salty deposits. With solvents, there is a chance of foulants being redeposited in the latter compressor stages. Even with good filtration, salt can collect in the compressor section. During the collection process of both salt and other foulants, an equilibrium condition is quickly reached, after which re-ingestion of large particles occurs. This re-ingestion has to be prevented by the removal of salt from the compressor prior to saturation. The rate at which saturation occurs is highly dependent on filter quality. In general, salts can safely pass through the turbine when gas and metal temperatures are less than 1000º F. Aggressive attacks will occur if the temperatures are much higher. During cleaning, the actual instantaneous rates of salt passage are very high together with greatly increased particle size. The following are some tips operators should follow during water washes: •

The water used should be demineralized. The use of nondemineralized water would harm the turbine.

•

On-line wash should be done whenever compressor performance diminishes by 2%–3%. It would be imprudent to let foulants build up before commencing water wash.

•

Stainless steel for tanks, nozzles, and recommended to reduce corrosion problems.

•

Spray nozzles should be placed where proper misting of the water would occur and minimize the downstream disturbance of the flow. Care should be taken that a nozzle would not vibrate loose and enter the flow passage.

•

After numerous water washes, the compressor performance will deteriorate and a crank wash will be necessary.

manifolds

are

After the use of many on-line water washes, the compressor usually requires an off-line wash, since the efficacy of the on-line washes is reduced. Off-line cleaning methods are definitely required and are more

| 598 |

| Maintenance Techniques | effective if a compressor is severely fouled. Failure to immediately remove all foulant from the rotor can cause catastrophic failure, since a large amount of unbalance can be instantaneously introduced in the rotor system. General guidelines for an off-line or “crank” washing procedure are as follows: •

Selection of proper solvent. This should be based on solubility test of foulant material.

•

Solvent properties should be well defined and the conductivity of the wash liquid noted.

•

Solvent should not be harmful to the compressor internals.

•

Warm solvent is often used.

•

Fill and vent case to be sure it is completely filled.

•

Compressor should be on turning gear.

•

Change wash fluid frequently and continue the wash until conductivity of wash liquid is equal to conductivity of initial values of the wash liquid.

COMPRESSOR BLADE COATING The main requirements of a coating are to protect blades against oxidation, corrosion, and cracking problems. Coatings are there to prevent the base metal from attack. Other benefits of coatings include thermal fatigue from cyclic operation, surface smoothness and erosion in compressor coatings, and heat flux loading, when one is considering thermal barriers. A secondary consideration, but perhaps rather more relevant to thermal barriers, is their ability to tolerate damage from light impacts, without spalling, to an unacceptable extent because of the resulting rise in the local metal temperatures. Coatings also extend life, provide protection by enduring the operational conditions, and protect the blades by being sacrificial by allowing the coating to be restripped and recoated on the same base metal. In the last few years, some users have investigated and used impeller and blade row coatings to reduce friction in lieu of continuous solvent injection. Based on the writer’s experience, this has met with mixed results. There has been success with coating impellers with Teflon-based material in ethylene cracked gas service.

| 599 |

| Centrifugal Compressors | Compressor coatings have been found to pay for themselves in energy savings in less than a year. This coating provides surface smoothness and prevents erosion on compressor blades. The coating provides a smooth surface and thus reduces the frictional forces on the blades as well as the pitting and spalling of the compressor blades.

COMPRESSOR FAILURES In the process industry there are three types of compressors that have very different maintenance problems. They are: 1.

Barrel-type Compressor

2.

Horizontal Split Casing Centrifugal Compressor with closed-face compressor.

3.

Air Compressors Integral Gear Type Compressor with open-faced impellers.

BARREL-TYPE COMPRESSORS Barrel-type compressors are being utilized in the process industry to an increased extent because the barrel design confines gases more effectively than horizontally split cases. This becomes a critical consideration in two areas: high pressure and low molecular weight gas compression. API-617, “Centrifugal Compressors for General Refinery Services” requires a barrel design based on the molecular percent of hydrogen contained in the process gas and the discharge pressure. Figure 14-8 details those requirements. The barrel design is essentially a compressor placed inside a pressure vessel. For higher pressures, some manufacturers have merely “beefed up” lower pressure barrel designs, while others have perfected unique designs such as the “shear ring” head design. All of these designs make extensive use of elastomer O-rings as sealing devices. There are several inherent maintenance problems with barrel-type compressors: Handling. Barrel-type machines must be removed from their foundations for total maintenance; many barrel machines weigh up to 30 tons, the handling problems becomes formidable. It is suggested that the

| 600 |

| Maintenance Techniques | shop be equipped with two 15-ton hoists six feet apart, to facilitate easy rigging. Hold-downs are provided in the floor for pulling the bundle from the barrel. Inner Casing Alignment. Since this type compressor consists of a bundle contained within the pressure walls of the barrel, alignment and

Figure 14-8: Compressor casing selection (Reference API-617).

positive positioning is often very poor, and the bundle is free to move to a certain extent. Bundle length is critical. Interstage leakage may occur if the bundle length is not correct, as shown in Figure 14-9. The sketch also shows bundle lengths supplied by the manufacturer after severe problems were encountered. Note the very critical tolerances. Assembly errors can be can be particularly severe in the case of a stacked diaphragm design, and care must be exercised to maintain proper impeller-diaphragm positioning. Since the bundle is subjected to discharge pressure on one end and suction pressure on the other, a force builds up that is transmitted from diaphragm to diaphragm, causing high loading on the inlet wall. The bundle lengths must be maintained carefully.

| 601 |

| Centrifugal Compressors | Internal leakage. A single O-ring normally separates the discharge and suction compartments of the inner bundle on a straight-through flow design. Compressors with side nozzles can have several bundles of Orings; excessive bundle-to-barrel clearance may cause leakage past the Orings. In addition, the O-ring at point “A,” as shown in Figure 14-9, are frequently pinched and cut as it passes across the suction nozzle opening in the barrel, a condition that is hard to prevent and doubly hard to detect if it occurs.

Figure 14-9: Bundle assembly installation.

Pressure differentials in excess of 400-500 psi, even using good design practice, can cause extrusion and failure of the O-rings. In many cases, back-up rings to the O-rings have been added to prevent failures. Grooves with O-ring ribbons have been added to the horizontal joints of the bundles of almost all of the machines to prevent interstage leakage.

| 602 |

| Maintenance Techniques | BEARING BRACKET ALIGNMENT In contrast to horizontally split compressors, where the bearing brackets are normally an integral part of the lower case half, barrel machines bearing brackets are bolted to the barrelheads. Both the bearing brackets and the head are removed during the disassembly operation, thus requiring all internal alignment to be re-established each time maintenance work is performed. This is a time consuming and exacting procedure that is not spelled out well in maintenance manuals. This procedure, commonly called “setting the lift,” must be done largely by “feeling” the rotor in the bundle. Experience on the part of the mechanic is very essential in this step. After the bundle and rotor are in place, the heads must be made up carefully to properly position the bundle and to align the bearing brackets. For bolt-on heads, the following procedures have proven to be satisfactory: •

Install a new head gasket and tape in place with “Scotch Tape.” Carefully align head to shaft and case studs. Push into bore.

•

Start and firmly drive up four diametrically opposite stud nuts. Remove hoisting equipment from the head. Run up remainder of the stud nuts. Sledge up on diametrically opposite nuts until all are equally tight and check out as follows: –

With a 0.0015-inch thickness gauge, check at least eight equidistant points between the head and the case. The gauge should read “no-go” at all of the points checked.

–

Check the same points with an 0.0010-inch thickness gauge. The gauge should read “go.” If necessary, sledge up on the stud nuts to achieve this clearance.

At this point there should be .010" to .015" clearance between the intake head and the inlet wall. If not, then the intake head must be machined on the “fingers” or the gasket surface to achieve this clearance. If there is too much clearance, the bundle will “leak” at each joint, reducing efficiency. Too little clearance can cause damage to the diaphragm sections. M AT E R I A L P R O B L E M S In order to limit the physical size of the case or pressure vessel and the rotor bearing span and to maximize the number of stages within the

| 603 |

| Centrifugal Compressors | heavy barrel, the gas path of a barrel is “squeezed” to a greater extent than in a horizontally split machine. This means the diaphragms and inlet guide vanes are intricate shapes with very small openings. Plain gray cast iron is normally used for these shapes because of casting ease and other economic reasons. The gray iron is not strong enough, in many instances, to withstand the pressure differentials imposed on them, resulting in failures. Inlet guide vanes have been especially troublesome. On several occasions, inlet guide vanes have been fabricated from wrought stainless and carbon steel materials. Replacement diaphragms and inlet guide vanes cast of nodular iron have also been used to alleviate some of these material problems.

COMPRESSORS IN GENERAL In addition to those problems inherent in barrel-type compressors, there are others that are common to all types of compressors. Some of the troubles and improved maintenance approaches could occur in these areas.

ROTOR THRUST CALCULATIONS Thrust loads in compressors, due to aerodynamic forces, are affected by impeller geometry, pressure rise through the compressor, and internal leakage due to labyrinth clearances. The impeller thrust is calculated, using correction factors, to account for internal leakage, and a balance piston size is selected to compensate for the impeller thrust load. The common assumptions made in the calculation are:

| 604 |

•

The radial pressure distribution along the outside of disc cover is essentially balanced.

•

Only the “eye” area is effective in producing thrust.

•

The pressure differential applied to the “eye” area is equal to the difference between the static pressure at the impeller tip, corrected for the “pumping action” of the disc, and the total pressure at inlet.

| Maintenance Techniques | These common assumptions are grossly erroneous and can be disastrous when applied to high-pressure, barrel-type compressors where a large part of the impeller-generated thrust is compensated by a balance piston. The actual thrust is about 50% more than the calculations indicate. The error is less when the thrust is compensated by opposed impellers, because the mistaken assumptions offset each other. The magnitude of the thrust is considerably affected by leakage at the impeller labyrinth seals. Increased leakage here produces increased thrust independent of balancing piston labyrinth seal clearance or leakage. The thrust errors are further compounded in design of the balancing piston, labyrinths, and line. API -617, “Centrifugal Compressors,” specifies that a separate pressure tap connection shall be provided to indicate the pressure in the balance chamber. It also specifies that the balance line shall be sized to handle balance piston labyrinth gas leakage at twice the initial clearance without exceeding the load ratings of the thrust bearing, and that thrust bearings for compressors should be selected at no more than 50% of the bearing manufacturer’s rating. Many compressor manufacturers design for a balancing piston leakage rate of about 11⁄2 –2% of the total compressor flow. Many users feel that the average barrel-type compressor, regardless of vendor, has a leakage rate of 3–4% of the total flow and the balance line must be sized accordingly. This design philosophy would dictate a larger balance line to take care of the increased flow than is normally provided. The balancing chamber in some machines is extremely small and probably highly susceptible to educting type action inside the chamber, which can increase leakage and increase thrust action. The labyrinth’s leakage should not be permitted to exceed a velocity of 10 feet/sec across the drum. The short balancing piston design of many designs results in a very high leakage velocity rate. Some of the balance piston leakage problems have been solved by the use of honeycomb labyrinths. The use of honeycomb labyrinths offers better control of leakage rates (up to 60% reduction of a straight pass type labyrinth). Honeycomb seals operate at approximately one-half the radial clearance of conventional labyrinth seals. The honeycomb structure is composed of stainless steel foil about 10 mils thick. Hexagonal shaped cells make a reinforced structure that provide a large number of effective throttling points, as shown in Figure 14-10. In addition, stainless steel honeycomb retains its strength at temperature and pressure levels that would cause weakening of an aluminum labyrinth.

| 605 |

| Centrifugal Compressors |

Figure 14-10: Schematic of honeycomb labyrinth seal used to control leakage rates.

Figure 14-11: Thrust forces in a centrifugal compressor.

| 606 |

| Maintenance Techniques | The leaks and the consequential pressure change across the balance piston destabilize the entire rotor system. A case in point was a four-stage rotor, as shown in Figure 14-11. The air to the right of the balance piston was low-pressure air from the compressor inlet. This air was originally bled off from the inlet to the compressor. After routine maintenance, it was decided that the air be taken from in front of the inlet air cooler rather than from behind it. This small change in pressure was enough to destabilize the system, and the rotor moved into the diaphragms where the heat generated, due to the diaphragm and the impeller clashing with each other, the welded the impeller shroud to the diaphragm, creating a major failure of the unit. THRUST COLLAR DESIGNS The thrust collar design of many compressors presents some problems. The minimum thrust capacity of a standard 8" (32.0 square inches) Kingsbury-type bearing with flooded lubrication at 10,000 rpm is well in excess of 61⁄2 tons. The thrust collar and its attachment method must be designed to accommodate this load. In most designs, the inboard bearing has a solid base ring, and the thrust collar must be installed after this thrust bearing is in place. The collar can be checked by revolving the assembled rotor in a lathe. The collar is removed subsequently for seal installation, and it must be checked for true, i.e., the face is normal to the axis of the bearing housing again after it is finally fitted to the shaft. The thrust bearing housing also generates a collar fit-up problem. In most cases a heavy puller cannot be attached to the collar outside diameter to remove it nor can heat be utilized because of space limitations. This limits the fit of the collar to about 1⁄2 to 1 mil loose. The poor fit of thrust collar to shaft, the use of a very small and weak collar key, and the use of a flimsy collar spacer and clamping sleeve arrangement all contribute to failures. In a failure experienced, the collar was broken in three pieces, in another, two pieces; in both instances, a corner of the keyway was involved in a fracture line. There have been several instances of thrust collars loosening on the shaft. These failures illustrate the reasons API-670 suggests two thrust position probes, one looking at the thrust collar and one at the shaft or both looking at the end of the shaft, as seen in Figure 14-12. As a solution, the thrust collar is frequently redesigned. A favorite design is an “ell” cross-section that permits increasing the fit area and, thus, collar rigidity. This is part of the bearing upgrade. The author is of the opinion that the best protection for the bearings and the

| 607 |

| Centrifugal Compressors | machinery from the movement of the axial shaft is the measurement of the rate of change of the thrust bearing metal temperatures. BEARING MAINTENANCE With high-speed machines, simple bearing failures are rare unless they are caused by faulty alignment, distortion, wrong clearance, or dirt. More common are failures caused by vibrations and rotor whirls. Some of these originate in the bearings; others can be amplified or attenuated by the bearings, the bearing cases, and the bearing support structure.

Figure 14-12: Actual probes for thrust-bearing monitoring.

During inspection, all journal bearings should be closely inspected. If the machine has not suffered from excessive vibrations or lubrication problems, the bearings can be reinstalled and utilized. Four places should be checked for wear during inspection periods: 1.

Babbitted shoe surface

2.

Pivoting shoe surface and seat in retaining ring

3.

Seal ring bore or end plates

4.

The shoe thickness at the pivot point or across ball and socket. All shoes should be within .0005% of the same thickness

While being inspected, the following checks should be made:

| 608 |

| Maintenance Techniques | •

All leading edges of shoes must have a uniform radius for the full length across the shoe. File the radii if necessary to obtain proper size.

•

Light scratches in the babbitt face do not necessarily require shoe replacement. If no wear is detected, scrape lightly with a sharp straight-edged scraper (plate type) to remove any upsetting caused by scratches.

•

Shoes should be replaced as sets only if: –

Radial clearance has increased more than one and one-half mils over nominal design clearance.

–

Leading or lagging edges of shoes show signs of wear.

•

The tilting-pad and support-ball combination spare parts should be lapped together, making them an integral unit. When a new or used bearing is disassembled for cleaning and inspection, care should be taken not to mix the tilting-pad and support-ball combinations.

•

On reassembly, care should be taken to return the tilting-pad and support-ball combination to the original location in the support ring. Changes in clearance and concentricity can result if the tilting-pad and support-ball combination is not returned to the same location. An eccentricity of as little as one mil can cause severe vibration problems.

CLEARANCE CHECKS •

Check housing OD and ID to be sure it is round.

•

Check bore and face-end plates for nicked edges, deep scratches, or scoring. Stone or scrape if necessary, and polish with very fine aluminum oxide polishing paper.

•

Check parting-line surfaces for full contact. Stone or lap if burrs or raised edges exist.

•

Check pivoting surfaces of shoe and housing ring for scratches, scoring, or erosion. Stone if necessary.

•

For tilting-pad bearings, blue-shoe the pivot surface, and check for contact area and position. The contacting surface must be in

| 609 |

| Centrifugal Compressors | the center only and at the bottom portion of the pivot bore in the retainer. •

Check to be sure that pins do not bottom-out in pads.

•

For ball-and-socket designs, check to be sure the ball seats properly and solidly in the counter bore.

•

Check for shaft clearance as follows: –

Select a stub mandrel in which the minimum diameter is the journal diameter plus minimum desired clearance (about 11⁄2 mil per inch of shaft diameter) and the larger diameter is journal diameter plus desired clearance (about 2 mils per inch of shaft diameter).

–

Assemble the bearing halves.

–

Slip the assembled bearing over the smaller diameter of the mandrel.

–

Tap the bearing lightly on the back of the housing, and slide the bearing down on the next larger diameter.

–

The mandrel should be rotated, and the OD of housing indicated.

THRUST BEARING MAINTENANCE Tilting pad-type thrust bearings are used in most major pieces of rotating equipment under the general term “Kingsbury Type,” as shown in Figure 14-13, and the rockers on which these bearing pads are placed are shown in Figure14-14. This type of bearing consists of pivoted segments or pads (usually six) against which the thrust collar revolves, forming a wedgeshaped oil film. This film, plus minute misalignment of the thrust collar and the bearing pads, causes movement and wear of the various bearing parts. The erroneous thrust calculations discussed earlier cause the bearing to be loaded heavier than desired. This accelerates the wear problem. There are seven wear points in the bearing, as shown in Figure 4-15.

| 610 |

1.

The soft babbitted shoe face

2.

The hardened steel shoe insert face (about Rockwell C 30-35 hardness)

| Maintenance Techniques | 3.

The face of the hardened steel upper leveling plate (about Rockwell C 47–50 hardness)

4.

The outer edge of the upper leveling plate

5.

The upper edge of lower leveling plate (about Rockwell C 47–50 hardness)

6.

The pivot point of the lower leveling plate (about Rockwell C 47–50 hardness)

7.

The inner face of the base ring (about Rockwell C 25–27 hardness)

Note: Hardness numbers are for a Kingsbury bearing. All of these points must be checked for wear. The leveling plates are normally surface hardened a few mils deep. Because the base ring is the softer component, it is likely to show the most wear. Also, a flat surface is more easily evaluated. Experience indicates that wear of about 6 mils here will cause “lock-up” of the leveling plates; therefore, by replacement of parts or a carefully supervised reworking of the entire bearing, wear in

Figure 14-13: Thrust bearing pads typical Kingsbury type bearings.

| 611 |

| Centrifugal Compressors |

Figure 14-14: Thrust bearing rocker typical Kingsbury type bearings.

Figure 14-15: Diagram of arrangement of tilting pad thrust bearing.

| 612 |

| Maintenance Techniques | excess of 11⁄2-2 mils at this point should be corrected. All shoe or pad thickness at the pivot point should be within .0005 inch of the overall thickness. The leveling plates wear is difficult to evaluate. Accumulated wear in excess of 25 mils (each bearing) in an opposed impeller, parallel flow catalytic cracking air blower can cause major problems. The normal thought is that thrust action is zero in this type of machine, yet wear is a problem. Because of the tremendous forces that are imposed on a thrust bearing, it must be in good shape. A thorough inspection may prevent machinery failure.

THRUST-BEARING FAILURE A thrust-bearing failure is one of the worst things that can happen to a machine, since it often wrecks the machine, sometimes completely. To evaluate the reliability of a thrust bearing arrangement, one must first consider how a failure is initiated and evaluate the merits of the various designs.

FAILURE INITIATION Failures caused by bearing overload during normal operation (design error) are rare today, but still far more thrust failures occur than one would expect, considering all the precautions taken by the bearing designer. The causes in the following list are roughly in sequence of importance. FLUID SLUGGING Passing a slug of fluid through a turbine or compressor can increase the thrust to many times its normal level—even if only a few gallons are involved. Instantaneous failures of the downstream bearing may result from fluid slugging.

| 613 |

| Centrifugal Compressors | BUILDUP OF SOLIDS IN ROTOR A N D / O R S TAT O R PA S S A G E S (“PLUGGING” OF TURBINE BLADES) This problem should be noticed from performance or pressure distribution in the machine (first-stage pressure) long before the failure occurs. O F F - D E S I G N O P E R AT I O N Especially from backpressure (vacuum), inlet pressure, extraction pressure, and moisture. Many failures are caused by overload, off-design speed, and flow fluctuations. COMPRESSOR SURGING Especially in double-flow machines. GEAR COUPLING THRUST A frequent cause of failure, especially of upstream thrust bearings. Thrust is high when alignment is perfect (friction coefficient 0.4–0.6), decreasing to a minimum when a small misalignment is present (about 0.1 at 25º angular misalignment). Friction increases rapidly again to 0.5 or more with an increase in misalignment. (These are rough numbers only, to show basic relationships.) The thrust is caused by friction in the loaded teeth that opposes thermal expansion. Therefore, thrust can get very high, since it has no relation to the normal thrust caused by pressure distribution inside the machine (for which the thrust bearing may have been dimensioned). The coupling thrust may act either way, adding to or subtracting from normal thrust. Much depends on tooth geometry and coupling quality. A straight-sided tooth can take misalignment only when the tooth fit has enough clearance to permit slanting of the male tooth inside the female teeth. For example, with vertical misalignment, the teeth on both sides will bind when the clearance is insufficient to allow for slanting. This can cause very high thrust; sometimes one can hear a “metallic sound” building up until the rotors finally slip with a very noticeable “bump.” Then the noise and vibration are gone, at least for a while. This phenomenon, of course, is torture for the thrust bearings, and it may cause failure in either direction. Dirt in the coupling can aggravate this situation or even cause it.

| 614 |

| Maintenance Techniques | DIRT IN OIL This is a common cause of failures, especially when combined with other factors. The oil film at the end of the oil wedge is only a small fraction of a thousandth thick. If dirt goes through, it can cause the film to rupture, and the bearing may burnout. Therefore, very fine filtering of the oil is required. But the best filter is no good if maintenance personnel leave the filter or bearing case open after inspection and the rain and sand blow in, or if they put the wet filter elements on the sandy floor or accidentally knock holes in the elements. It happens far too often. Once a machine is wrecked, it is difficult to reconstruct. M O M E N TA RY L O S S O F O I L P R E S S U R E Sometimes encountered while switching filters or coolers. To avoid this, a three-way filter system is recommended.

FAILURE PROTECTION Fortunately, accurate and reliable instrumentation is now available to monitor thrust bearings well enough to assure safe continuous operation and to prevent catastrophic failure in the event of an upset to the system. Temperature sensors, such as RTDs (resistance temperature detectors), thermocouples, and thermistors can be installed directly in the thrust bearing to measure metal temperature. The installation shown in Figure 1416 has the RTD embedded in the babbitted surface. It is in the most sensitive zone of the shoe—70% from the leading edge and 50% radially. The position of the sensor is critical in establishing the safe operating limits. As long as the probe is generally in the zone of maximum temperature, it will be highly sensitive to load, although the level of temperature may vary considerably, as can be seen in Figure 14-17. The temperature is also dependent on the pad-backing material. At 500-psi load, the center sensor at A-II registers 200ºF while the sensor at B-I registers 280ºF in a steel-backed bearing. Again, these temperatures are typical and will vary with size, type, speed, and lubrication from bearing to bearing. The difference in a copper-backed bearing can be seen to be quite significant, with A-II reading 185ºF and B-I reading 205ºF. The position of the sensor with respect to the surface is less significant in this bearing than in the steel-backed bearing. Again, position in the sensitive zone is important in establishing safe operating limits with respect to temperature.

| 615 |

| Centrifugal Compressors |

Figure 14-16: RTD embedded in bearing surface.

Figure 14-17: Temperature distribution in bearing surfaces.

| 616 |

| Maintenance Techniques | Axial proximity probes are another means of monitoring rotor position and the integrity of the thrust bearing, as mentioned earlier. This method detects thrust collar runout and also rotor movement. In most cases, this ideal positioning of the probes is not possible. Many times the probes are indexed to the rotor or other convenient locations and, thus, do not truly show the movement of the rotor with respect to the thrust bearing. A critical installation should have the metal temperature sensors in the thrust pad. Axial proximity probes may be used as a backup system. If metal temperatures are high and the rate of change of those temperatures begins to alter rapidly, thrust bearing failure should be anticipated.

IMPELLER PROBLEMS The high-speed rotation of the impeller of a centrifugal compressor imparts the vital aerodynamic velocity to the flow within the gas path. The buffeting effects of the gas flow can cause fatigue failures in the conventional fabricated shrouded impeller due to vibration induced alternating stresses. These may be of the following types: •

Resultant vibration in a principal mode.

•

Forced, undamped vibration, associated with aerodynamic buffeting or high acoustic energy levels.

The vibratory mode most frequently encountered is of the plate-type and involves either the shroud or the disc. Fatigue failure generally originates at the impeller outside diameter, adjacent to a vane, often due to the vibratory motion of the shroud or disc, as shown in Figure 14-18. The fatigue crack propagates inward along the nodal line, and, finally, a section of the shroud or disc tears out. To eliminate failures of the covered impellers operating at highdensity levels, the impellers are frequently scalloped between vanes at the outside diameter, as shown in Figure 4-19. The consequent reduction in disc friction also causes a small increase in impeller efficiency; however, it does not seem to affect the overall efficiency, due to higher losses in the diffuser. In fact, there may be a slight reduction in efficiency. The reduction in efficiency, however, is very small and the advantage from a mechanical point very large.

| 617 |

| Centrifugal Compressors |

Figure 14-18: Vibration in the covered impeller shroud and disc.

Figure 14-19: Scalloped rotor.

| 618 |

| Maintenance Techniques | Several rotors have been salvaged by scalloping the wheels after a partial failure has occurred. To accomplish this, the rotor was un-stacked. Each wheel was set up in the milling machine and scalloped. Then each wheel was individually balanced on a mandril. The rotor was restacked and the machine returned to service in slightly over a week’s time. There are some OEMs that have new impellers designed with the scallops as part of the original design.

COMPRESSOR SEALS The extent of the leakage past the seals where the shaft comes through the casing frequently limits the running time of the compressor, yet the seals and the seal systems are not given adequate treatment in the maintenance manuals or the operating instructions furnished by the compressor manufacturer. API Standard 617, “Centrifugal Compressors for General Refinery Services,” divides shaft seals into the following categories: •

Labyrinth

•

Restrictive carbon rings

•

Mechanical (contact) type

•

Liquid film or floating bushing type

•

Liquid film type with pump bushings

The first two seal categories are usually operated dry, and the last three categories require seal oil consoles, either separately or as part of the lube system. While each of these seal designs has their own characteristic maintenance difficulties, a discussion of the liquid film bushing type will illustrate some of them. First, a review of how the seal functions is in order. A liquid film or bushing seal is simply a close-clearance sleeve surrounding the shaft. Sealing is accomplished by two sleeves which normally are pressurized at a midpoint with 2–4 gpm of seal oil, about 5–15 psi greater than the process gas chamber pressure. Seal oil flows in two directions—through the inboard sleeve to a high-pressure drain area and out through a trap (sour oil) and through an outboard sleeve to an atmospheric drain (sweet oil). The inboard seal oil will absorb some process gas in the drain area; therefore, the inboard oil flow must be kept down to a few gallons per day if a separate oil system is not used. The bulk

| 619 |

| Centrifugal Compressors | of the oil circulation passes through the outer sleeve which is somewhat longer than the inner sleeve. This breaks down the greater pressure differential. Since all the oil returned to the sweet oil system passes through the sleeve, the oil temperature will be much higher than a bearing because of extreme shearing of the oil film. The sleeves are usually made of babbitt-lined steel with 0.002–0.006 inch diametral shaft clearance. The sleeves must float radially, that is, they must follow the shaft movements. Therefore, the sealing face between the end of the sleeve and its housing is very important. The entire assembly, sleeves and housing, is sealed into the casing with suitable O-rings and gaskets to avoid gas leakage and to conduct seal oil to proper compartments. These concepts, oil flows, and critical clearances are not spelled out well in either the operating instructions or maintenance manuals. Because of this, several maintenance technique improvements are needed. RADIAL CLEARANCES Radial clearance between the bushing and the shaft and the length of the bushing must be selected to obtain minimum leakage without exceeding fluid temperature limitations. The clearance recommended by some equipment manufacturers result in exceeding good design temperature levels. Often equipment is run with greater clearances than those recommended by the manufacturer in order to provide good lubrication and cooling for the bushings. QUALITY CONTROL The flatness, parallelism, and surface finish of the mating sleeve faces must also be carefully controlled to obtain maximum seal effectiveness. Poor quality control by the manufacturer over these parts requires that each part be carefully checked and frequently remanufactured. Such simple things as an incorrect O-ring groove depth can cause malfunctioning of the total seal. AXIAL CLEARANCES Axial clearance between the bushing or sleeve and the housing is critical but is completely ignored by most manufacturers. There should be

| 620 |

| Maintenance Techniques | 12-15 mils clearance per bushing between the bushing or sleeve and the housing. Where the sleeves are mounted back-to-back, there will be 25 to 30 mils clearance total for the seal. SEAL DESIGN In higher-pressure seals, more than one outboard (i.e., high differential) sleeve may be used. Generally, it is desirable to use a single sleeve because the inboard sleeve operates with up to 80% of the total pressure drop across it. The outer sleeve with the lower differential causes lubrication and cooling problems that can shorten the life of one or both sleeves. In some cases, the inner bushings are altered so as to allow more oil flow to the outer sleeve. TRAINING FOR SEAL MAINTENANCE Guidelines should be explicit in indicating oil flow rates and the interaction of various components. The oil flow rates, vital to the operation of the seal, are usually buried in a single drawing (the combined lube and seal layout) in the manuals supplied with the machines.

RULES OF THUMB There are a few rules of thumb that help in understanding seal operation and maintenance: •

•

The oil flow rate will vary as follows: –

Directly with the differential pressure and the wetted perimeter of the sleeve

–

With the cube of the radial clearance

–

With the square of the eccentricity of the sleeve and shaft

–

Inversely with oil viscosity, temperature, and length of the sleeve

Shear work done on the sealing fluid during its passage through the sleeve raises its temperature to a much higher level than may be expected.

| 621 |

| Centrifugal Compressors | ROTOR REPAIRS Repair work on the rotating elements of compressors and turbines has traditionally been the field of the original equipment manufacturer, not the equipment owner. Many experiences have pushed users into rotor repair in order to improve reliability and availability of the rotors and the machinery. Returning a rotor to a vendor for repairs could mean a 12–15 month period, without a backup rotor in a critical process. This would be intolerable, and, thus, users have been forced to find other than OEM shops. The decision to send the rotor to the original equipment manufacturer (OEM), or to a local machine shop with rotor repair capabilities is made by calculating the ability of shops. The time and cost are usually the controlling factors. In fact, the turnaround time is the governing factor in most cases. R E PA I R S I N FA C T O RY When the rotor is returned to a vendor’s factory, a consultant is often hired to be an on-site representative. His services can be shared with other companies. This has been a successful approach as it gives considerable expertise available at the factory over long periods of time without depleting experienced manpower from the users’ plants. A detailed repair list is part of the purchase order, and both the consultant and the vendor provide reports. R E PA I R S I N L O C A L S H O P S When the rotor is sent to a local shop, the rotor crew foreman makes periodic visits to analyze the work progress and the quality of the repairs. When needed, he can call upon the services of several consultants within the company as well as his supervisors to aid in the decision-making. Many local shops have excellent capabilities for rotor repairs. Impellers have been spin-tested in local facilities to prove their integrity. Local highspeed balancing facilities have also been utilized to check rotor dynamics characteristics in troublesome cases.

| 622 |

| Maintenance Techniques | R E PA I R S I N U S E R ’ S S H O P The repair work at the user’s shop varies depending on the equipment and manpower available. The most prominent features are a stacking pit about eight feet deep with a hydraulic service station lift in the bottom. Use of this lift keeps the working height of the rotor about 36 inches off the floor at all times during the restacking operation. Balancing machines should be nearby, and the shop should have drill presses, lathes, and presses. REBLADING The reblading of steam turbine blades (especially the last few stages) due to water erosion is a problem that needs to be addressed in a routine manner. A simple leading edge repair can be accomplished by inserting a protection strip on the leading edge.

COMPRESSOR RERATING Compressor rerating is often required in the process and petrochemical industries, to meet the design flow rate and/or head. This is due usually to change in molecular weights of the process and, in the case of re-injection compressors, to the change in the suction pressure of the compressor. These changes, in many cases, need the compressor to be “redesigned.” This includes anything from a simple speed change to removal or trimming of impellers. In the case of the vaned diffuser being used in the compressor, the compressor surge margin is greatly reduced because of changes in molecular weight and suction pressures. This is due to the angle of exit from the impeller having a large angle of attack on the diffuser blade. In this section, some ways to increase the head and flow of existing compressors are discussed. Compressor rerating requires an in-depth knowledge of the compressor and its various components. In this section, some of the highlights that would be required to fully understand the mechanism of the changes to be incorporated are discussed. No compressor rerate would be complete without a full mechanical review since the changes in flow and pressure delivered change the rotor dynamic characteristic as well as the thrust loadings. This may require a

| 623 |

| Centrifugal Compressors | change in the bearings or the balance piston flows so that upset conditions can be avoided. Shortcuts can get the operator into a lot of trouble. The extra money spent in the proper evaluation up front will pay major dividends in the reduction of start up as well as operational problems on compressor rerates.

CHANGES IN INLET PRESSURE OR TEMPERATURE Changes in the inlet conditions of the gas affect the head delivery capacity of the system. Changes in temperature are more significant than changes in pressure. An increase in temperature will decrease the head delivery, while a decrease in inlet temperature would increase the head delivered. Figure 14-20 shows the compressor maps on a corrected basis. It is obvious from this that, as we changed temperature, we would be moving from one aerothermal speed line to the other. Thus, for a compressor operating at a constant mechanical speed, changes in inlet temperature would be tantamount to changes in speed. The effect of pressure changes due to the environmental changes is more subtle.

Figure 14-20: A centrifugal compressor performance map corrected for temperature and pressure.

| 624 |

| Maintenance Techniques | INCREASE OR DECREASE IN THE HEAD DELIVERED The requirements for a higher head or lower head at the same flow would require changes in machine speed and, in some cases, the trimming of the wheels. Figure 14-21 shows the effect of speed changes on the head delivered. The unit would operate along the system characteristic curve. This curve is a function of the frictional characteristics of the system. Trimming the compressor wheels would also reduce the head produced, as seen in Figure 14-22. A speed increase after the wheels have been trimmed would then increase the head delivered at a higher flow. Removal of a wheel would also reduce the head delivered. This latter technique is used more when larger flow increases are required. A stability analysis will indicate that, as the number of impellers increases, the stability decreases and with it comes a reduction in the capacity limit. In most machines, there will be only two, or, at the maximum, three different wheel designs. Therefore, even though the first impeller is operating at design, the other impellers would be operating at off-design conditions. Figure 14-23 shows the effect of the number of wheels on both the stability and capacity limit. It should be noted that, just as the number of wheels in a casing reduce the stability, so does the number of sections of compression or the number of casings in series.

Figure 14-21: A centrifugal compressor map showing operational characteristics.

| 625 |

| Centrifugal Compressors |

Figure 14-22: Effect of trimming the exit diameters of the centrifugal compressor impellers.

Figure 14-23: Effect of number of impellers on stability and capacity limit.

| 626 |

| Maintenance Techniques | INCREASE OR DECREASE IN FLOW DELIVERED The change in flow requirements is the most common requirement for a compressor rerate. When the flow requirements are decreased at a constant head, the compressor will move towards surge and may surge. To avoid this, the compressor is placed in a recirculation mode; this is a technique to “fool” the compressor into believing it is delivering more flow than it really is. An intercooler must be introduced in the recirculation line so that the recirculated gas does not heat up the incoming gas. Heating up the incoming gas will cause the compressor to surge earlier, thus requiring more recycling, a problem which can build on itself. A very responsive surge control system is required in the cases where the flow is reduced significantly to necessitate recirculation. Surge control systems that require a ten to fifteen percent margin could cost the operator millions of dollars in wasted energy costs compared to a more sensitive system that can operate the unit within 2–4% of surge. Increase in flow requirements cause the unit to be operating in a choke region. In many cases, the solution for such a problem can be the removal of a stage. The head delivered by the compressor will be reduced, however, the speed line characteristic will be flatter, as seen in Figure 14-24. After removal of the wheel, increasing the speed will result in a head increase at

Figure 14-24: Effect of removal of wheel from a centrifugal compressor.

| 627 |

| Centrifugal Compressors | a higher flow rate, thus delivering more flow at the original head, as seen in Figure 14-25. The removal of the wheel can be accomplished by placing a “bridge” to direct the flow from one diffuser to the next, as shown in Figure 14-26. In some cases, a dummy wheel could be placed, as seen in Figure 1427. Since these techniques require speed increases, care must be taken to ensure that wheel stresses are not exceeded.

Figure 14-25: Increase in flow after removal of wheel and speed up.

Another way to increase the flow is by using an inducer on each wheel or increasing the throat areas in the inducer. This is a much more costly way to achieve the goals but may be necessitated due to the restrictions on the increase in the speed. In many cases, the changes in the first few stages are the only ones required. Note that new impeller designs require much care. Area changes in the impeller are critical. Changes must be smooth; otherwise, acceleration or deceleration of the flow will cause separation and early surge. Diaphragms will also have to be changed out in the first few stages to incorporate these new designs. Metallurgical properties of the wheel must be clearly stated. Heat treatment must be clearly specified. Improper heat treatment has led to many impeller failures. When dealing with stainless steel AISI 405, one must be extra careful, because many shops cannot handle this material. Throttling the inlet, however, causes a major reduction in the head without affecting the surge point, as shown in Figure 14-28.

| 628 |

| Maintenance Techniques |

Figure 14-26: Removal of a wheel from a centrifugal impeller and replacing it with a bridge.

Figure 14-27: Multistage centrifugal compressor.

| 629 |

| Centrifugal Compressors |

Figure 14-28: The effect of inlet throttling on a centrifugal compressor.

COUPLING MAINTENANCE The major inspections should also include detailed inspections of any couplings in the train. Gear couplings should be disassembled and teeth inspected for indications of problems. The most common failures encountered with continuous lubrication type gear couplings are as follows:

| 630 |

•

Wear

•

Corrosive wear

•

Coupling contamination

•

Scoring and welding

| Maintenance Techniques | Couplings with sealed lubrication systems tend to have wear problems similar to couplings with continuous lubrication, but they must also be checked for fretting corrosion and cold flow. These problems result from normal coupling operation. If for some reason excessive misalignment exists, additional damage can be revealed by tooth breakage, scoring, and pitting. Disc couplings should be checked to ensure there are no cracks in the discs or connecting shaft. If the damage to the disks can not be repaired, new disks should be used and the coupling should be rebalanced prior to installation.

REPAIR AND REHABILITATION OF TURBOMACHINERY FOUNDATIONS In many instances, vibration problems in turbomachinery can be attributed to faulty support. Once the problem areas have been identified, correcting defects can be a logical procedure. What is novel is that this result can often be accomplished through the proper selection and application of adhesives. Most turbomachinery is mounted on structural steel platforms sometimes referred to as base plates or skids. These platforms are then installed on a mass of concrete at the job site (either by direct grouting or mounting on sole plates) to become the machinery foundation. Platforms should always be considered as part of the foundation rather than as part of the machinery. Problems with platforms fall into one or both of the following categories: •

Improper installation

•

Insufficient mass and/or rigidity

Improper installation is not a design weakness. This defect can be corrected rather easily in the field at any time after installation. Insufficient mass or rigidity is a design weakness brought about by the complexity of the origin of vibration in high-speed rotating machinery and its sensitivity to vibration. Nevertheless, mass and rigidity can be increased in the field, but it is more of a task to do so than the mere correcting of installation defects.

| 631 |

| Centrifugal Compressors | INSTALLATION DEFECTS A typical compressor train containing a turbine and two compressor stages is shown in Figure 14-29. The I-beams on the platform are grouted to the concrete structure. When proper grouting techniques are carried out during the original installation, the grout should contact the entire lower surfaces of all longitudinal and transverse I-beams. Cement-based grouts will not bond well to the platform load-bearing surfaces. Over a period of time, lubricating oils will severely degrade both cement groups and concrete. This problem is further aggravated because most platforms are not designed with oil drains. On several occasions, as much as six to eight inches of oil has been found trapped within the platform cavities. This condition not only provides head-pressure for an increased rate of oil penetration but also creates a severe fire hazard. All platforms, regardless of the type of grout to be used, should be designed with oil drains. Epoxy grouts are recommended on platform installations because they provide an excellent oil barrier for the concrete below. Cement grouts should be used only for temporary installations. When differences in vibration amplitudes can be detected between the lower flange of the platform beams and the concrete structure, the decision to bond the entire lower surfaces of the platform to the concrete structure should be made. Bonding can be accomplished using a technique known as pressure grouting. With this technique, holes are drilled through the lower flange at locations near the web on centers of approximately 18 inches. These holes are then tapped, and ordinary grease fittings are installed.

Figure 14-29: A typical compressor train containing a turbine and two compressor stages.

| 632 |

| Maintenance Techniques | Pressure grouting can then be carried out with either automatic injection equipment or with conventional grease guns. Some manufacturers recommend that their platforms be installed on rails or sole plates, which have been grouted to the concrete foundation. Occasionally the installation will be either poorly designed, or the contractor will fail to clean the plates before grouting. Loss of adhesion may result in excessive vibration or movement of the plate in the grout. When this problem occurs, pressure grouting can be accomplished with a relatively high degree of success if proper techniques are used. The following are some main points to consider when designing and grouting a sole plate. •

Check to see that the block between the equipment base and the sole plate is adequate to transmit the load.

•

Corners on the edges of the sole plate should have at least a 2-inch radius to prevent the creation of stress risers and subsequent cracking of the corners.

•

There should be a sufficient amount of aggregate in the epoxy mixture. Insufficient quantities of aggregate will lead to a layer of unfilled epoxy on the surface of the mortar. The linear coefficient of thermal expansion of the unfilled epoxy can be expected to be on the order of magnitude of 6-8 x 10-5 inches per inch of thickness per ºF. The linear coefficient of thermal expansion for the epoxy mortar below can be expected to be on the order of magnitude of 2 x 10-5 inches per inch of thickness per ºF. This difference in thermal expansion rates will encourage crack propagation, particularly on cooling when the system is subjected to cyclic temperatures such as between day and night.

•

Make sure that a foamy surface does not exist immediately below the sole plate. A foamy surface is caused by an insufficient quantity of aggregate in preparing the epoxy mortar. The epoxy adhesive has a density of about nine pounds per gallon. The aggregate has a bulk density of about 14 pounds per gallon, which assumes about 25–30% voids. In preparing an epoxy mortar, the resin and hardener components are always mixed together before the addition of aggregate. When the aggregate is added to the mix, it obviously falls to the bottom and introduces air into the mix. If a soupy mortar is prepared, the air will simply rise to create a weak, foamy surface.

| 633 |

| Centrifugal Compressors | INCREASING MASS AND RIGIDITY When excessive vibration is detected in the gear box of a compressor train, as seen in Figure 14-29, the force is transferred to the platform below, a dampening effect can be created by increasing the rigidity of the support below. This effect can be accomplished by first filling the platform cavity and then the gearbox support with epoxy mortar. When the turbine and compressor support pedestals are such that little cross-sectional area is provided, it is doubtful that appreciable stiffening can be obtained. Consequently, the objective is to concentrate on increasing the mass of the pedestals. This increase is accomplished by filling the cavities with a special mortar prepared with epoxy and steel shot. The density of this special mortar can be in excess of 300 pounds per cubic foot. To inject this special mortar, a pipe has been installed in the access hole that was drilled in the side of the pedestal near the top. These same techniques can be employed to stabilize the foundations under much smaller equipment.

GEARS AND GEAR BOXES The mounting of a gearbox into trains is a precision job that should be performed carefully. Gear unit installation is one of the most important factors to be considered for long, trouble-free operation. No matter how accurately the gear unit is manufactured, it can be destroyed in a few hours of operation when improperly installed. The same care should be taken when installing a gearbox as with any high-speed machinery. The mounting surface should be a flat, level, single-plane surface of finished steel at a height that will permit the shimming necessary to align the gear unit properly to connecting shafts. The shims should be of a size at least equal to the width of the unit footpad. Then the gear unit should be placed on the foundation in the approximate required position. Uneven supports can distort the gearbox and adversely affect the gear tooth contact. Shaft alignment is very important for long gear life. Poor alignment can cause unequal distribution of tooth loads and distortion of the gear elements from overhung moments. A 2.0 mil shaft vibration level on the gear unit produced by misalignment is equivalent to a gear pitch-line runout of 2.0 mils. The spectrum signature plots at synchronous speeds and highfrequency spectra reveal an interesting set of information. A high-running

| 634 |

| Maintenance Techniques | speed amplitude can indicate problems such as unbalance. The spectrum showing this unbalance is in Figure 14-30. Misalignment problems can be analyzed also. Figure 14-31 shows a plot obtained from a casing-mounted pickup and the classic high, twice-per-revolution radial vibration. A second order component of the running speed taken from a probe mounted in the axial direction is nearly a 100% sign of misalignment.

Figure 14-30: A typical unbalance signature plot.

Figure 14-31: A typical misalignment signature plot.

| 635 |

| Centrifugal Compressors | The gear housing must be properly supported to maintain proper internal gear alignment. When a gear unit is installed, the support pads must be maintained in the same plane as used by the manufacturer during assembly when gear face contact was obtained in the plant. Before startup, gear face contact should be checked using high-spot bluing and rotating or rocking the pinion or lighter element back and forth sharply within the confines of the backlash. Inspection of this blued area should show approximately 90% face contact. If this contact is not obtained, the gear housing can be shimmed under the proper corner until an acceptable face contact is achieved. Many large, high-hardness or wide-face width gears are manufactured with helix angle modifications to account for torsional and bending deflection. When the helix angle has been modified, good face contact will not be obtained under light load. In this case, the gear supplier should furnish data on percent of face contact versus load to be used as a guide during installation and startup. Also, many gears have a short area of easeoff on each end of the teeth to prevent end loading, and this area usually will not show contact under light load. The larger the gear unit, the more important this check becomes, since large housings tend to be more flexible. Also, the use of baseplates furnished by the original equipment manufacturer does not eliminate face contact problems, and these inspection procedures should be carried out. After the gear checks have been made, the foundation bolts should be uniformly tightened, and the alignment rechecked. It may be necessary to repeat the shimming and tightening of foundation bolts to obtain final, correct, cold alignment. Alignment of high-speed gear units should always be hot-checked and adjustments made as necessary. Temperatures vary so greatly throughout the housing and shafting that it is impossible to calculate a thermal growth accurately, and, therefore, an alignment check must be made in the hot condition. When the alignment is complete, the baseplate or bed should be grouted in as close to the gear housing as possible. Journal bearings are used on the gear shafts and proper oil flow must be maintained. The oil system should therefore be checked thoroughly prior to startup. The gear lube system is normally flushed prior to any operations. The usual procedure is to seal off the gearbox components to which acid would be harmful, acid flush the system, and follow with a neutralizing flush before filling with the lube oil.

| 636 |

| Maintenance Techniques | Gear mesh spray nozzles should be checked to be sure dirt was not pumped through the system by observing the sprays or by introducing high-pressure air into the spray nozzle manifold. When possible, gears should be run-in on initial startup. Speeds and loads should be increased in percentage increments. Lube oil temperature and pressure and bearing temperatures should be observed and adjustments made to the lube system, as required. The number of adjustments made will depend on the complexity of the system. Oil pressure is of primary importance. When an auxiliary pump has been provided, oil should be circulated before the actual start. If not, the pump should be primed, and the journals wetted with oil. Primer holes are sometimes provided or alternate journals can be oiled through the holes provided for bearing temperature detectors. It is recommended that warning devices be provided to eliminate as much human error as possible, and the set points should be checked carefully. As with any startup, vibrations should be monitored and recorded. The vibration monitoring system should include at least one accelerometer to detect any vibrations generated at gearmesh frequency. The recorded data should be saved to provide baseline vibration data for future reference. The use of accelerometers for diagnosing problems is very effective, since, in many cases, the high frequency spectra give much more information than the low-frequency spectra obtained from proximity probes. An example can be seen in Figure 14-32, which has been based on

Figure 14-32: Gear box signature using a proximity probe showing only the low end frequency component.

| 637 |

| Centrifugal Compressors | using proximity probes; the two gear signals are alike, and, based on that, one would conclude that the two gear drives are in good mechanical condition. Figure 14-33 shows the high-frequency accelerometer signatures. These indicate a problem with gear A, indicating a cracked or chipped tooth. Thus, it is obvious that gear mesh problem accelerometers should be used. Shutdowns, as well as startups, require care and attention. Shutdown of a unit that has been operating in a humid atmosphere can result in considerable condensation and subsequent rusting of the gears, shaft journals, and housing in a very short time. When water contacts clean steel, it begins to etch the steel immediately. When shutdowns in such conditions are necessary, provisions to prevent condensation must be furnished. Under normal operating conditions, the oil should be changed every 2500 operating hours or every six months, whichever occurs first. Where operating conditions warrant, this period may be extended; conversely, severe operating conditions may make it necessary to change oil at more frequent intervals. Such conditions may occur with the rapid rise and fall in temperature, which produces condensation when operating in moist or dusty atmospheres or in the presence of chemical fumes. In any case, the lubrication supplier should be consulted when determining a lubricant maintenance program. It may be necessary to analyze the oil periodically until a reasonable program can be established.

Figure 14-33: A gear box signature using an accelerometer thus examining the high frequency component.

| 638 |

| Maintenance Techniques | Gearing problems are due to case distortion, improper gear cooling, or high backlash on the gears. Also, misalignment is a great contributor to this problem. Gears should be checked for proper fit; in some cases, lapping is advised. Care is always needed to prevent lapping compound from entering the lubrication system and bearings. Cooling of the highspeed gears is accomplished by directing a jet of lubrication oil on the gears as they become unmeshed. In very high-speed applications, oil should be directed at the casing to reduce thermal distortion of the casing. Coupling problems are also directly or indirectly caused by improper lubrication or a high level of misalignment. Couplings of the gear type should have a continuous lubrication system rather than be greasepacked. Grease tends to separate at high speeds; however, new greases being developed may change the whole coupling picture. In many cases, disc-type couplings are replacing gear couplings. This type of coupling is more forgiving of angular alignment problems and also does not require any type of lubrication. The piping to the system can often aggravate misalignment problems in the system. Pipe stresses can run very high, and, under these forces, the compressor could move, thus causing high misalignments in the system. Seal problems can give rise to high leakages and thrust problems. The high leakages reduce the efficiency of the unit and can also lead to contamination of the lubricant. Thrust problems are created by air leakage past seals, causing an unbalance of the thrust forces on the system. The previous problems are some of the more common types encountered on a gas turbine-compressor train. Regular and preventive maintenance is the key to a successful operation. Problems will arise, but by proper monitoring of the aerothermal and mechanical problems, preventive maintenance can often avert major or catastrophic failures.

SYNCHRONOUS AND HARMONIC SPECTRA ALIGNMENT PROBLEMS Good shaft alignment must be maintained under dynamic conditions to achieve trouble-free operation. To accomplish this, the thermal growth of the individual elements must be taken into account when aligning

| 639 |

| Centrifugal Compressors | compressor equipment in cold conditions. For instance, the diameter and length of a compressor case will increase due to the heat of compression. Normally, the casing is supported at the horizontal centerline to allow the case to grow in diameter without changing the position of the shaft. The mounting feet on one end of the case are bolted and doweled to the foundation, as shown in Figure 14-29. The opposite end of the case must be allowed to move axially as the casing expands. This has been accomplished in the traditional compressor mount with hold down shoulder bolts that allow the case to slide axially upon lubricated shims but limit the vertical movement of the case. With this arrangement, a vertical key and keyway are located between the case and foundation on the vertical centerline to prevent transverse movement of the case. This mounting requires a special foundation to support the vertical key and also lubrication of the shims. GEAR NOISE Gear noise is a very important factor in detecting problems with gears and gearboxes. Noise from an operating gear set is a function of roundness and concentricity of operating elements (both gearing and shafting), accurate balance, and in particular, control of tooth spacing errors and uniformity of mesh stiffness to reduce meshing frequency excitation. It is significant that, for submarine gears where the ultimate in quietness is essential, the hallmarks are moderate tooth loading, fine pitch, high-helix angle, and low-pressure angle—all diametrically opposite from the usual and necessary practice for single-helical, hardened, and ground gearing, which have low-helix angle (for minimum thrust), very coarse pitch teeth (to get adequate strength), and high loading (because of the carburized hardening). Some factors causing gear noise can be attributed, but not limited, to the following:

| 640 |

•

Tooth spacing or involute error

•

Contact ratio

•

Surface finish

•

Wear on tooth flanks or pitting

•

Excessive or too little backlash

•

Gear, shaft, or housing resonance

| Maintenance Techniques | •

Tooth deflections

•

Pitch-line runout

•

Load intensity on gearing

•

Clutches and couplings

•

Lube oil pump and piping

•

Transmitted noise from driven or driving machinery

REFERENCES Boyce M. P. Gas Turbine Engineering Handbook. Second Edition Butterworth and Hienemann. December 2001. Boyce, M. P. “Managing Power Plant Life Cycle Costs.” International Power Generatio., July 1999, pp. 21-23. Nakajima, Seiichi. s Productivity Press, Inc. Sohre, J. “Operating Problems with High-Speed TurbomachineryCauses and Correction.” 23rd Annual Petroleum Mechanical Engineering Conference. Sept. 1968. Sohre, J. “.Reliability Evaluation for Trouble-Shooting of High-Speed Turbomachinery.” ASME Petroleum Mechanical Engineering Conference. Denver, CO. Herbage, B. S. “High Efficiency Film Thrust Bearings for Turbomachinery.” Proceedings of the 6th Turbomachinery Symposium. Texas A&M Univ. 1977. pp. 33-38. VanDrunen, G. and Liburdi, J. “Rejuvenation of Used Turbine Blades by Host Isostatic Processing.” Proceedings of the 6th Turbomachinery Symposium. Texas A&M Univ, 1977. pp. 55-60. Nelson, E. “Maintenance Techniques for Turbomachinery.” Proceedings of the 2nd Turbomachinery Symposium. Texas A&M Univ. 1973.

| 641 |

List of Acronyms AC

Alternating Current

AGMA

American Gear Manufacturers Association

AISI

American Iron and Steel Institute

API

American Petroleum Institute

ASME

American Society of Mechanical Engineers

BTU

British Thermal Unit

CFD

Computer Fluid Dynamics

DC

Direct Current

DFLP

Double-Flow, Low-Pressure

DLE

Dry Low Emissions

DLN

Dry Low NOx

DS

Directional Solidification

EPA

Environmental Protection Agency

ESD

Emergency Shutdown Valve

FFT

Fast Fourier Transform

FOD

Foreign Object Damage

GUI

Graphic User Interface

HP

High Pressure

IP

Intermediate Pressure

LCF

Low-Cycle Fatigue

LP

Low Pressure

LSD

Low-Solidity Diffusers

MI

Maintainability Improvement

MW

Megawatt

MP

Maintenance Prevention

NEMA

National Electrical Manufacturers Association

| xxxiii |

| Centrifugal Compressors | OSHA

Occupational Safety and Health Administration

PID

Propitional Integral Derivative

PTC

Performance Test Code

PSI

Pounds Per Square Inch

PTPM

Performance-Based Total Productive Maintenance

RPM

Revolutions Per Minute

RTD

Resistance Temperature Detectors

RTS

Rise to Surge

TDH

Total Differential Head

TEFC

Totally Enclosed Fan Cooled

TEWAC

Totally Enclosed Water Air Cooled

TPM

Total Productive Maintenance

TSIP

Turbine Steam Inlet Pressure

TSIT

Turbine Steam Inlet Temperature

UHC

Unburned Hydrocarbon

| xxxiv |

List of Constants SPEED fpm 88 --------mph

mph 0.6818 --------fps

m/s 0.5144 --------knot

m/s 0.3048 --------fps

m/s 0.44704 --------fps

fps 1.467 --------mph

mph 1.152 --------knot

fps 1.689 --------knot

cm/min 152.4 --------------ips

FORCE, MASS dynes 444,820 -----------lbf

oz 16 --------lbm

lbm 32.174 --------slug

lbf 1000 --------kip

poundals dynes 32.174 ------------------ 980,665 -----------lbf gmf

lbm 2000 --------ton kg 14.594 ---------slug

grains 7000 -----------lbm gm 28.35 --------oz

lbm 2.205 --------kg

N 9080665 --------kgf

kg 14.594 --------slug

N 4.4482 --------lbf

gm 453.6 ------lbm

dynes 105 -----------N

kilopond 1 ------------------kg

gmole 453.6 -----------pmole

kg kg 907.18 -------- 1000 --------------------ton metric ton

PRESSURE psi 14.696 -------atm N/m2 47.88 ----------psf

N/m2 101,325 ----------atm

kg 13.6 ----------------------mm Hg(0°C)

mm Hg(0°C) 51.715 ----------------------psi

in Hg(0°C) in H2O(60°F) 29.921 ------------------- 13.57 ------------------------atm in Hg(60°F)

ft H2O(60°F) N/m2 6894.8 ----------- 33.934 ----------------------psi atm kg/cm2 torr 0.0731 ------------- 760 -------psi atm 9.869 atm ----------------------------2 7 dyne/cm 10

psi 14.504 ------bar

kg/m2 703.07 ----------------psi

psi 0.0361 -----------------------in H2O(60°F)

bar dynes/cm2 psi 1.01325 ------- 106 ---------------------- 0.4898 --------------------atm bar in Hg(60°F)

N/m2 ft H2O(60°F) mm Hg(0°C) 133.3 ---------- 33.934 -----------------------760 -----------------------torr atm atm

in H2O(39.2°F) 406.79 ---------------------------atm

dyne/cm2 0.1 ----------------N/m2

kg/cm2 1.0332 -------------atm

| xxxv |

| Centrifugal Compressors | ENERGY AND POWER ft–lb Btu hp–hr 778.16 --------- 2544.4 ------------- 505 --------------Btu hp–hr ft–lb ft–lb 550 ---------hp–s

Btu 42.4 -------------hp–min

ft–lb 33,000 --------------hp–min

Btu/lb 1.8 --------------cal/gm

Btu 3412.2 -----------kW/hr

1.6021 erg ft–lb ------------------ 737.562 ---------------12 ev kW–min 10 11.817 ft–lb J -------------------1.3558 --------12 M ev ft–lb 10

J J 1 --------- --------W–s N–m kW–s 1 -----------kJ

16.021 J ------------ --------MeV 1012

Btu/pmole 1800 ---------------------kcal/gmole

Btu 56.87 ---------------kW–min

cal 251.98 -------Btu

bar–dm3 0.01 ---------------J

V–amp 1 ------------W–s

Btu 2.7194 -------------atm–ft3

kJ 4.1868 ------------kW–hr

ergs 107 --------J

kW 0.746 --------hp

ENTROPY, SPECIFIC HEAT, GAS CONSTANT Btu/pmole–°R 1 --------------------------cal/gmole–K Btu/pmole–°R 0.2389 --------------------------J/gmole–K

Btu/lb–°R 1 -----------------------cal/gmole–K

Btu/lb–°R 1 -------------------kcal/kg–K

kJ/kg–K 4.187 ------------------Btu/lb–°R

UNIVERSAL GAS CONSTANT atm–ft3 .07302 ----------------pmole°R

ft–lb 1545.32 ------------------pmole–°R

kJ 8.314 ---------------------kmole–K

atm–cm3 82.057 ------------------gmole–K

cal 1.9859 ----------------gmole–K

psi–ft3 10.731 ------------------pmole–°R

bar–cm3 83.143 ------------------gmole–K

J 83.143 ----------------gmole–K

erg 8.3149 x 107 -----------------gmole–K

atm–m3 0.08206 ----------------gmole–K

bar–1 0.83143 ----------------gmole–K

| xxxvi |

| List of Constants | NEWTON’S PROPORTIONALITY CONSTANT k (as a conversion unit) lb 32.174 fps2 --------slug

(

lb 386.1 ips2 ---------p sin

)

m dynes ------------9.80665 ------s2 gm

(

(

)

m N 9.80665 -----------s2 kg

( )

)

MISCELLANEOUS CONSTANTS Speed of Light m c = 2.9979 x 108 ---s Planck Constant h = 6.6256 x 10-34J-s Gravitational Constant N–m2 G = 6.6.670 x 10-11 ----------kg2

Avogadro Constant molecules NA = 6.02252 x 1023 -------------------gmole Boltzmann Constant J k = 1.38054 x 10-23 ----K Normal Mole Volume m3 2.24136 x 10-2 -----------gmole

| xxxvii |

List of Figures FIGURE 1-1:

Principal Types of Compressors . . . . . . . . . . . . . . . 2

FIGURE 1-2:

Performance Characteristics of Different Types of Compressors . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3

FIGURE 1-3:

Variation of Adiabatic Efficiency with Specific Speed for Three Types of Compressors . . . . . . . . . 4

FIGURE 1-4:

Operating Characteristics of a Compressor. . . . . . 5

FIGURE 1-5:

A High-Pressure Ratio Turbine Rotor. . . . . . . . . . . 7

FIGURE 1-6:

Coordinate System for Axial Flow Compressor . . 7

FIGURE 1-7:

Variation of Flow and Thermodynamic Properties in an Axial Flow Compressor . . . . . . . . . . . . . . . . . 8

FIGURE 1-8:

Aerodynamic and Thermodynamic Properties in a Centrifugal Compressor Stage . . . . . . . . . . . . . . 10

FIGURE 1-9:

Flow in a Vaned Diffuser . . . . . . . . . . . . . . . . . . . . 10

FIGURE 1-10:

Theoretical Head Characteristics as a Function of the Flow in a Centrifugal Impeller . . . . . . . . . . 12

FIGURE 1-11:

A Small Radial Flow Gas Turbine Cutaway Showing the Turbine Rotor . . . . . . . . . . . . . . . . . . 15

FIGURE 1-12:

A Small Aeroderivative Gas Turbine. ST30 Marine and Industrial Gas Turbine Engine . . . . . 16

FIGURE 1-13:

A Compact Microturbine Schematic. . . . . . . . . . . 17

FIGURE 1-14:

An Integral Geared Centrifugal Compressor, Showing Intercoolers Below the Compressor Base Plate. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18

FIGURE 1-15:

Horizontally Split Centrifugal Compressor, With Closed-Face Impellers . . . . . . . . . . . . . . . . . 19

FIGURE 1-16:

Horizontal Casing Centrifugal Compressor . . . . 20

FIGURE 1-17:

Horizontal Casing Centrifugal Compressor . . . . 21

FIGURE 1-18:

The Inner Bundle of a Barrel Centrifugal Compressor Opened . . . . . . . . . . . . . . . . . . . . . . . 21

| ix |

| Centrifugal Compressors |

|x|

FIGURE 1-19:

The Long Train is a Typical Arrangement for Acid Nitric Plants (process UHDE) . . . . . . . . . . . 22

FIGURE 1-20:

An Integral Geared Centrifugal Compressor Rotor Assembly Showing Intercoolers Below the Compressor Base Plate . . . . . . . . . . . . . . . . . . . . . 23

FIGURE 1-21:

Two Sets of Closed-Face Impellers Used in a Centrifugal Compressor. . . . . . . . . . . . . . . . . . . . . 23

FIGURE 1-22:

Axial Flow Compressors and Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24

FIGURE 1-23:

Various Configurations of Centrifugal Compressors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25

FIGURE 1-24:

An Integral Geared Single Stage Centrifugal Compressor, With a Cast Scroll Exit, for Mechanical Vapor Recompression Service . . . . . 27

FIGURE 1-25:

Schematic of an Integral Geared Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27

FIGURE 1-26:

Schematic of a Typical Rotor / Bull Gear Configuration for a Six Stage Compressor . . . . . 28

FIGURE 1-27:

An Integral Geared Centrifugal Compressor, Showing Intercoolers Below the Compressor Base Plate . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29

FIGURE 1-28:

Straight Flow-Through Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30

FIGURE 1-29:

Back-to-Back Compressor with Double-Flow Inlet . . . . . . . . . . . . . . . . . . . . . . . . . 30

FIGURE 1-30:

Back-to-Back Centrifugal Compressor . . . . . . . . 31

FIGURE 1-31:

Side-Loaded Compressor . . . . . . . . . . . . . . . . . . . 31

FIGURE 1-32:

Open-Faced Impeller Fabrication . . . . . . . . . . . . 33

FIGURE 1-33:

Closed-Faced Impeller . . . . . . . . . . . . . . . . . . . . . 34

FIGURE 1-34:

Several Fabrication Techniques for Centrifugal Impellers. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34

FIGURE 2-1:

An Ideal Convergent Divergent Nozzle. . . . . . . . . 45

| List of Figures | FIGURE 2-2:

Compressibility Factor Chart for a Simple Fluid . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 47

FIGURE 2-3:

Flow Characteristics in a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 53

FIGURE 2-4:

Theoretical Head as a Function of Flow for Various Types of Blade Angle at the Impeller Exit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 56

FIGURE 2-5:

Enthalpy-Entropy Diagram for a Compressor . . . 58

FIGURE 2-6:

Relationship Between Polytropic and Adiabatic Efficiency for a Compressor . . . . . . . . . . . . . . . . . 60

FIGURE 2-7:

Variation of Adiabatic Efficiency with Specific Speed for Three Types of Compressors . . . . . . . . 61

FIGURE 2-8:

Compressor Selection Map . . . . . . . . . . . . . . . . . . 61

FIGURE 2-9:

The Effect of Flow Coefficient and Mach Number on the Polytropic Efficiency of Centrifugal Compressors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 62

FIGURE 2-10:

Effect on Polytropic Head as Function of Flow Coefficient, Blade Angle, and Degree of Reaction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 62

FIGURE 2-11:

A Typical Centrifugal Compressor Performance Map. . . . . . . . . . . . . . . . . . . . . . . . . . 63

FIGURE 2-12:

A modified Compressor Performance Map Where Speed Lines are a Function of Power and Flow Rate . . . . . . . . . . . . . . . . . . . . . . 64

FIGURE 2-13:

Variation of Power and Pressure Rise as a Function of Flow . . . . . . . . . . . . . . . . . . . . . . 65

FIGURE 2-14:

The Effect of Pressure Ratio on the Operating Range of Centrifugal Compressor. . . . . . . . . . . . . 66

FIGURE 2-15:

Effect of Load Characteristics on Compressor Performance. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 66

FIGURE 2-16:

Effect of Temperature on the Performance of a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . 67

FIGURE 2-17:

Typical Compressor Performance Map for a Multistage Centrifugal Compressor . . . . . . . . . . . 68

| xi |

| Centrifugal Compressors |

| xii |

FIGURE 3-1:

Air Blower Fluid Catalytic Cracking Unit . . . . . . 74

FIGURE 3-2:

Recommended System for Future Power Recovery Installation in Fluid Catalytic Cracking Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 74

FIGURE 3-3:

Single Case for Gas Recovery Compressor in PCC Unit. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 76

FIGURE 3-4:

A Two-Barrel Compression Train with a Knockout Drum Between the Two Units . . . . . . 76

FIGURE 3-5:

Nitric Acid Ammonia Train . . . . . . . . . . . . . . . . . . 77

FIGURE 3-6:

Turboexpander-Compression System. . . . . . . . . . 78

FIGURE 3-7:

Refrigeration System . . . . . . . . . . . . . . . . . . . . . . . 79

FIGURE 3-8:

Compressor Map Showing Limitation of Casings and Impellers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83

FIGURE 3-9:

Performance Characteristics of Different Types of Compressor Drives . . . . . . . . . . . . . . . . . . . . . . 88

FIGURE 3-10:

Rotor Response Plot . . . . . . . . . . . . . . . . . . . . . . 106

FIGURE 4-1:

Centrifugal Compressor Map . . . . . . . . . . . . . . . 124

FIGURE 4-2:

Single Stage Centrifugal Compressor . . . . . . . . . 125

FIGURE 4-3:

Multistage Centrifugal Compressor . . . . . . . . . . 125

FIGURE 4-4:

Flow in a Vaned Diffuser . . . . . . . . . . . . . . . . . . 126

FIGURE 4-5:

Open Faced Impeller . . . . . . . . . . . . . . . . . . . . . 126

FIGURE 4-6:

Closed Faced Impellers . . . . . . . . . . . . . . . . . . . 127

FIGURE 4-7:

Pressure and Velocity through a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 127

FIGURE 4-8:

The Two Major Planes in a Centrifugal Impeller. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 129

FIGURE 4-9:

Flow Map in a Blade-to-Blade Plane . . . . . . . . . 129

FIGURE 4-10:

Flow in the Meridional Plane, Also Known as the Hub-to-Shroud Plane. . . . . . . . . . . . . . . . . 130

FIGURE 4-11:

Velocity Distribution in the Meridional Plane of a Centrifugal Impeller . . . . . . . . . . . . . . . . . . . 130

| List of Figures | FIGURE 4-12:

Velocity Distribution in the Blade-to-Blade Plane of a Centrifugal Impeller . . . . . . . . . . . . . . . . . . . 131

FIGURE 4-13:

Types of Entry-Inducer Systems . . . . . . . . . . . . . 132

FIGURE 4-14:

Adjustable Inlet Guide Vane . . . . . . . . . . . . . . . . 133

FIGURE 4-15:

Inlet Velocity Diagrams . . . . . . . . . . . . . . . . . . . . 134

FIGURE 4-16:

Estimated Effect of Inlet Prewhirl . . . . . . . . . . . 135

FIGURE 4-17:

Prewhirl Distribution Patterns . . . . . . . . . . . . . . 136

FIGURE 4-18:

Euler Work Distribution at an Impeller Exit, Showing the Effects of Pre-rotation . . . . . . . . . . 136

FIGURE 4-19:

Inducer in a Centrifugal Compressor . . . . . . . . 137

FIGURE 4-20:

Open-Faced Impeller with Splitter Blades . . . . 138

FIGURE 4-21:

Splitter Blade and Tandem Inducer . . . . . . . . . . 139

FIGURE 4-22:

Velocity Triangles for Various Types of Impeller Blading . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 139

FIGURE 4-23:

Head Flow Characteristics for Various Outlet Blade Angles. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 140

FIGURE 4-24:

Velocity Triangles at the Exit of the Impeller . . . 141

FIGURE 4-25:

Forces and Flow Characteristics in a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 143

FIGURE 4-26:

Coriolis Circulation . . . . . . . . . . . . . . . . . . . . . . . 144

FIGURE 4-27:

Boundary Layer Development. . . . . . . . . . . . . . . 144

FIGURE 4-28:

Laminar Flow Control in a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 145

FIGURE 4-29:

Effect on Exit Velocity Triangles by Various Parameters . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 147

FIGURE 4-30:

Slip Flow as a Function of the Flow Coefficient . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 147

FIGURE 4-31:

Flow Characteristics in a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 151

FIGURE 4-32:

Streamlines and Potential Flow Lines Describing the Flow in the Meridional Plane. . . 152

FIGURE 4-33:

Optimum Mach Number and Prewhirl Angle . . . 155

| xiii |

| Centrifugal Compressors |

| xiv |

FIGURE 4-34:

Flow Through a Centrifugal Impeller . . . . . . . . . 156

FIGURE 4-35:

Design Flow Chart For Centrifugal Compressor Impeller and Diffuser. . . . . . . . . . . . . . . . . . . . . . 160

FIGURE 4-36:

Meridional Plane Area Distribution Showing Streamline and Potential Flow Lines . . . . . . . . . 164

FIGURE 4-37:

Velocities In a Centrifugal Impeller— Inviscid Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . 165

FIGURE 4-38:

Energy Loss in an Impeller . . . . . . . . . . . . . . . . . 165

FIGURE 4-39:

Computational Mesh for 3D Viscous Flow Calculation Impeller . . . . . . . . . . . . . . . . . . . . . . 166

FIGURE 4-40:

Velocities In a Centrifugal Impeller— Viscous Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . 178

FIGURE 4-41:

Comparisons in Velocities in the Meridional Plane in a Centrifugal Impeller Using an Inviscid Flow Solution as Compared to a Viscous Flow Solution . . . . . . . . . . . . . . . . . . . . . 179

FIGURE 4-42:

Comparisons in Velocities In the Blade-to-Blade Plane in a Centrifugal Impeller using an Inviscid Flow Solution as Compared to a Viscous Flow Solution. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 179

FIGURE 5-1:

Vaneless Diffuser Shapes. . . . . . . . . . . . . . . . . . . 184

FIGURE 5-2:

Adjustable Air-foil Style Diffuser with Variable Guide Vanes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 185

FIGURE 5-3

Island Diffuser . . . . . . . . . . . . . . . . . . . . . . . . . . . 185

FIGURE 5-4:

Low Solidity Vaned Diffuser . . . . . . . . . . . . . . . . 186

FIGURE 5-5:

Schematic of a Rib Diffuser . . . . . . . . . . . . . . . . 186

FIGURE 5-6:

Cascade Diffuser—Showing Three Stages of Diffusion. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 187

FIGURE 5-7:

Flow Wake Leaving the Impeller. . . . . . . . . . . . . 188

FIGURE 5-8:

Moody Diagram Showing the Relationship of the Frictional Coefficient of Diffusers as a Function of the Reynolds Number . . . . . . . . . . . 191

FIGURE 5-9:

Flow Regions of the Vaned Diffuser . . . . . . . . . . 192

| List of Figures | FIGURE 5-10:

Schematic of a Supersonic Diffuser . . . . . . . . . . 193

FIGURE 5-11:

Separation Limit for Two-Dimensional Straight Walled Diffusers . . . . . . . . . . . . . . . . . . . . . . . . . . 194

FIGURE 5-12:

Airfoil Diffuser Geometry . . . . . . . . . . . . . . . . . . 196

FIGURE 5-13:

Straight Vaned Flat Plate Diffuser showing the Diffusion Angle δ and the Air Angle Entering the Diffuser α . . . . . . . . . . . . . . . . . . . . . . . . . . . . 197

FIGURE 5-14:

A Five Stage Horizontally Split Centrifugal Compressor, Showing the Return Channels. . . . 199

FIGURE 5-15:

Continuous Diffuser Return Channel . . . . . . . . . 200

FIGURE 5-16:

Two Stage Closed Wheel Centrifugal Compressor Showing the Return Passage and the Exit Volute . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 201

FIGURE 5-17:

Flow Patterns in a Volute . . . . . . . . . . . . . . . . . . 203

FIGURE 6-1:

Compressor Performance Map . . . . . . . . . . . . . . 208

FIGURE 6-2:

Energy Loss in an Impeller . . . . . . . . . . . . . . . . . 209

FIGURE 6-3:

Characteristic Dimensions in a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 210

FIGURE 6-4:

Leakage Affecting Clearance Loss . . . . . . . . . . . 212

FIGURE 6-5:

Secondary Flow at the back of the Impeller to Determine Disk Friction Loss . . . . . . . . . . . . . . . 213

FIGURE 6-6:

Shock Wave at the Impeller Eye. The Oblique Shock Shown is the Desirable Shock as Opposed to the Standing Bow Shock . . . . . . . . . 214

FIGURE 6-7:

Inlet Velocity Triangles at Non-Zero Incidence to Determine Incidence Loss . . . . . . . . . . . . . . . 215

FIGURE 6-8:

Schematic of Secondary Flow Circulation in a Centrifugal Impeller. . . . . . . . . . . . . . . . . . . . . . . 217

FIGURE 6-9:

Diffuser Recirculating Loss . . . . . . . . . . . . . . . . . 219

FIGURE 6-10:

Flow Trajectory in a Vaneless Diffuser. . . . . . . . 220

FIGURE 6-11:

Flow in a Vaned Diffuser . . . . . . . . . . . . . . . . . . . 221

FIGURE 6-12:

Losses in a Centrifugal Compressor . . . . . . . . . . 223

| xv |

| Centrifugal Compressors |

| xvi |

FIGURE 7-1:

Initiation of Surge in Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 226

FIGURE 7-2:

A Typical Compressor Performance Map. . . . . . 228

FIGURE 7-3:

Flow Reversal in a Centrifugal Compressor as the Compressor Approaches Surge . . . . . . . . . . 230

FIGURE 7-4:

Lift as a Function of the Angle of Incidence . . . 231

FIGURE 7-5:

Schematic of the Flow Around an Airfoil With Increasing Angle of Attack of the Airfoils . . . . . 231

FIGURE 7-6:

Rotating Stall . . . . . . . . . . . . . . . . . . . . . . . . . . . . 232

FIGURE 7-7:

Typical Centrifugal Compressor Performance Map . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 234

FIGURE 7-8:

Performance Map with System Curves and Surge . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 235

FIGURE 7-9:

Head Flow Rate Characteristics for Various Outlets Blade Angles . . . . . . . . . . . . . . . . . . . . . . 238

FIGURE 7-10:

Typical Compressor Head Losses for Backward Curve Blades . . . . . . . . . . . . . . . . . . . . . . . . . . . . 239

FIGURE 7-11:

Effect of Number of Impellers on Stability and Capacity Limit . . . . . . . . . . . . . . . . . . . . . . . . . . . 241

FIGURE 7-12:

Typical Overall Stability vs. Individual Stability for Two Casings Operating in Series. . . . . . . . . . 242

FIGURE 7-13:

Effect of Blade Angle on Stability. . . . . . . . . . . . 243

FIGURE 7-14:

Design Condition Velocity Triangles. . . . . . . . . . 244

FIGURE 7-15:

Flow Trajectory in Vaneless Diffuser . . . . . . . . . 245

FIGURE 7-16:

Flow in a Vaned Diffuser . . . . . . . . . . . . . . . . . . . 246

FIGURE 7-17:

Effect of Guide Vane Setting (Stationary or Variable) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 246

FIGURE 7-18:

Constant Speed Machine with Variable Inlet Guide Vanes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 248

FIGURE 7-19:

Effect Of Gas Composition on the Characteristics of a Centrifugal Impeller . . . . . . 249

| List of Figures | FIGURE 7-20:

Process Operating Points, Flow 60% (A) & Flow 100% (D), Surge Point B, & Control Points C. . . . . . . . . . . . . . . . . . . . . . . . . . 252

FIGURE 7-21:

A Centrifugal Compressor-Inlet Throttling. . . . . 254

FIGURE 7-21B:

Centrifugal Compressor Variable Speed. . . . . . . 254

FIGURE 7-21C:

Axial Flow Compressor Variable Angle Stator Vanes. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 255

FIGURE 7-21D:

Axial Flow Compressor Variable Speed . . . . . . . 255

FIGURE 7-22:

Anti-Surge Control Scheme, Based on Inlet Volume Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . 257

FIGURE 7-23:

Centrifugal Compressor Antisurge Control J vs. Pd . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 258

FIGURE 7-24:

Wet Weight Flow vs. Brake Horsepower and Discharge Pressure . . . . . . . . . . . . . . . . . . . . . . . 258

FIGURE 7-25:

Anti-Surge Control Based on Typical Flow / Differential Pressure System. . . . . . . . . . . . . . . . 259

FIGURE 7-26:

Alternate Flow Differential Pressure Type Anti-Surge Control System . . . . . . . . . . . . . . . . . 261

FIGURE 7-27:

Discharge Volume Flow (P&T Compensated), Antisurge Control Variable Speed Centrifugal. . . 261

FIGURE 7-28:

Surge & Control Lines Discharge Volume Flow (P&T Compensated) Antisurge Control . . . . . . . 261

FIGURE 7-29:

Discharge Flow Antisurge Control Variable Speed Centrifugal Compressor . . . . . . . . . . . . . . 262

FIGURE 7-30:

Surge Control Lines, Discharge Flow Control, Variable Speed Centrifugal . . . . . . . . . . . . . . . . . 262

FIGURE 7-31:

Surge and Control Lines, Flow Delta-P, AntiSurge Control, and Normal Gas . . . . . . . . . . . . . 265

FIGURE 7-32:

Unacceptable and Acceptable Equipment Arrangement. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 267

FIGURE 7-33:

Location of the Check Valve . . . . . . . . . . . . . . . . 268

FIGURE 7-34:

Series Operation of Compressors . . . . . . . . . . . . 269

FIGURE 7-35:

Parallel Operation of Compressors. . . . . . . . . . . 270

| xvii |

| Centrifugal Compressors |

| xviii |

FIGURE 7-36:

Simulation Study Diagram. . . . . . . . . . . . . . . . . . 273

FIGURE 7-37:

Compression System Model . . . . . . . . . . . . . . . . 275

FIGURE 7-38:

Compressor Flow Diagram . . . . . . . . . . . . . . . . . 275

FIGURE 7-39:

Compressor Characteristics as the Gas Turbine Trips at Compressor Full Load . . . . . . . . . . . . . . 277

FIGURE 7-:40

Compressor Characteristics as the Gas Turbine Trips at Compressor Full Load . . . . . . . . . . . . . . 277

FIGURE 7-41:

Block Valve Failure at Compressor Steady State Flow at Full Capacity. . . . . . . . . . . . . . . . . . . . . . 278

FIGURE 7-42:

Block Valve Failure at Steady State Compressor Flow at Full Capacity. . . . . . . . . . . . . . . . . . . . . . 279

FIGURE 7-43:

Block Valve Failure at Steady State Flow at Full Capacity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 279

FIGURE 8-1:

Development of Engine Pressure Ratio Over the Years . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 283

FIGURE 8-2:

Trend in Improvement in Firing Temperature . . . 283

FIGURE 8-3:

Overall Cycle Efficiency . . . . . . . . . . . . . . . . . . . 285

FIGURE 8-4:

Performance Map of a Simple Cycle Gas Turbine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 286

FIGURE 8-5:

Performance Map of a Regenerative Gas Turbine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 286

FIGURE 8-6:

A Frame Type Gas Turbine with Can-Annular Combustors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 289

FIGURE 8-7:

Frame Type Gas Turbine with Silo Type Combustors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 290

FIGURE 8-8:

A Cross Section of an Aeroderivative Gas Turbine Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . 291

FIGURE 8-9:

A Medium Sized Industrial Gas Turbine . . . . . . 293

FIGURE 8-10:

A Recuperative Medium Sized Industrial Gas Turbine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 294

FIGURE 8-11:

A Small Radial Flow Gas Turbine Cutaway Showing the Turbine Rotor . . . . . . . . . . . . . . . . . 295

| List of Figures | FIGURE 8-12:

A Small Aeroderivative Gas Turbine . . . . . . . . . 296

FIGURE 8-13:

A High-Pressure Ratio Turbine Rotor . . . . . . . . 298

FIGURE 8-14:

Variation of Flow and Thermodynamic Properties in an Axial Flow Compressor . . . . . . 298

FIGURE 8-15:

A Typical Plate and Fin Type Regenerator for an Industrial Gas Turbine . . . . . . . . . . . . . . . . . . 299

FIGURE 8-16:

A Typical Combustor Can With Straight Through Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . 301

FIGURE 8-17:

Air Distribution in a Typical Combustor . . . . . . 302

FIGURE 8-18:

A Typical Reverse Flow-Can Annular Combustor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 305

FIGURE 8-19:

A Typical Single Can Side Combustor . . . . . . . . 306

FIGURE 8-20:

A Typical Combustor Showing the NOx Production Zone. . . . . . . . . . . . . . . . . . . . . . . . . . 309

FIGURE 8-21:

The Effect of Flame Temperature on Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 311

FIGURE 8-22:

Correlation of Adiabatic Flame Temperature with NOx Emissions . . . . . . . . . . . . . . . . . . . . . . 311

FIGURE 8-23:

Effect of Fuel / Air Ratio on Flame Temperature and NOx Emissions . . . . . . . . . . . . . . . . . . . . . . . 312

FIGURE 8-24:

A Schematic Comparison of a Typical Dry Low Emission NOx Combustor and a Conventional Combustor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 313

FIGURE 8-25:

Schematic of a Dry Low Emission NOx Combustor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 315

FIGURE 8-26:

Shows the Staging of Dry Low Emissions Combustor as the Turbine is Brought to Full Power . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 315

FIGURE 8-27:

Schematic Temperature Profiles for Catalyst (Substrate) and Bulk Gas in a Traditional Catalytic Combustor . . . . . . . . . . . . . . . . . . . . . . 319

| xix |

| Centrifugal Compressors |

| xx |

FIGURE 8-28:

Schematic Temperature Profiles for Catalytica Catalytic Combustion System in which the Wall Temperature is Limited and Complete Combustion Occurs after the Catalyst . . . . . . . . 320

FIGURE 8-29:

Schematic on a Full Scale Catalytic Combustor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 321

FIGURE 8-30:

Variation of Utilization factor with U/V . . . . . . . 324

FIGURE 8-31:

Turbine Performance Map. . . . . . . . . . . . . . . . . . 325

FIGURE 8-32:

Performance Map Showing the Variation of Heat Rate and Power as a Function of Turbine Speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 325

FIGURE 8-33:

Schematic of an Axial Flow Turbine . . . . . . . . . 326

FIGURE 8-34:

Schematic of an Impulse Turbine showing the Variation of the Thermodynamic and Fluid Mechanic Properties . . . . . . . . . . . . . . . . . . . . . . 328

FIGURE 8-35:

Schematic of a Reaction Type Turbine Showing the Distribution of the Thermodynamic and Fluid Mechanic Properties . . . . . . . . . . . . . . . . . 329

FIGURE 8-36:

The Effect of Exit Velocity and Air Angle on the Utilization Factor for a 50% Reaction Axial Turbine Stage . . . . . . . . . . . . . . . . . . . . . . . . . . . . 331

FIGURE 8-37:

Types of Cooling Schemes Used in Gas Turbine for Cooling the Hot Section. . . . . . . . . . . . . . . . . 332

FIGURE 8-38

Cooling Based on Convection and Impingement Cooling. Temperature Distribution Based on Uncooled ºF and (Cooled ºF) Blades . . . . . . . . . 335

FIGURE 8-39:

Cooling Based on Convection and Film Cooling. Temperature Distribution Based on Uncooled ºF and (Cooled ºF) Blades . . . . . . . . . . . . . . . . . . 336

FIGURE 8-40:

Internal of the Frame FA Blades, Showing Cooling Passages . . . . . . . . . . . . . . . . . . . . . . . . 338

FIGURE 8-41:

The Effect of Various Types of Cooling on Turbine Blade Efficiency . . . . . . . . . . . . . . . . . . . 339

| List of Figures | FIGURE 8-42:

Hot Gas Expander Driving a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 340

FIGURE 8-43:

Cantilever-Type Radial Inflow Turbine . . . . . . . . 341

FIGURE 8-44:

Mixed Flow Type Radial Inflow Turbine . . . . . . 341

FIGURE 8-45:

Components of a Radial Inflow Turbine . . . . . . 343

FIGURE 8-46:

Firing Temperature Increase with Blade Material Improvement . . . . . . . . . . . . . . . . . . . . . 343

FIGURE 8-47:

Larson Miller Parameter for Various Types of Blades. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 346

FIGURE 8-48:

A Schematic of a Gas Turbine Driving a Centrifugal Compressor. . . . . . . . . . . . . . . . . . . . 350

FIGURE 8-49:

A Typical Centrifugal Compressor Performance Map . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 351

FIGURE 8-50:

A Modified Compressor Performance Map Where Speed Lines are a Function of Power and Flow Rate . . . . . . . . . . . . . . . . . . . . . . . . . . . 353

FIGURE 8-51:

Gas Turbine Performance Map . . . . . . . . . . . . . . 353

FIGURE 8-52:

Efficiency as a Function of Pressure, RPM, and Enthalpy in a Turbine . . . . . . . . . . . . . . . . . . . . . 354

FIGURE 8-53:

Compressor Map Showing Limitation of Casings and Impellers . . . . . . . . . . . . . . . . . . . . . 354

FIGURE 8-54:

A Performance Map Showing Horsepower Developed by Turbine and Required by the Compressor for Constant Pressures. . . . . . . . . . 355

FIGURE 8-55:

Effect of Compressor Inlet Temperature on the Performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . 355

FIGURE 8-56:

Two Shaft Gas Turbine Performance Curve. . . . 356

FIGURE 9-1:

Steam Turbine Driving a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 360

FIGURE 9-2:

Schematic of a Steam Turbine Power Plant. . . . 361

FIGURE 9-3:

Pressure-Volume Diagram of a Typical Steam Turbine Power Plant . . . . . . . . . . . . . . . . . . . . . . 361

| xxi |

| Centrifugal Compressors |

| xxii |

FIGURE 9-4:

Temperature-Entropy Diagram of a Typical Steam Turbine Power Plant. . . . . . . . . . . . . . . . . 362

FIGURE 9-5:

Schematic of a Regenerative-Reheat Steam Turbine Power Plant . . . . . . . . . . . . . . . . . . . . . . 364

FIGURE 9-6:

Temperature-Entropy Diagram of a Regenerative-Reheat Steam Turbine Power Plant . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 364

FIGURE 9-7:

Radial Flow or Ljongstrom Turbine . . . . . . . . . . 369

FIGURE 9-8:

Steam Turbine Losses as a Function of Mach Number . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 371

FIGURE 9-9:

Partial Entry Steam Turbine Performance Properties . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 372

FIGURE 9-10:

A Simple Impulse Steam Turbine . . . . . . . . . . . . 374

FIGURE 9-11:

A Curtis Type Steam Turbine . . . . . . . . . . . . . . . 375

FIGURE 9-12:

Pressure Compounded Steam Turbine-Ratteau Turbine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 376

FIGURE 9-13:

Reaction Turbine . . . . . . . . . . . . . . . . . . . . . . . . . 377

FIGURE 9-14:

A Schematic of Arrangements of Steam Turbines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 378

FIGURE 9-15:

A Typical Compound Turbine . . . . . . . . . . . . . . . 379

FIGURE 9-16:

Typical Types of Blade Roots . . . . . . . . . . . . . . . 383

FIGURE 9-17:

Construction of a Typical Steam Turbine Diaphragm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 384

FIGURE 9-18:

Nomenclature for a Typical LP Turbine Blade . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 385

FIGURE 10-1:

A Range of Speed and Horsepower for Various Types of Motors . . . . . . . . . . . . . . . . . . . . . . . . . . 396

FIGURE 10-2:

Squirrel Cage Induction Motor Drive through a High Speed Gear Driving a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 398

FIGURE 10-3:

Synchronous Motor Driving a Three-Section Centrifugal Compressor. . . . . . . . . . . . . . . . . . . . 401

| List of Figures | FIGURE 10-4:

A Typical Motor vs. Driven Compressor Speed–Torque Characteristic Curve . . . . . . . . . . 405

FIGURE 10-5:

Typical Compressor Speed–Torque Curve . . . . . 407

FIGURE 10-6:

Nomogram for Locked Motor Currents for Induction Motors . . . . . . . . . . . . . . . . . . . . . . . . . 411

FIGURE 11-1:

Rotor Response Plot . . . . . . . . . . . . . . . . . . . . . . 416

FIGURE 11-2:

Rigid Supports . . . . . . . . . . . . . . . . . . . . . . . . . . . 418

FIGURE 11-3:

Flexible Supports. . . . . . . . . . . . . . . . . . . . . . . . . 418

FIGURE 11-4:

Critical Speed Map. . . . . . . . . . . . . . . . . . . . . . . . 420

FIGURE 11-5:

Rigid Supports and a Flexible Rotor . . . . . . . . . 421

FIGURE 11-6:

Flexible Supports and Rigid Rotors . . . . . . . . . . 421

FIGURE 11-7:

Characteristic of Forced Vibration or Resonance in Rotating Machinery . . . . . . . . . . . 425

FIGURE 11-8:

Characteristics of Instabilities or Self-Excited Vibration in Rotating Machinery . . . . . . . . . . . . . 427

FIGURE 11-9:

Hysteretic Whirl . . . . . . . . . . . . . . . . . . . . . . . . . . 428

FIGURE 11-10:

Dry Friction Whirl . . . . . . . . . . . . . . . . . . . . . . . . 429

FIGURE 11-11:

Oil Friction Whirl . . . . . . . . . . . . . . . . . . . . . . . . . 430

FIGURE 11-12:

Aerodynamic Cross Coupling . . . . . . . . . . . . . . . 431

FIGURE 11-13:

Whirl from Fluid Trapped in the Rotor. . . . . . . . 432

FIGURE 11-14:

Campbell Diagram . . . . . . . . . . . . . . . . . . . . . . . . 434

FIGURE 11-15:

Accelerometer Locations on Impeller Tested . . . 435

FIGURE 11-16:

Impeller Showing Nodal Points . . . . . . . . . . . . . 436

FIGURE 11-17:

Campbell Diagram of Tested Impeller . . . . . . . . 437

FIGURE 11-18:

The Showing of Formation of Oil Whirl in a Rotor System . . . . . . . . . . . . . . . . . . . . . . . . . . . . 438

FIGURE 11-19:

Severity Charts . . . . . . . . . . . . . . . . . . . . . . . . . . . 439

FIGURE 11-20:

Comparison of General Bearing Types . . . . . . . . 442

FIGURE 11-21:

Tilting-Pad Bearing Preload. . . . . . . . . . . . . . . . . 446

FIGURE 11-22:

Discharge Temperature Criteria . . . . . . . . . . . . . 447

| xxiii |

| Centrifugal Compressors |

| xxiv |

FIGURE 11-23:

Babbitt Fatigue Characteristics . . . . . . . . . . . . . 448

FIGURE 11-24:

Comparison of Thrust Bearing Types . . . . . . . . . 440

FIGURE 11-25:

Various Types of Thrust Bearings . . . . . . . . . . . . 449

FIGURE 11-26:

Thrust Bearing Temperature Characteristics . . . 451

FIGURE 11-27:

Difference in Total Power Loss Data–Test Minus Catalog Frictional Losses vs. Shaft Speed for 6 x 6 Pad Double-Element Thrust Bearings . . . . 452

FIGURE 11-28:

Various Configurations of Labyrinth Seals . . . . . 454

FIGURE 11-29:

Theory Behind the Knife-Edge Arrangement . . . 455

FIGURE 11-30:

Mechanical Contact Shaft Seal . . . . . . . . . . . . . . 457

FIGURE 11-31:

Single Dry Gas Seal . . . . . . . . . . . . . . . . . . . . . . . 463

FIGURE 11-32:

Spiral Grooved Mating Ring . . . . . . . . . . . . . . . . 463

FIGURE 11-33:

Flexible Coupling Operating Spectrum . . . . . . . 471

FIGURE 11-34:

Gear Coupling (Male Teeth Integral With the Hub) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 472

FIGURE 11-35:

Gear Coupling (Male Teeth Integral With the Spool) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 472

FIGURE 11-36:

Schematic of Gear Used in Coupling Applications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 473

FIGURE 11-37:

Recommended Limits of Misalignment vs. Operating Speed. . . . . . . . . . . . . . . . . . . . . . . . . . 474

FIGURE 11-38:

Schematic of a Typical Diaphragm Coupling . . . 477

FIGURE 11-39:

Axial Deflection in a Disc . . . . . . . . . . . . . . . . . . 478

FIGURE 11-40:

Stress Distribution Under Axial Deflection . . . . 479

FIGURE 11-41:

Typical Metal-Flexing Disc Coupling . . . . . . . . . 480

FIGURE 11-42:

Frictional Damping in a Metal-Disc Coupling . . . 480

FIGURE 11-43:

A Typical Lube Oil System. . . . . . . . . . . . . . . . . . 482

FIGURE 11-44:

Lube Oil Reservoir . . . . . . . . . . . . . . . . . . . . . . . . 483

FIGURE 11-45:

Cooler-Filter Arrangement . . . . . . . . . . . . . . . . . 485

FIGURE 12-1:

Forward and Feedback Control Loop . . . . . . . . 492

| List of Figures | FIGURE 12-2:

Steam Turbine Driven Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 495

FIGURE 12-3:

Schematic of Steam Turbine–Driven Centrifugal Compressor . . . . . . . . . . . . . . . . . . . 496

FIGURE 12-4:

Gas Turbine Driven Centrifugal Compressor . . . 496

FIGURE 12-5:

Start-Up Characteristics of a Gas Turbine . . . . . 497

FIGURE 12-6:

Compressor Characteristics Showing the Start-Up Characteristics . . . . . . . . . . . . . . . . . . . 498

FIGURE 12-7:

A Turbine Torque Characteristic Curve . . . . . . . 499

FIGURE 12-8:

Typical Simple Motor-Driven Centrifugal Compressor Unit . . . . . . . . . . . . . . . . . . . . . . . . . 500

FIGURE 12-9:

Schematic of Electric Turbine-Driven Centrifugal Compressor System . . . . . . . . . . . . . 500

FIGURE 12-10:

A Typical Compressor Speed-Torque Characteristic Curve . . . . . . . . . . . . . . . . . . . . . . 502

FIGURE 12-11:

A Typical Motor vs. Driven Compressor SpeedTorque Characteristic Curve . . . . . . . . . . . . . . . . 502

FIGURE 12-12:

Volume Reduction Per Stage, Compression Ratio=1.5. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 506

FIGURE 12-13:

Volume Reduction Per Stage, Compression Ratio=2.0. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 507

FIGURE 12-14:

Instrumentation for Monitoring and Diagnostics on an Industrial Type Gas Turbine Driver . . . . . 516

FIGURE 12-15:

Instrumentation for Monitoring and Diagnostics on a Centrifugal Compressor . . . . . . . . . . . . . . . 516

FIGURE 12-16:

Exhaust Temperature Profile for a Typical Gas Turbine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 518

FIGURE 12-17:

Comparison of Computed Gas Turbine Firing Gas Temperature . . . . . . . . . . . . . . . . . . . . . . . . . 519

FIGURE 12-18:

A Typical Compressor Characteristic Curve Showing a Relationship Between the Corrected Compressor Flow vs. the Corrected Polytropic Head . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 520

| xxv |

| Centrifugal Compressors |

| xxvi |

FIGURE 12-19:

A Typical Compressor Characteristic Curve Showing a Relationship Between the Corrected Compressor Flow vs. the Corrected Power . . . . 520

FIGURE 12-20:

A Typical Compressor Characteristic Curve Showing a Relationship Between the Corrected Compressor Flow vs. the Corrected Exit Temperature . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 521

FIGURE 12-21:

A Typical Compressor Characteristic Curve Showing a Relationship Between the Corrected Compressor Flow vs. the Corrected Discharge Pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 521

FIGURE 12-22:

Decomposition of a Time Signal Into a Sum of Oscillatory Functions from which a Spectrum is Obtained . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 523

FIGURE 12-23:

Limitations on Machinery Vibration Analysis System and Transducers . . . . . . . . . . . . . . . . . . . 526

FIGURE 12-24:

Vibration Nomograph and Severity Charts . . . . 526

FIGURE 12-25:

Life Cycle Costs of a Gas Turbine-Centrifugal Compressor Train . . . . . . . . . . . . . . . . . . . . . . . . 528

FIGURE 13-1:

Flow Equalizers and Straighteners . . . . . . . . . . . 541

FIGURE 13-2:

Inlet and Discharge Configuration . . . . . . . . . . . 542

FIGURE 13-3:

Short Inlet Pipe . . . . . . . . . . . . . . . . . . . . . . . . . . 542

FIGURE 13-4:

Diffusing Volute Discharge with Nonsymmetric Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 543

FIGURE 13-5:

Typical Test Cell Setup for Centrifugal Compressors . . . . . . . . . . . . . . . . . . . . . . . . . . . . 544

FIGURE 13-6:

Closed Loop Test Arrangement 1 . . . . . . . . . . . . 545

FIGURE 13-7:

Closed Loop Test Arrangement 2 . . . . . . . . . . . . 546

FIGURE 13-8:

Closed Loop for Condensable Gases . . . . . . . . . 547

FIGURE 13-9:

Traverse Points in Pipe . . . . . . . . . . . . . . . . . . . . 551

FIGURE 13-10:

Discharge Nozzle on an Open Loop, Subcritical Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 553

| List of Figures | FIGURE 13-11:

Discharge Nozzle on an Open Loop, Critical Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 554

FIGURE 13-12:

Inlet Nozzle on an Open Loop . . . . . . . . . . . . . . 554

FIGURE 13-13:

Typical Compressor Map . . . . . . . . . . . . . . . . . . . 557

FIGURE 13-14:

Relationship Between Polytropic and Adiabatic Isentropic Efficiency for a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 559

FIGURE 13-15:

Isentropic Z Function (Low Range) . . . . . . . . . . 560

FIGURE 13-16:

Isentropic Z Function (High Range) . . . . . . . . . 561

FIGURE 13-17:

Generalized Compressibility Factor for Simple Fluid . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 564

FIGURE 13-18:

Generalized Compressibility Factor Correction from Deviation from Simple Fluid . . . . . . . . . . . 565

FIGURE 13-19A:

Constants of Benedict-Webb-Rubin Equation . . . 566

FIGURE 13-19B:

Constants of Benedict-Webb-Rubin Equation . . . 566

FIGURE 13-19C:

Constants of Benedict-Webb-Rubin Equation . . . 567

FIGURE 13-19D: Constants of Benedict-Webb-Rubin Equation . . . 567 FIGURE 14-1:

Total Performance Based Condition Monitoring System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 573

FIGURE 14-2:

Machinery Life Cycle Characteristics . . . . . . . . . 573

FIGURE 14-3:

Breakdown Countermeasures. . . . . . . . . . . . . . . 574

FIGURE 14-4:

Responsibilities of the Operations and Maintenance Departments . . . . . . . . . . . . . . . . . 574

FIGURE 14-5:

Contributions of Various Major Components to Pump Downtime . . . . . . . . . . . . . . . . . . . . . . . 587

FIGURE 14-6:

Leakages in a Labyrinth Seal of a Centrifugal Compressor and Contamination of Gases . . . . . 593

FIGURE 14-7:

Areas of Foulant Build-up in a Centrifugal Impeller. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 594

FIGURE 14-8:

Compressor Casing Selection . . . . . . . . . . . . . . . 601

FIGURE 14-9:

Bundle Assembly Installation . . . . . . . . . . . . . . . 602

| xxvii |

| Centrifugal Compressors | FIGURE 14-10:

Schematic of Honeycomb Labyrinth Seal Used to Control Leakage Rates . . . . . . . . . . . . . . . . . . 606

FIGURE 14-11:

Thrust Forces in a Centrifugal Compressor. . . . 606

FIGURE 14-12:

Actual Probes for Thrust-Bearing Monitoring . . . 608

FIGURE 14-13:

Thrust Bearing Pads Typical Kingsbury Type Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 611

FIGURE 14-14:

Thrust Bearing Rocker Typical Kingsbury Type Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 612

FIGURE 14-15:

Diagram of Arrangement of Tilting Pad Thrust Bearing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 612

FIGURE 14-16:

RTD Embedded in Bearing Surface . . . . . . . . . . 616

FIGURE 14-17:

Temperature Distribution in Bearing Surfaces . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 616

FIGURE 14-18:

Vibration in the Covered Impeller Shroud and Disc . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 618

FIGURE 14-19:

Scalloped Rotor . . . . . . . . . . . . . . . . . . . . . . . . . . 618

FIGURE 14-20:

A Centrifugal Compressor Performance Map Corrected for Temperature and Pressure. . . . . . 624

FIGURE 14-21:

A Centrifugal Compressor Map Showing Operational Characteristics. . . . . . . . . . . . . . . . . 625

FIGURE 14-22:

Effect of Trimming the Exit Diameters of the Centrifugal Compressor Impellers . . . . . . . . . . . 626

FIGURE 14-23:

Effect of Number of Impellers on Stability and Capacity Limit . . . . . . . . . . . . . . . . . . . . . . . . 626

FIGURE 14-24:

Effect of Removal of Wheel from a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 627

FIGURE 14-25:

Increase in Flow After Removal of Wheel and Speed Up . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 628

FIGURE 14-26:

Removal of a Wheel from a Centrifugal Impeller and Replacing it With a Bridge . . . . . . . . . . . . . . 629

FIGURE 14-27:

Multistage Centrifugal Compressor . . . . . . . . . . 629

FIGURE 14-28:

The Effect of Inlet Throttling on a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 630

| xxviii |

| List of Figures | FIGURE 14-29:

A Typical Compressor Train Containing a Turbine and Two Compressor Stages . . . . . . . . . 632

FIGURE 14-30:

A Typical Unbalance Signature Plot . . . . . . . . . . 635

FIGURE 14-31:

A Typical Misalignment Signature Plot. . . . . . . . 635

FIGURE 14-32:

Gear Box Signature using a Proximity Probe Showing Only the Low End Frequency Component . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 637

FIGURE 14-33:

A Gear Box Signature Using an Accelerometer thus Examining the High Frequency Component . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 638

| xxix |

List of Tables TABLE 1-1:

Compressor Characteristics . . . . . . . . . . . . . . . . . . 5

TABLE 1-2:

Impeller Designs—Advantages and Disadvantages . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12

TABLE 1-3:

Applications of Centrifugal Compressors . . . . . . 14

TABLE 1-4:

Industrial Centrifugal Compressor Classification Based on Casing Design . . . . . . . . . . . . . . . . . . . . 18

TABLE 2-1:

Dimensions of Major Variables . . . . . . . . . . . . . . . 39

TABLE 3-1:

Vendor Requirements to be Provided by the User for a Compressor Train . . . . . . . . . . . . . . . . . . . . . 84

TABLE 3-2:

Points to Consider in a Centrifugal Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 86

TABLE 3-3:

Driver Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . 89

TABLE 3-4:

Points to Consider in a Gas Turbine. . . . . . . . . . 113

TABLE 3-5:

Points to Consider in a Steam Turbine. . . . . . . . 114

TABLE 4-1:

Advantages and Disadvantages of Various types of Blade Types . . . . . . . . . . . . . . . . . . . . . . . . . . . 140

TABLE 7-1:

Surge Control System . . . . . . . . . . . . . . . . . . . . . 253

TABLE 8-1:

High Temperature Alloys . . . . . . . . . . . . . . . . . . 345

TABLE 9-1:

Important Blade Material Characteristics . . . . . 387

TABLE 10-1:

Synchronous vs. Induction Motor Prices . . . . . . 403

TABLE 11-1:

Forces Acting on Rotor Bearing Systems. . . . . . 422

TABLE 11-2:

Characteristics of Forced and Self-Excited Vibration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 424

TABLE 11-3:

Characteristics of Rotor Instabilities . . . . . . . . . 433

TABLE 11-4:

Disc, Diaphragm, and Gear Couplings . . . . . . . . 471

TABLE 13-1:

Allowable Departure From Specified Operating Conditions For Class 1 Test . . . . . . . . . . . . . . . . 538

| xxxi |

| Centrifugal Compressors | TABLE 13-2:

Allowable Fluctuation of Test Readings During a Test Run for All Test Classes 1, 2, and 3 . . . . . 539

TABLE 13-3:

Allowable Departure From Specified Design Parameters for Class 1 and Class 3 Tests . . . . . . 539

TABLE 13-4:

Departure of Gas Properties from Perfect Gas Laws of Test and Specified Gas Permissible for Class 2 Tests . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 540

TABLE 13-5:

Analysis of Natural Gas . . . . . . . . . . . . . . . . . . . . 562

TABLE 14-1:

Benefits of Various Maintenance Systems Maintenance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 571

TABLE 14-2:

Performance Test Codes . . . . . . . . . . . . . . . . . . . 580

TABLE 14-3:

Mechanical Specifications . . . . . . . . . . . . . . . . . . 582

| xxxii |

INDEX

Index Terms

Links

A Acceleration transducers Acronyms Adiabatic efficiency

525-526 xxxiii-xxxiv 58

60-61

164-165

88

290-292

9-10

35-68

206

338-339

430-431

310-312 loss calculations Aeroderivative gas turbines installation cost

164-165 16 291-292

adaptation to remote control

292

maintenance

292

Aerodynamics (turbomachinery) properties

9-10

dimensional analysis

36-41

nozzles and diffusers

42-45

turbomachinery

45-46

equation of state

46

ideal gas

46-48

compressibility effect

48-49

energy equation

50

continuity equation

50-53

momentum equation

53-56

degree of reaction

56-57

efficiencies

58-68

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Aerodynamics (turbomachinery) (Cont.) speed

206

whirl

430-431

Aerothermal analysis

510

Air blower

73-75

Air compressors

69-70

inlet filter Air free of oil Air pollution problems

512 596

596 72 307-318

smoke

307

unburnt hydrocarbons

307

carbon monoxide

307

nitrogen oxides

307-318

nitrogenoxide prevention

310-312

dry low nitrogen combustor

312-318

Alarm/system logs

111

511-512

Alignment problems

425

603-604

shaft

425

bearingbracket

603-604

spectra

639-640

Alternating current induction motors

397-403

three phase

399

single phase

399

synchronous

400-403

Amplification factor Analogies (prediction) Analytical approach programs ANSI standards

109

415-416

36 36-37

513

513 102-103

This page has been reformatted by Knovel to provide easier navigation.

639-640

Index Terms Antisurge control

Links 228-229

methods

228-229

systems

251-269

constant speed

256-259

variable speed

259-265

flow/delta P system

259-261

discharge volume flow

260-264

API standards Applications (compressor)

101-106

112

115

117

119

13-14

73-79

73-75

gas recovery unit

75-76

reformer recycle compressor

75

ammonia plant

77

cryogenic expanderand recompression

77-78

refrigeration system

78-79

Arc of peripheral admission

367

full admission turbines

371-372 372-373

single-stage/simple-impulse turbines

373-374

velocity-pressure stage combination

374 374-375

pressure-stage/Rateau-type impulse turbine

376

reaction type

377

general flow arrangement ASME standards/codes Asymmetry (rotor) Availability (compressor) Axial clearances

371-380

372

turbine stages in series multi-stage impulse-type turbines

264-265

97-98

air blower

partial admission turbines

251-269

377-380 22

93-96

425 70 620-621

This page has been reformatted by Knovel to provide easier navigation.

99-101

Index Terms

Links

Axial flow

368

Axial-flow compressors

2-8

Axial-flow turbines

24

297-299

326-327

B Backpressure Back-to-back compressors

64 29-31

Backward-curved vane impeller

140

Balancing

109

Balje slip factor

149

Barrel-type compressor handling

19-21 600-601

inner casing alignment

601

internal leakage

602

Bearing bracket alignment material problems Bearing lubrication oil contamination Bearings

600-602

603-604 603-604 466 438-466

488-489

608-609 instabilities

438-440

functions

440-441

journal bearings

441-444

tilting pad bearings

443-447

load

445-446

materials

447-448

thrust bearings

448-466

lubrication oil

466

bracket alignment

603-604

maintenance

608-609

Benedict-Webb-Rubin equation

603-604

565-567

This page has been reformatted by Knovel to provide easier navigation.

603-604

Index Terms

Links

Blade angle (surge)

242-243

Blade coatings

347-349

Blade cooling concepts

332-350

convection

333

impingement

333

film

333

transpiration

334

steam

334

design

334

convection/impingement cooling(strutinsert)

599-600

334-336

film/convection cooling

334

336-337

steam cooling

334

337-338

Blade damping

388-389

Blade loading analysis

149-152

method

167

velocity gradient in meridional plane

168-175

derivation of equations (blade to-blade solution)

175-179

Blade materials

342-346

material characteristics

387-389

construction materials

389-392

surface treatments

392

mechanical efficiency

392

turbine efficiency

392

Blades (turbine) loading analysis design

166-179

387-392

149-152

166-179

242-243

332-350

382-392

599-600

149-152

166-179

152

167

angle

242-243

cooling

332-350

materials

342-346

387-392

This page has been reformatted by Knovel to provide easier navigation.

175-179

Index Terms

Links

Blades (turbine) (Cont.) coatings

347-349

high pressure

382-383

intermediate pressure

382-383

low pressure

385-387

damping

388-389

Blade-to-blade design solution

152

599-600

167

175-179

175-179

Boundary-layer development

144-146

Butterfly valve

247-248

504

C Campbell diagram

433-438

Can-annular/annular combustors

304-305

Capacity control

72

surge control

253-255

Carbon monoxide

307

Cascade diffuser

186-188

Case studies (instrumentation controls)

529-531

Casings compressor alignment Catalytic combustion

253-255

17-18

83

17-18

83

601 318-331

features

320-322

turbine expander section

322-325

axial-flow turbines

326-327

impulse turbine

327-329

reaction turbine

329-331

This page has been reformatted by Knovel to provide easier navigation.

601

Index Terms Centrifugal compressor components

Links 124-151

impeller

128-131

inlet guide vanes

132-137

inducer

137-139

centrifugal section of impeller

138-142

slip factor

142-143

impeller slip

143-149

impeller design

149-151

Centrifugal compressor standards

103-106

API Std 672 (1996)

103-106 2-5

8-34

60-63

97-98

103-106

124-151

applications

13-14

gas turbines

15-17

process industry

17-22

internal configuration

22-31

impeller fabrication

32-34

selection

60-63

standards

103-106

components

124-151

Centrifugal pumps

148-149

103-106

API Std 617 (1995) Centrifugal flow compressors

138-151

102

Centrifugal section (impeller)

138-142

Classification (steam turbines)

367-370

steam flow direction

367-369

steam passage between blades

367

369-370

arc of peripheral admission

367

turbine stages in series

367

372-373

stage type

367

373-377

general flow arrangement

367

377-380

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Clearance checks

609-610

Clearance loss

211-212

Closed-face impeller

32

Closed-loop control system

492-493

Closed-loop piping

545-547

Coating materials (blade)

347-349

Combustion analysis Combustor design

304-306 304-305

tubular/side combustors

305-306

Combustors

300-306

design

304-306

nitrogen oxides

312-318 48-49

Compressibility factor

46-47

Compressor aerodynamics

35-68

dimensional analysis

36-41

nozzles and diffusers

42-46 46

ideal gas

46-48

compressibility effect

48-49

energy equation

50

continuity equation

50-53

momentum equation

53-56

degree of reaction

56-57

efficiencies

56-68

Compressor applications

312-318

378-379

Compressibility effect

equation of state

599-600

510

canannular/annular

Compound flow/tandem compound

34

73-79

air blower

73-75

gas recovery unit

75-76

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Compressor applications (Cont.) reformer recycle compressor

75

ammonia plant

77

cryogenic expander and recompression

77-78

refrigeration system

78-79

Compressor definition and standards overview air compressors

69-122 69 69-70

conditionand life assessment

70

availability

70

reliability

71

power requirement

71

floor space

71

foundation

71

control of capacity

72

air free of oil

72

life cycle cost

72

load factor

72

multiple stages

72-73

process gas compressor

73

applications of compressors

73

air blower

73-75

gas recovery unit

75-76

reformer recycle compressor

75

ammonia plant

77

cryogenic expander and recompression

77-78

refrigeration system

78-79

gas characteristics

79

compressor sizing

80-82

compressor casings

83

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Compressor definition and standards (Cont.) compressor evaluation

84-86

plant location/site configuration

86-87

driver selection

87-90

fuel type condenser type

90 90-91

enclosures

91

plant operation mode

91

start-up techniques

92

performance standards

92-96

mechanical parameters

96-103

centrifugal compressor standards

103-106

drives

107-120

references

121-122

Compressor design characteristics

123-181

centrifugal compressor components

124-151

one dimensional flow

151-165

three dimensional blade loading analysis

166-179

references

180-181

Compressor drives Compressor failures barrel-type compressor

87-90 600-602 600-602

Compressor load

407-409

Compressor loss

208-223

externallosses

211-213

rotor losses

214-218

stator losses

218-223

Compressor mechanics

1-34

selection

2-5

axial-flow compressors

6-8

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Compressor mechanics (Cont.) centrifugal flow compressors

8-34

applications

13-14

gas turbines

15-17

process industry

17-22

internal configuration

22-31

impeller fabrication

32-34

Compressor module (simulation)

274-275

Compressor performance testing

535-568

test planning

536-537

classification of tests

537-538

piping arrangements

538-541

inlet piping

541-543

discharge piping

543-544

closed-loop piping

545-547

data acquisition

547-548

pressure measurement

548-551

temperature measurement

550-552

flow measurement

552-554

power measurement

555

speed measurement

555-556

test procedures

556-557

test computations

557-567

references Compressor performance

568 63-68

84-86

205-237

535-568

592-596

600-602

characteristics

63-68

evaluation

84-86

calculations

205-208

operation

233-237

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Compressor performance (Cont.) testing

535-568

problems

592-596

failures

600-602

Compressor problems

592-596

fouling

592-596

fouling prevention

594-596

failures

600-602

Compressor rerating

623-624

Compressor seals

619-621

radial clearances

620

quality control

620

axial clearances

620-621

seal design

621

training for seal maintenance

621

rules of thumb

621

Compressor system simulation

600-602

229

271-279

20

22

518-529

70

509-517

586-588

monitoring system

509-517

586-588

Condition monitoring system

509-517

529-531

Compressor train gas turbine

518-528

steam turbine

528-529

Compressor types

297-299

Compressor water wash (online)

597-599

Compressor-turbine matching

350-356

Computer fluid dynamics Condenser type Condition and life assessment

166 90-91

586-588 aerothermal analysis

510

512

This page has been reformatted by Knovel to provide easier navigation.

570

Index Terms

Links

Condition monitoring system (Cont.) combustion analysis

510

vibration analysis

510

mechanical analysis

510

diagnosis

510

trendin gand prognosis

510

what if analysis

510-511

gas/fluid types

514

fuel type

514

materials in hot sections

514

performance maps

514

instrumentation location/types typical results case study Conservation law Constant speed antisurge control Constants

514-517 517 529-531 50-51 256-259 xxxv-xxxvii

speed

xxxv

force

xxxv

mass

xxxv

pressure

xxxv

energy and power

xxxvi

entropy

xxxvi

specific heat

xxxvi

gas constant

xxxvi

universal gas constant

xxxvi

Newton’s proportionality constant k

xxxvii

speed of light

xxxvii

Planck constant

xxxvii

gravitational constant

xxxvii

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Constants (Cont.) Avogadro constant

xxxvii

Boltzmann constant

xxxvii

normal mole volume

xxxvii

Contamination (seal)

465-466

process gas

466

bearing lubrication oil

466

seal gas supply

466

Continuity equation

42

Continuous diffuser return channel Continuous-flow compressors

45

50-53

274-279

349-350

198-201 2

Continuously-lubricated couplings

476-477

Control equipment

265-269 491-533

Control factors (compressor)

504

butterfly valve

504

adjustableinlet guide vane

504

Control hardware evaluation

276

Control system module (simulation)

274-275

Control systems (compressor)

274-275

simulation Control valve Control-vortex prewhirl

278 492-493

274-275 265-269

276-279

136

Convection cooling

333-338

Cooled-turbine aerodynamics

338-339

Cooling design (turbine blade)

334-338

Coriolis circulation

143-144

Corrosion fatigue resistance

388

Corrosion resistance

388

This page has been reformatted by Knovel to provide easier navigation.

540-541

Index Terms Cost (life cycle) analysis Coupled gas/diesel engine Coupling maintenance Couplings (gear) oil filled

Links 72

492

527-528

570

513 89 630-631 11

103

489-490

630-631

475-476

continuously lubricated

476-477

metal diaphragm

477-479

metal disc

479-481

maintenance

630-631 387

Critical speed

420-421

Cross compound flow

378-380

Curtis-type impulse turbine

469-481

475

grease packed

Creep strength/creep-fatigue resistance

513

375

D Data acquisition

547-548

Data management

513

Defects

572

Degree of reaction Design characteristics

56-57 123-181

183-204

413 centrifugal compressor

123-181

diffuser

183-204

surge

240-248

rotating equipment Detection and control (surge)

413 251-253

This page has been reformatted by Knovel to provide easier navigation.

240-248

Index Terms Diagnosis (monitoring) Diesel engine

Links 510

512

89

Diffuser channel

194-201

Diffuser design

183-204

performance characteristics

183-189

radial vaneless diffuser

189-191

vaned diffuser

191-198

continuous diffuser return channel

198-201

scroll and volutes

201-203

references

204

Diffusers and nozzles

9-10

13

183-204 diffuser design

183-204

channel

194-201

Diffusion blading loss (rotor)

216-217

Dimensional analysis

36-41

Dimensional system

38-39

Directional solidification

344

Dirt in oil

615

Disc friction loss

213

Discharge piping

543-544

Discharge volume flow

260-264

P and T compensated

260-261

non-compensated

262-264

Displacement transducers

524

Double flow

380

Double-entry inlet

132

Downtime

572

Drawing/print file

590

263

This page has been reformatted by Knovel to provide easier navigation.

42-45

Index Terms Drive types

Links 107-120

gas turbines

107-113

steam turbines

112-115

gears

115-117

lubrication systems

117-118

vibration measurements

119-120

specifications Drives selection

120 87-90 87-90

types

107-120

gas turbines

107-113

steam turbines

112-115

gears

115-117

lubrication systems

117-118

vibration measurements

119-120

specifications Dry gas seals

120 462-466

operatingrange

464

materials

464

systems

464-465

degradation

465-466

process gas contamination

466

bearing lubrication oil contamination

466

seal gas supply contamination

466

Dry low emission combustor

312-318

Dry-friction whirl

428-429

Ductility/impact strength Dynamic simulation (compressor)

107-120

388 2

229

This page has been reformatted by Knovel to provide easier navigation.

271-279

Index Terms

Links

E Efficiencies adiabatic efficiency

56-68 58

polytropic efficiency

59-60

centrifugal compressors election

60-63

compressor performance characteristics

63-68

Electric motors

87-91

98-99

395-412

500-501 enclosure

91

characteristics

395-396

speed/type

397-403

power-factor correction

404-407

inertial load (compressor)

407-409

applications

410-412

references

412

Electric turbine

500-501

Enclosures

91

turbine

91

motor

91

Energy balance Energy equation

412

412 412

164-165 42

45

175 Energy loss

208-223

Enthalpy

58

Entropy

58

Environmental considerations Equation of state Equipment efficiency/effectiveness Erosion resistance

110

287-288

45-46 572-575 388

This page has been reformatted by Knovel to provide easier navigation.

50

Index Terms

Links

Euler equation

141

Exhaust temperature

518

Exit loss (stator)

222-223

Expander section (turbine)

322-325

External causes/effects (surge)

250-251

External losses

211-213

clearance eloss disc friction loss Extraction flow

211-212 213 380

F Factory repair

622

Failure initiation (thrust bearing)

613

Failure protection (thrust bearing)

615-617

Fatigue resistance

387-388

Fatigue strength

388

Film cooling

333

Firing temperature

519

Flexible supports (machinery) Floor space Flow arrangement (steam turbine)

418-421 71 367

Flow delivered

627-630

Flow measurement

552-554

flowmeters Flow/delta P system Fluid catalytic cracking

336-338

377-380

552-554 259-261

264-265

73-76

FCC unit

73-75

Fluid slugging

613

Fluid trapped whirl

432-433

Force applied (rotor)

422-423

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Force generated (rotor motion)

422-423

Force transmitted (casing/foundations)

422-423

Forced/resonant vibration

424-426

Forced-vortex prewhirl Foreman Forward-curved vane impeller

136 580-581 140

Fouling (compressor)

592-596

prevention

594-596

Fouling prevention

594-596

process gas compressors

594-595

online solvent injection

595-596

Foundation repair/rehabilitation

71 631

Fourier transform analysis

523

Free-vortex prewhirl

135

Froude number Fuel systems Fuel type

631

40-41 111-112 90

99

112

514 Full admission turbines

371-372

G Gas composition (surge)

248-251

Gas engine

89

100

Gas properties

79

96

248-249

507

514

540

Gas recovery unit Gas turbine components

75-76 297-318

compressor types

297-299

axial-flow compressors

297-299

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Gas turbine components (Cont.) regenerators

299-300

combustors

300-306

air pollution problems

307-318

Gas turbine compressor train

518-528

spectrum analysis

522-523

vibration measurement

524

displacement transducers

524

velocity transducers

524-525

acceleration transducers

525-526

life cycle costs

527-528

Gas turbine materials directional solidification single-crystal blades Larson-Miller parameter coatings Gas turbines

312-318

342-349 344 344-345 346 347-349 6

8

15-17

87-89

91

98-101

107-113

281-357

497-498

518-528 industrial/heavy-duty

288-290

aeroderivative

290-292

medium range

292-294

small

294-296

major components

297-318

catalytic combustion

318-331

turbine blade cooling concepts

332-350

materials

342-349

matching gas turbines and compressors

350-356

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Gas turbines (Cont.) references compressor train Gear coupling thrust Gear noise Gears/gearing

housings

357 518-528 614 640-641 24

27-29

101-102

110-111

115-117

466-487

489

614

634-641

124-126

132-137

468

couplings

469-481

lubrication oil system

481-486

lubricant selection

486-487

coupling thrust

614

gear boxes

634-639

noise

640-641

General flow arrangement (turbine) single flow/single casing

377-380 378

compound flow/tandem compound

378-379

cross compound flow

378-380

double flow

380

extraction flow

380

GPA Standard 2145-94

96

Graphic user interface

511

Grease-packed couplings Guide vanes

475-476 11 504

adjustable

504

H Head (velocity/pressure)

55-57

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Head calculation

Links 207

Head delivered

625-626

Heat rate

365-366

Heavy-duty gas turbines

288-290

High-pressure turbine blades

382-383

Historical data management Horizontally-split compressor Hot gas expanders mixed-flow turbine

513 19-21 339-343

528-529

341-343

Hot section materials

514

Housings (gear)

468

Hub-shroud design

167

Hydrostatic test

112

Hysteretic whirl

426-428

I Ideal gas

46-48

Impact strength

388

Impeller design

12

Impeller slip

143-149

Coriolis circulation

143-144

boundary-layer development

144-146

leakage

146

number of vanes

146

vane thickness

146-147

Stodola slip factor

148

Stanitz slip factor

148-149

Balje slip factor

149-151

149

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Impellers

Links 1

11-12

23-24

32-34

124-131

138-164

617-619 design

12

materials

32

fabrication

32-34

centrifugal section

138-142

causes of slip

143-149

problems

617-619

Impingement cooling

333-335

Impulse-reaction turbine Impulse-type turbines

370 327-329

impulse-reaction turbine

370

Curtis type

375

Rateau type

376

Incidence loss (rotor)

215-216

Inducer

137-139

Induction motor applications

410-412

number of starts motor enclosures applications

410-412

Industrial/heavy-duty gas turbines

288-290

Inertial load (compressor)

407-409

motor burnouts

408-409

410-412

596 11

124-126

504 adjustable Inlet piping

373-376

412 397-403

Inlet guide vanes

370

410-411

Induction motors

Inlet filter

149-151

504 541-543

This page has been reformatted by Knovel to provide easier navigation.

132-137

Index Terms

Links

Inlet pressure/temperature changes

624

Inlet velocity

132

Inspections condition and life assessment

586-588 586-588

Installation cost (gas turbine)

291-292

Installation defects

632-633

Instrumentation controls

349-350

blade cooling

349-350

control systems

492-493

start-up procedures (compressor)

493-503

control factors

504-509

total condition monitoring system

509-511

monitoring software

511-517

location/types

514-517

gas turbine compressor train

518-528

steam turbine compressor train

528-529

case studies

529-531

references

532-533

Intermediate-pressure turbine blades Internal configuration (compressor) Internal losses (surge)

24 22-31 237-251 240-248

gas composition

248-251

external causes/effects

250-251

ISO standards

27-29

382-383

design factors

Isentropic efficiency

491-533

504

modes of operation

Integral gears

134

58-59 96

98

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

J Journal bearings

441-444

K Kinetic energy

42

L Labyrinth seals

107-108

Laminar flow control

145-146

Larson-Miller parameter Lead machinist Leakage

346 580-581 146

602

619-621

509-511

513-517

72

492

513

527-528

570

impeller

146

compressor

602

seal Life assessment

453-456

619-621 70 586-588

Life cycle cost analysis Lift (airfoil) Load characteristics Load factor

513 230-231 66-67 72

Loss calculations

164-165

Low-pressure turbine blades

385-387

Lubricant selection

486-487

Lubrication systems

117-118

481-486

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Lubrication

Links 102

109-110

481-487

490

systems

117-118

481-486

lubricant selection

486-487

M Mach number Machinist training Maintenance communications

40-41 580-581 583-584 589-592

operation/service manuals

590

drawing/print file

590

training materials

590-591

training in-house rotor crew

591

pocket guide

591

newsletter/internal memos

591-592

Maintenance department requirements

579-584

training of personnel

579

types of personnel

580-581

types of training

581-584

Maintenance engineer

580

Maintenance program

575-579

implementation

576-579

Maintenance scheduling

588-589

Maintenance techniques

292

gas turbine

583-584

569-641

292

equipment efficiency/effectiveness

572-575

organization structures (maintenance program)

575-579

maintenance department requirements

579-584

tools and shop equipment

584-592

This page has been reformatted by Knovel to provide easier navigation.

117-118

Index Terms

Links

Maintenance techniques (Cont.) scheduling

588-589

compressor problems

592-596

fouling prevention

594-596

air compressors

596

onlin ecompressor water wash

597-599

compressor blade coating

599-600

compressor failures

600-602

bearing bracket alignment

603-604

rotor thrust calculations

604-610

thrust bearing maintenance

610-613

thrust bearing failure

613-617

impeller problems

617-619

compressor seals

619-621

rules of thumb

621

rotor repairs

622-623

compressor rerating

623-624

inlet pressure/temperature changes

624

head delivered increase/decrease

625-626

flow delivered increase/decrease

627-630

coupling maintenance

630-631

turbomachinery foundation repair/rehabilitation installation defects

631 632-633

mass increase

634

rigidity increase

634

gears/gear boxes

634-639

synchronous and harmonic spectra

639-641

references Mass flow Mass increase

641 205-206 634

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Matching (turbine/compressor)

350-356

Material characteristics (blade)

342-346

creep strength/creep-fatigue resistance

387

tensile strength

387

corrosion resistance

388

ductility/impact strength

388

fatigue strength

388

corrosion fatigue resistance

388

notch sensitivity

388

erosion resistance

388

blade damping

388-389

blade construction

389-392

Materials (hot section)

514

Materials (seals)

464

Mathematical model

273-276

process module

273-275

compressor module

274-275

control system module

274-275

Mechanical analysis

510

Mechanical efficiency

392

Mechanical parameters/specifications references Mechanical seals

96-103

120-122

457-461 458

pressure

458-459

temperature

459

lubricity

459

abrasion

459-460

toxicity

512

121-122

product

corrosion

387-392

460-461

460 460-461

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Mechanical test

Links 112

Medium-range gas turbines

292-294

Metal diaphragm couplings

477-479

Metal disc couplings

479-481

Metallurgy

32

Microturbine

17

Millwright

581

Mixed flow

369

Mixed-flow compressors Mixed-flow turbine

2 341-343

Momentum equation

45

Monitoring software

511-517

graphic user interface

511

alarm/system logs

511-512

perform ancemaps

512

analysis programs

512

aerothermal analysis

512

mechanical analysis

512

diagnostic analysis

512

optimization analysis

512-513

life cycle analysis

513

historical data management

513

condition monitoring system results Monitoring system

513-517 517 509-511

aerothermal analysis

510

combustion analysis

510

vibrationanalysis

510

mechanical analysis

510

diagnosis

510

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Monitoring system (Cont.) trending and prognosis what if analysis Motor speed/type

510 510-511 397-403

alternating current/squirrel cage induction motors

397-400

synchronous alternating-current motors

400-403

wound-rotor induction motors

403

Motor starting method

406-407

Motor torque selection

405-406

Motor voltage

406-407

Motor-driven unit

500-501

Motors (electric)

395-412

characteristics

395-396

speed/type

397-403

power-factor correction

404-407

size selection

404-405

torque selection

405-406

starting method

406-407

voltage

406-407

inertial load (compressor)

407-409

burnouts

408-409

applications

410-412

references Multistage compressor Multi-stage impulse-type turbines

500-501

413 28

68

374

N Natural frequency Natural gas

413-415 96

This page has been reformatted by Knovel to provide easier navigation.

72-73

Index Terms

Links

Navier-Stokes equation

53-56

NEMA MG 2

98-99

Newsletter/internal memos

591-592

Newtonian fluid

57

Nitric acid train

77

Nitrogen oxides combustor

312-318

Nitrogen oxides

307-318

prevention

310-312

dry low combustor

312-318

Noise reduction Non-compensated flow

110 262-264

Non-contacting seals

452

Notch sensitivity

388

Nozzles (turbine)

15

and diffusers

42-45

42-45

O Off design operation Off-design performance characteristics

614 205-223

compressor performance

205

performance calculations

205-208

compressor loss

208-223

external losses

211-213

rotor losses

214-218

stator losses

218-223

Oil pressure loss Oil whirl Oil-filled couplings One-dimensional flow stream tube coordinates

615 429-430 475 151-165 151-165

This page has been reformatted by Knovel to provide easier navigation.

384-385

Index Terms On-stream operation Open-face impeller Open-loop control system Operating margin

Links 504-509 32-33 492 64

Operating range (seals)

464

Operation/service manuals

590

Operational modes

504-509

startup control action

504-508

shutdown control action

508-509

Optimization analysis

512-513

Organization structures (maintenance program)

575-579

implementation Oxides of nitrogen

138

576-579 307-318

prevention

310-312

dry low combustor

312-318

P P and T compensated flow

260-261

Parallel compressors stability

269-279

dynamic simulation

271-279

control hardware

276

Parsons turbine

370

Partial admission turbines

372

263

278

Performance based total producti vemaintenance system Performance calculations (compressor) corrected compressor mass flow

570-575 205-208 205-206

corrected compressor volume flow

206

corrected (aerodynamic)speed

206

corrected power

206

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Performance calculations (compressor) (Cont.) corrected head (adiabatic orpolytropic)

207

corrected pressure

207

corrected temperature

207

performance map

207-208

Performance characteristics (diffusers)

183-189

Performance characteristics (off design)

205-223

compressor

205-208

calculations

205-208

compressor loss

208-223

external losses

211-213

rotor losses

214-218

Performance characteristics

63-68

diffusers

183-189

off design

205-223

Performance standards ASME PTC 19.1 (1988)

92-96

183-189

121

93

ASME PTC 10 (1997)

93-94

ASME PTC 22 (1997)

94

ASME PTC 6 (1996)

95

ASME PTC 12.2 (1983)

95-96

ISO 6976-1983(E)

96

GPA Standard 2145-94

96

references Performance test Performance-based maintenance program implementation Periodic loading

121 93-96

112

575-579 576-579 426

This page has been reformatted by Knovel to provide easier navigation.

205-223

Index Terms Personnel training

Links 579

581-584

621 update training practical training

581-583 583

machinist

583-584

materials

590-591

rotor crew

591

seal maintenance

621

Personnel types maintenance engineer

580-581 580

foreman

580-581

lead machinist

580-581

machinist

580-581

millwright Piping arrangements (compressor) Plant location/site configuration Plant operation mode Pocket guide Polytropic efficiency Polytropic head Positive-displacement compressor Potential flow lines Power calculation Power loss (thrust bearing)

581 538-541 86-87 91 591 59-60 60 1-5 151-152 206 451-452

Power measurement

555

Power requirement

71

Power-factor correction (motors)

62-63

404-407

motor size selection

404-405

motor torque selection

405-406

motor voltage and starting method

406-407

This page has been reformatted by Knovel to provide easier navigation.

590-591

Index Terms Practical training

Links 583

Prandtl number

40-41

Pressure ratio

65-67

Pressure transducers Pressure

349 47

49

52

65-67

207

349

521

548-551

static

47

49

total

47

49

ratio

65-67

calculation

207

transducers

349

measurement

548-551

Prevention (nitrogen oxides)

310-312

Prewhirl (impeller inlet)

132-137

free vortex

135

forced vortex

136

control vortex

136

Process gas compressors

73

Process gas contamination Process industry Process module (simulation) Proportional integral derivative

594-595

466 17-22 273-275 493

Q Quality control (seal)

620

R Radial clearances

620

This page has been reformatted by Knovel to provide easier navigation.

52

Index Terms Radial flow Radial vane impeller

Links 368-369 140

Radial vaneless diffuser

189-191

Rankine cycle

360-363

Reaction turbine

329-331

370

377

121-122

180-181

204

280

357

393-394

413

488-490

532-533

568

641

Reblading (rotor)

623

Reciprocating compressor

1-2

Recirculation/wake mixing loss (stator) Recompression processing References

Reformer recycle compressor Regenerative-reheat cycle

218-219 77-78

75-76 363-366

heat rate

365-366

steam rate

365-366

turbine components

366

Regenerators

299-300

Reheat cycle

363-366

heat rate

365-366

steam rate

365-366

turbine components Reliability Remote control (gas turbine) Repair rotor foundation

366 71 292 622-623

631

622-623 631

Rerating (compressor)

623-624

Response time (surge control)

268-269

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Return channel Reynolds Number Rigid supports (machinery)

Links 198-201 40-41 417-418

Rigidity increase

634

Rise to surge

236

Rotary compressor

1-2

Rotating equipment

413-490

rotor dynamics

413-440

design considerations

413

natural frequency

413-414

unbalances

414-416

application to rotating machines

417-421

forces acting on rotor-bearing system

421-440

bearings

438-466

seals

452-466

gears/gearing

466-487

references

488-490

Rotating machines

417-421

rigid supports

417-418

flexible supports

418-421

Rotating stall Rotor crew training Rotor dynamics design considerations

420-421

420-421

230-233 591 413-440 413

natural frequency

413-415

unbalances

414-416

application to rotating machines

417-421

forces acting on rotor bearing system

421-440

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Rotor losses shock

Links 214-218 214

incidence

215-216

diffusion blading

216-217

skin friction Rotor motion

218 423

Rotor repair

622-623

factory

622

local shops

622

user’s shop

623

reblading

623

Rotor system instabilities self-excited Rotor thrust calculations

423-424 426-433 604-610

thrust collar designs

607-608

bearing maintenance

608-609

clearance checks

609-610

Rotor-bearing system (forces)

421-440

transmitted to casing/foundations

423

generated by rotor motion

423

applied to rotor

423

system instabilities

423-424

forced/resonant vibration

424-426

periodic loading

426-433

426

self-excited instabilities

426-433

Campbell diagram

433-438

bearing and shaft instabilities

438-440

Rules of thumb (seals)

426-433

621

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

S Scroll and volutes

201-203

Seal abrasion

459-460

Seal arrangement considerations

461-466

equipment dry gas seals Seal corrosion environment

462 462-466 460-461 461

Seal degradation

465-466

process gas

466

bearing lubrication oil

466

seal gas supply

466

Seal design

621

Seal gas supply contamination

466

Seal leakage

619-621

radial clearances

620

quality control

620

axial clearances

620-621

seal design

621

training for seal maintenance

621

rules of thumb

621

Seal lubricity

459

Seal maintenance training

621

Seal pressure

458-459

Seal product

458

Seal temperature

459

Seal toxicity

460

460-461

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Seals (thrust bearing) noncontacting seals

Links 452-466 452

labyrinth seals

453-456

mechanical seals

457-461

abrasion

459-460

corrosion

460-461

seal environment

461

seal arrangement

461-466

systems

464-465

degradation

465-466

leakage

619-621

design

621

maintenance training

621

Selection (compressor) centrifugal compressor Self-excited instabilities

2-5

60-63

60-63 108

hysteretic whirl

426-428

dry-friction whirl

428-429

oil whirl

429-430

aerodynamic whirl

430-431

fluid trapped inrotor whirl

432-433

Seminars

592

Service manuals

590

Shaft instabilities

425

Shock loss (rotor)

214

Shop repair

489

426-433

438-440

622-623

Side-loaded compressors

30-31

Simple-impulse turbines

373-374

This page has been reformatted by Knovel to provide easier navigation.

619-621

Index Terms Simulation (compressor system)

Links 229

process module

273-275

compressor module

274-275

control system module

274-275

example

274-279

Single flow/single casing Single-crystal blades Single-entry inlet Single-stage turbines

378 344-345 132 373-374

Site configuration

86-87

Sizing (compressor)

80-82

Skin friction loss (rotor)

218

Slip causes (impeller)

142-149

slip factor

142-143

Corioliscirculation

143-144

boundary-layer development

144-146

leakage

146

number of vanes

146

vane thickness

148

Stanitz slip factor

148-149 149

Slip factor

142-143

Small gas turbines

294-296

Smoke problem Software (condition monitoring) graphic user interface

146-149

146-147

Stodola slip factor Balje slip factor

271-279

146-149

307 511-517 511

alarm/system logs

511-512

performance maps

512

analysis programs

512

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Software (condition monitoring) (Cont.) aerothermal analysis

512

mechanical analysis

512

diagnostic analysis

512

optimization analysis

512-513

life cycle analysis

513

historical data management

513

condition monitoring system Solids buildup/plugging

513-517 614

Solvent injection (online)

595-596

Spare parts inventory

585-586

Specifications

96-103

drive

120

performance

121

mechanical

121-122

Spectrum analysis (gas turbine compressor)

522-523

Speed measurement

555-556

Speed-torque curve

501-503

Splitter blade

138-139

Squirrel cage induction motors

397-400

three phase

399

single phase

399

Stability (surge control systems/parallel compressors) dynamic simulation (compressor system) control hardware evaluation Stage mismatching Stage type

269-279 271-279 276

278

507-508 367

Stall (rotating)

230-233

Stanitz slip factor

148-149

Starting system

120-122

373-377

111

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Start-up procedures

Links 92

111

410-411

47

49

52

48-49

51-52

493-503 turbine

92

system

111

motor

410-411

compressor

493-503

Static pressure Static temperature Stator losses recirculation/wake mixing

218-223 218-219

vaneless diffuser

220

vaned diffuser

221

exit loss

222-223

Steam cooling

333-334

Steam engine Steam flow direction axial flow radial flow

89 367-369 368 368-369

tangential flow

368

mixed flow

369

Steam passage between blades

369-370

impulse turbine

370

reaction turbine

370

Parsons turbine

370

impulse-reaction combination turbine

370

Steam rate

365-366

Steam turbine classification

367-370

Steam turbine compressor train

528-529

hot gas expanders

528-529

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Steam turbines

Links 87-91

98

112-115

359-394

495-496

498

501

528-529

Rankine cycle

360-363

regenerative-reheat cycle

363-366

components classification arc of peripheral admission

366 367-370 367

characteristics

380-381

turbine blades

382-392

turbine nozzles

384-385

advantages/disadvantages

392-393

references

393-394

compressor train

528-529

Steel alloy/metallurgy

32

Stodola slip factor Stone walling Straight flow-through compressors Stream tube coordinates

148 65 151-165 151-152

adiabatic efficiency (loss calculations)

164-165

Streamlines

151-152

Strut insert design

334-338

Supersonic diffuser

193-194 392

Surge control systems stability

269-279

dynamic simulation

271-279

control hardware

234

29-30

streamlines

Surface treatments

371-380

276

278

This page has been reformatted by Knovel to provide easier navigation.

498-499

Index Terms Surge control systems

Links 251-279

detection and control

251-253

capacity control methods

253-255

antisurge control

256-265

system piping/configuration

264-268

response time

268-269

stability

269-279

Surge control

225-280

definition

229-230

rotating stall

230-233

compressor operation

233-237

effects of internal losses

237-251

design factors

240-248

systems

251-279

stability of systems

269-279

references

280

Surge design factors

240-248

Surge/pumping limit

233-234

Surging (compressor)

188-189

225-280

614 surge control

225-280

Synchronous alternating-currentmotors

400-403

Synchronous and harmonic spectra

639-641

alignment problems

639-640

gear noise

640-641

System piping/configuration (surge)

264-268

T Tandem inducer

138-139

Tangential flow

368

This page has been reformatted by Knovel to provide easier navigation.

508-509

Index Terms

Links

Temperature static total effect

48-49

51-52

67-68

207

518-522

550-552

48-49

51-52

48

52

67-68

calculation

207

exhaust

518

firing

519

measurement Tensile strength

550-552 387

Terminology (components)

124-151

Test classification

537-538

Test computations

557-567

Test planning

536-537

Test procedures

556-557

Thermodynamic properties Three-dimensional blade loading analysis method

8 166-179 167

velocity gradient in meridional plane

168-175

derivation of equations (blade-to-blade solution)

175-179

Thrust bearing design factors

9-10

450-466

power loss

451-452

seals

452-466

Thrust bearing failure

613-617

failure initiation

613

fluid slugging

613

solids buildup/plugging

614

off design operation

614

compressor surging

614

gear coupling thrust

614

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Thrust bearing failure (Cont.) dirt in oil

615

oil pressure loss

615

failure protection

615-617

Thrust bearing maintenance

610-613

Thrust bearing power loss

451-452

Thrust bearing seals

452-466

non contacting seals

452

labyrinth seals

453-456

mechanical seals

457-461

environment arrangement considerations Thrust bearings

461 461-466 448-466

functions

448-450

power loss

451-452

design factors

450-466

seals

452-466

maintenance

610-613

failure

613-617

Thrust collar designs

607-608

Tilting pad bearings

443-447

Time lapse (surge control)

268-269

Tools and shop equipment

584-592

spare parts inventory

585-586

inspections

586-588

maintenance scheduling

588-589

maintenance communications

589-592

seminars and workshops Total performance condition monitoring Total pressure

610-617

592 492 47

49

This page has been reformatted by Knovel to provide easier navigation.

Index Terms Total productive maintenance program implementation Total temperature Training (personnel)

Links 575-579 576-579 48

52

579

581-584

621 update training practical training

581-583 583

machinist

583-584

materials

590-591

rotor crew

591

seal maintenance

621

Training materials Transducers

590-591 349

pressure

349

displacement

524

velocity

524-525

acceleration

525-526

Transpiration cooling

333-334

Trending and prognosis

510

Tubular/side combustors

305-306

Turbine blade cooling

332-350

convection

333-338

impingement

333-335

film

524-526

333

transpiration

333-334

steam

333-334

cooling design

334-338

strut insert design

334-338

cooled-turbine aerodynamics

338-339

hot gas expanders

339-343

336-338

This page has been reformatted by Knovel to provide easier navigation.

590-591

Index Terms

Links

Turbine blade cooling (Cont.) blade materials

342-346

blade coatings

347-349

instrumentation and controls

349-350

Turbine blades

332-350

cooling

332-350

high-pressure turbine

382-383

intermedia tepressure turbine

382-383

nozzles

384-385

low-pressure turbine

385-387

blade material characteristics

387-389

blade materials

387-392

Turbine efficiency

392

Turbine expander section

322-325

Turbine nozzles

384-385

Turbine rotor Turbine stages in series Turbine-compressor matching Turbomachinery aerodynamics

7 367

372-373

350-356 35-68

thermodynamics

35

aero-thermodynamics

35

dimensional analysis

36-41

nozzles and diffusers

42-45

turbomachinery

45-46

equation of state

46

ideal gas

46-48

compressibility effect

48-49

energy equation

382-392

50

continuity equation

50-53

momentum equation

53-56

This page has been reformatted by Knovel to provide easier navigation.

Index Terms

Links

Turbomachinery aerodynamics (Cont.) degree of reaction

56-57

efficiencies

56-58

Turbomachinery

4

aerodynamics foundation repair/rehabilitation

35-68

569-641

414-416

424

35-68 631

maintenance

569-641

Unbalances (equipment)

108-109

U

Unburnt hydrocarbons Update training

307 581-583

V Vane loading

146

Vane number

146

Vane thickness

146-147

Vaned diffuser

9-10

124-126

184-186

191-198

221

244-247 189-191

loss (stator)

221

Vaneless diffuser

125

183-184

220

243-244

loss (stator) Variable speed antisurge control

220 259-265

flow/delta P system

259-261

discharge volume flow

260-264

Velocity distribution

130-131

264-265 164-165

Velocity gradient in meridional plane

168-175

This page has been reformatted by Knovel to provide easier navigation.

178-179

Index Terms

Links

Velocity transducers

524-525

Velocity-pressure stage combination

374-376

Curtis-type

375

Rateau type

376

Vibration

103

119-120

510

524-26

103

119-120

524

analysis

119-120

510

524-526

Volume flow

206

measurement

Volutes

201-203

543-544

W Wake mixing loss (stator)

218-219

What if analysis

510-511

Whirl instability

426-433

hysteretic

426-428

dry friction

428-429

oil

429-430

aerodynamic

430-431

fluid trapped

432-433

Workshops

592

Wound-rotor induction motors

403

cost

403

This page has been reformatted by Knovel to provide easier navigation.

424-440