"Compact Heat Exchangers in Clean Energy ... - Wiley Online Library

Compact Heat Exchangers in Clean Energy Systems Shan-Tung Tu and Guo-Yan Zhou East China University of Science and Technology, Shanghai, P. R. China

1 INTRODUCTION With the sustained economic growth, the dependence on oil and electricity has made energy a vital component of our everyday needs. The hike in oil and gas prices has promoted everyone to take a careful look at the issues dealing with our energy supply and demand. Reducing the use of fossil fuels would considerably reduce the amount of carbon dioxide (Ram and Gupta, 2010) and other pollutants produced (Goddard Institute for Space Studies, 2009). This can be achieved by either using less energy altogether or replacing fossil fuels with renewable fuels. Hence, the advancing technologies emitting less carbon (e.g., high efficiency combustion) or no carbon are developing rapidly. Simultaneously, energy transformation and utilization efficiency are improved and carbon dioxide emitted during fossil fuel combustion is sequestered using higher efficient heat exchangers. Two examples of using compact heat exchangers (CHEs) are discussed in the following sections.

1.1

Aero engine systems

The air traffic is expanding remarkably over the past several decades. The growing tendency is expected to continue into the coming decades. During this period, the environment impact of air traffic such as gaseous and noise emission problems will therefore gain increasing importance. The government and related authorities will further tighten emission and noise regulations for air traffic vehicles, and these regulations are directly related with the high cost of fuel consumption. Environmental issues and the high cost of fuel consumption require gas turbine manufacturers to produce Handbook of Clean Energy Systems, Online © 2015 John Wiley & Sons, Ltd. This article is © 2015 John Wiley & Sons, Ltd. This article was published in the Handbook of Clean Energy Systems in 2015 by John Wiley & Sons, Ltd. DOI: 10.1002/9781118991978.hces119

environment friendly aero engines with lower emissions and higher specific fuel consumption ratings. The requirements can be met if highly compact and efficient heat exchangers are incorporated into aero-engines (Boggia and Rüd, 2005; McDonald et al., 2008). Hence, the concept of intercooled and recuperative aero engine (IRAE) has been forward, as shown in Figure 1 (Boggia and Rüd, 2005). Intercoolers and recuperators are currently used in land-based and marine propulsion gas turbines. Owing to the added weight, system complexity, and uncertainty in terms of structure, they are not incorporated in civil aero engines thus far (Min et al., 2009). In recent years, the ever-increasing environmental protectionism and the difficulty in increasing the cycle efficiency more made the research of intercoolers and recuperators for aero engine focused. Funded by the NEWAC (Wilfert et al., 2007) (new aero engine concepts) and CLEAN (Wilfert et al., 2005) (component validator for an environmentally friendly aero engine) programs, civil aero engine industries have also been working on the IRAE concept. The major components of IRAE are fan, gear, low pressure compressor, intercooler, high pressure compressor, high pressure turbine, intermediate pressure turbine, low pressure turbine, and recuperator. The hot air from the low pressure compressor is transferred to the intercooler and cooled by the cold air. The volume flow rate decreases at the intercooler and then goes into the high pressure compressor. This process reduces the work required to compress air between low and high pressure compressors because a compressor work is proportional to the rate of volume flow. Hence, they increase the net work output (Ito and Nagasaki, 2011). Meanwhile, the air from the high pressure compressor is transferred through a long duck along the outer side of the engine core into the recuperator that mounted at the end of the engine core. The air is heated by the exhaust hot air and then goes back into the combustor chamber that mounted before the high pressure turbine. The recuperating process decreases the fuel

2

Energy Efficiency Improvement

1.2 Intercooler Fan I-C HEX Gear

Nuclear power systems

Turbine Recuperator

Compressor

Figure 1. The concept of intercooled and recuperative aeroengine. Source: Reproduced with permission from Boggia and Rüd, 2005. © AIAA – American Institute of Aeronautics and Astronautics, Inc.

consumption because the fuel needed is less for the higher temperature of the combustor inlet air. Therefore, the overall fuel consumption reduces. Considering the harsh operating conditions such as high temperature and high pressure, the intercoolers and recuperators must have the characteristics of compact and ultra light, high effectiveness, minimum pressure loss to main performance benefit, very high pressure and temperature capability, structural integrity to cope with large temperature difference, and low cost. Based on these requirements, MTU developed a new type of recuperator with profile tube heat exchanger, as shown in Figure 2 (Boggia and Rüd, 2005; Schönenborn et al., 2006). Through the intercooled-recuperated structure, the targets include a 20% reduction in fuel consumption and CO2 emissions, an 80% cut in NOx , and halving the noise are expected to reach (Albanakis et al., 2009). However, the total weight of the intercooled-recuperated system was still estimated to be much more than 1000 kg per aero engine (Wilfert et al., 2007); this significantly adds the weight of the engine, so higher efficiency and ultra light of heat exchange system are needed for the future development of aero engine.

Since the first nuclear power plant was commercially operated in 1956, it has brought the electricity economically for the world. The industrially-advanced countries have developed sequentially several kinds of nuclear power plants, which use the heat produced by nuclear fission to generate steam that drives turbines. The high temperature gas-cooled reactor (HTGR) (Brey, 2001; Schleicher, Raffray, and Wong, 2000), with helium gas turbine cycle (Brayton cycle), can run for many months without interruption, providing reliable and predictable supplies of electricity. It is especially suitable for large-scale, continuous electricity demand that requires reliability (i.e., base-load), and hence ideally matched to increasing urbanization worldwide.1 In the HTGR system, reactor, turbine, high pass (HP) compressor, band pass (BP) compressor, and heat exchangers (recuperator, intercooler, and precooler) are the main components. The helium (He), usually at around 850◦ C and 7 MPa, is used as the core cooling medium and is introduced directly into a gas turbine for generating electricity. The basic process cycle is shown in Figure 3 (Pra et al., 2008). The heat of the helium gas from the power turbine is transferred by the recuperator to the helium gas from the high pressure compressor before returning to the reactor. The precooler and intercooler can reduce the work consumption by compressors to increase the cycle efficiency. The helium Brayton cycle with recuperator, precooler, and intercooler can achieve a high net thermal efficiency. Therefore, they, especially recuperator, are the key components to ensure a high net thermal efficiency of the HTGR gas turbinedirected cycle. A high heat exchange coefficient and small size should be designed to meet the work requirements.

Air in from compressor Damping wire Exhaust gas flow

Air out to combustor

Alternate rows of 3 / 4 profile tubes

Gas flow from turbine (a)

(b)

Figure 2. (a-b) Recuperator designed by MTU. Source: (a) Reproduced with permission from Boggia and Rüd, 2005. © AIAA – American Institute of Aeronautics and Astronautics, Inc. (b) Reproduced with permission from Schönenborn et al., 2006. © ASME. Handbook of Clean Energy Systems, Online © 2015 John Wiley & Sons, Ltd. This article is © 2015 John Wiley & Sons, Ltd. This article was published in the Handbook of Clean Energy Systems in 2015 by John Wiley & Sons, Ltd. DOI: 10.1002/9781118991978.hces119

Compact Heat Exchangers in Clean Energy Systems 3

490°C 7.07 MPa

HP circuit Reactor

BP circuit

510°C 2.64 MPa

Generator

MP circuit Turbine

850°C 7.01 MPa

105°C 7.22 MPa

Recuperator

125°C 2.59 MPa

HP compressor Intercooler

Water cooling

Precooler

Water cooling 26°C 2.57 MPa

BP compressor

Figure 3. High temperature gas-cooled reactor. Source: Reproduced from Pra et al., 2008. © Elsevier.

2 COMPACT HEAT EXCHANGERS CHEs are characterized by having a comparatively large area density. Area density is the ratio of heat transfer surface to heat exchanger volume. Their large area density, indicating small hydraulic diameter for fluid flow (Figure 4), results in a higher efficiency than conventional shell-and-tube heat exchanger in a significantly smaller volume. A compact heat exchanger has been arbitrarily defined by Shah (Shah, 1981; Mehendale, Jacobi, and Shah, 1999) that

20 Hydraulic diameter dn (mm)

The conventional tubular designs limited large recuperator effectiveness to 81–82%; the large plate-fin recuperators and other made possible designs of compact helium recuperators can get effectiveness up to 96% (Schleicher, Raffray, and Wong, 2000). The development of CHE will improve the performance of HTGR power conversion system. The above-mentioned and other clean energy systems, including wind energy, biomass energy, and tidal energy, require the use of heat exchange device having superior performance and reliable mechanical characteristics at high pressure and high temperature to guarantee the cycle efficiency. Moreover, geometric constraints are also important for such applications. Thus, CHE technologies are expected to be one of the solutions for this new generation clean energy system.

S+THX PFHE

10

Compact S+T brazed PHE, other PHEs

PCHE

Marbond HE Micro heat exchangers

5 0.5 0 0

100

1000

10,000 20,000

Area density (m2/m3)

Figure 4. Heat exchanger area density and hydraulic diameters. Source: Reproduced from Reay, 2002. © Elsevier.

having a surface area density greater than about 700 m2 /m3 or a hydraulic diameter Dh ≤ 6 mm if at least one fluid is gas, and in excess of 400 m2 /m3 when operating in liquid or multiphase streams. Some microscale heat exchangers under development, having an area density greater than about 15000 m2 /m3 or hydraulic diameter 1 μm ≤ Dh ≤ 100 μm, are more compact (Shah and Sekulic, 2003; Shah, 1991). The efficiency is necessarily improved by various heat transfer enhancement techniques, which have brought in a

Handbook of Clean Energy Systems, Online © 2015 John Wiley & Sons, Ltd. This article is © 2015 John Wiley & Sons, Ltd. This article was published in the Handbook of Clean Energy Systems in 2015 by John Wiley & Sons, Ltd. DOI: 10.1002/9781118991978.hces119

4


variety of compact heat exchangers. Some types of CHEs have been in routine use for many decades. Others have recently been introduced into the market or are still being tested in the lab. The necessary reduction of energy consumption, capital investment minimization, and improvement of adaptability of components has made the CHEs widely used in industry especially as gas-to-gas or liquid-to-gas heat exchangers; some examples are vehicular heat exchangers, condensers, and evaporators in air-conditioning and refrigeration industry, aircraft oil coolers, automotive radiators oil coolers, unit air heaters, intercoolers of compressors, and aircraft and space applications. CHEs are also used in cryogenics process, electronics, energy recovery, conservation and conversion, and other industries.

2.1

Plate heat exchangers (PHE)

The plate heat exchanger (PHE) was one of the first compact exchangers to be used in the United Kingdom process industries, being originally introduced in 1923 for milk pasteurization; the first plates were made of gunmetal. It is widely used in the food and drink processing industries, and is selectively used in the chemical processing industries. It is currently second to the shell and tube heat exchanger in terms of market share. The most common variant of the plate heat exchanger consists of a number of pressed, corrugated metal plates compressed together into a frame. These plates are provided with gaskets, partly to seal the spaces between adjacent plates and partly to distribute the media between the flow channels. Stainless steel (SS) is a commonly used metal for the plates because of its ability to withstand high temperatures, its strength, and its corrosion resistance.

2.1.1

Construction

As shown in Figure 5, the PHE is basically a series of individual plates pressed between two heavy end covers. These plates are gasketed, welded, or brazed together depending on the application of the heat exchanger (Kakac and Liu, 2002). The basic geometry of plates used in PHE is shown in Figure 6. The entire assembly of 10–100 plates is held together by the tie bolts, which gives 5–50 channels per fluid (Jogi and Lawankar, 2012; Thonon and Breuil, 2001). Individual plates are hung from the top carrying bar and are guided by the bottom carrying bar. For single-pass circuiting, hot and cold side fluid connections are usually located on the fixed end cover. Multi-pass circuiting results in fluid connections on both fixed and moveable end covers. The heat transfer surface consists of a number of thin corrugated plates pressed out of a high grade metal. The plates are pressed to form troughs at right angles to the direction of flow of the liquid that runs through the channels in the heat exchanger. These troughs are arranged so that they interlink with the other plates, which forms the channel with gaps of 1.3–1.5 mm between the plates. The hydraulic diameter is between 2 and 10 mm for most plates. The pressed pattern on each plate surface induces turbulence and minimizes stagnant areas and fouling. Unlike shell and tube heat exchangers, the plates for plate heat exchangers are mass-produced using expensive dies and presses. Therefore, all PHEs are made with what may appear to be a limited range of plate designs. Although the PHEs are made from standard parts, each one is custom designed. Variation in the trough angle, flow path, or flow gap can alter the NTU (number of transfer unit) of the heat exchanger. When the trough angle is 90◦ , the

Top bar

Follower Gasket

Flow plate Conector grid

End support

Bottom bar

Fixed head

Nut Washer Tie bolts

Figure 5. Basic structure of plate heat exchanger (PHE). Source: Reproduced with permission from Kakac and Liu, 2002. © Taylor and Francis. Handbook of Clean Energy Systems, Online © 2015 John Wiley & Sons, Ltd. This article is © 2015 John Wiley & Sons, Ltd. This article was published in the Handbook of Clean Energy Systems in 2015 by John Wiley & Sons, Ltd. DOI: 10.1002/9781118991978.hces119


Dp

Developed dimension

β

Protracted dimension

Lp

Lv t b

Lw

p

Pc Corrugation pitch Pc

Lh

Figure 6. Basic geometry of chevron plate. Source: Reproduced with permission from Kakac and Liu, 2002. © Taylor and Francis.

troughs run vertically. The flow passage made of such plates would resemble a collection of vertical tubes with low NTU characteristics. As the trough angle is reduced from 90◦ , the path becomes more tortuous and offers greater hydrodynamic resistance giving rise to high NTU characteristics. A combination of different plates may be used to create an intermediate NTU passage, which can be used to meet a specific NTU requirement. The plate pack is clamped together in a frame suspended from a carrying bar. Gaskets are fitted to seal the plate channels and interfaces. The frame consists of a fixed frame plate at one end and a moveable pressure plate at the other. The moveable plate facilitates access for cleaning or exchanging the heat transfer surfaces. A feature of this type of heat exchanger is the ability to add or remove surface area as necessary. The plates are group into passes with each fluid being directed evenly between the paralleled passages in each pass. Whenever the thermal duty permits, it is desirable to use single pass, counter flow for an extremely efficient performance. Although PHE can accept more than two streams, this is unusual. Two-pass arrangements are, however, common.

2.1.2

Characteristics

The area density of PHEs ranges from 120 to 660 m2 /m3 (Shah and Sekulic, 2003). PHEs are more compact than other types of heat exchangers. This means that they can exchange a greater amount of heat per a particular volume

of heat exchanger. A special feature of PHEs is that the ports for the hot and cold streams are incorporated in the plate form, thus obviating the need for header arrangements. They are cheaper to manufacture than other types of heat exchanger because their internal geometry is relatively simple. Using blanking plates within the stack, multi-passing can be accommodated, allowing for increased flow length and hence reduced temperature approach. Simultaneously, they are easier to dissemble and hence easier to clean and maintain. On the other hand, PHEs tend to leak more than other types of heat exchanger. This is because they tend to be made of cheaper materials and there is a greater length of joins that can potentially break. They also block more easily than other types of heat exchanger, because of the narrow gaps between the plates. This can be a problem if oxidants build up in central heating systems. They also generate a higher pressure drop than comparable heat exchangers, and so the high energy consumption of pumps is also a disadvantage (Wang, Sundén, and Manglik, 2007).

2.1.3

Operating limits

The operating limits of PHEs vary slightly from manufacturer to manufacturer. A summary of typical operating ranges for all different types of PHEs is given in Table 1 (Li et al., 2011). Typically, the operating temperature range of the gasketed metal PHE is from −35◦ C to +200◦ C. Design pressures up to 25 bar can be tolerated, with test pressures to



Washboard Wide-Gap Herringbone Zigzag

Herringbone

Herringbone Zigzag

Herringbone Dimpled Herringbone Dimpled Hybrid Herringbone

Wide-gap

Double-wall

Brazed

Semi-welded

AlfaRex

Maxchanger Compabloc

Hybrid

Packinox

b

Maximum differential temperature. Maximum differential pressure.

Herringbone Zigzag

Gasketed conventional

a

Plate Patterns

Type

Table 1. Typical operating ranges for PHEs.

550

−200 to 900

−195 to 540 (538)a −195.5 to 350 (170)a

−50 to 350

−45 to 220

−195 to 225

−35 to 200

−35 to 200

Rubber: −35 to 200 Graphite: −20 to 250

Operating Temperature (◦ C)

— —

120(40)b

15 4000

700

970

160

200

2000

0–5768

Flow rate (m3 /h)

80

115 45

40

40

45

16

16

35

Maximum Pressure (bar)

1000–20000/unit

6–8000/unit

250/unit 2.056–8.4/plate 4/unit 0.7–840/unit 0.061–0.989/unit

2500/unit 0.16–1.82/plate

75/unit

400/unit

1472.5/unit 0.28–1.56/plate

0.1–3800/unit 0.02–5/plate

Heat Transfer Area (m2 )

Alfa Laval

GEA-Ecoweld-Flex APV-Hybrid

Tranter Tranter-Ultramax GEA-Ecoweld-Bloc Alfa Laval-Compabloc

Alfa Laval

APV-Paraweld Alfa Laval Tranter GEA-EcoFlex-LWC

APV Alfa Laval GEA-EcoBraze

Alfa Laval Tranter-GD GEA-Safetytherm APV-Duo Safety

GEA-Ecoflex-Free Flow Tranter-Superchanger-GF APV-Easy flow

GEA-Ecoflex APV-ParaFlow Tranter-Superchanger-GC& GL

Main Products in the Market

6 Energy Efficiency Improvement

Compact Heat Exchangers in Clean Energy Systems 7 40 bar. Heat transfer areas range from 0.02 to 4.45 m2 (per plate). Flow rates of up to 3500 m3 /h can be accommodated in standard units, rising to 5000 m3 /h with a double port entry. Approach temperatures as low as 1◦ C are feasible with PHEs. The surface pattern on the plates tends to induce good mixing and turbulence, and in general this type of heat exchanger has a low propensity for fouling. Fouling resistances of typically 25% of those for shell and tube heat exchangers have been measured by the Heat Transfer Research Incorporated (HTRI) in the United States. Where fouling is a concern, the gap between the plates can be widened. For example, one manufacturer offers plates with a 13 mm gap and coarse contours for viscous liquids and fluids containing fibers, solids, crystals, pulp, and so on.

2.2

Plate-fin heat exchangers (PFHEs)

Originally developed for the aircraft industry in the 1940s, for use in environmental control and oil-fuel heat exchange duties, the aluminum plate-fin heat exchanger (PFHE) is extensively used in the cryogenics, or gas separation and liquefaction, industries, where the good low temperature properties of aluminum are paramount. Widespread use is also made in ethylene production plant. In recent decades, high temperature applications of PFHEs have attracted more and more attentions owing to their high compactness and efficiency. PFHEs are a matrix of flat plates and corrugated fins in a sandwich construction. Tube plates (i.e., parting sheets) provide the primary heat transfer surface. Tube plates are positioned alternatively with the layers of fins in the stack to form the containment between individual layers. These elements are built into a complete core and then vacuum brazed to form an integral unit. A section through a typical PFHE core is shown in Figure 7.

2.2.1

The vacuum-brazed construction technique also allows for multi-streaming, of up to 12 streams, to be incorporated into a single core, saving much weight and cost of the system. The high area density (hydraulic diameter of the order of 1–2 mm) also allows for the low temperature differences necessary for efficiency operation, especially at cryogenic temperatures, at which the power requirements are strongly influenced by Second Law constraints. The various fin types used are illustrated in Figure 8. Design variations in the configuration of the heat exchanger matrix can accommodate an almost unlimited range of flow options, including counterflow, crossflow, parallelflow, multi-pass, and multi-stream formats. The heat transfer fins provide the secondary heating surface for heat transfer. Fin types, densities, and heights can be varied to ensure that exchangers are tailor-made to meet individual customer requirements in terms of heat transfer performance versus pressure drop.

2.2.2

Characteristics

Brazed PFHEs exhibit certain features and characteristics that distinguish them from other types of heat exchangers. These include (Energy Efficiency Office, 2000): 1.

2.

3.

4.

A very large heat transfer area per unit volume of heat exchanger. This surface area is composed of primary and secondary (finned) surfaces. Typically, the effective surface area is over five times greater than that of a conventional shell and tube heat exchanger. Area densities range from 850 to 1500 m2 /m3 . A single heat exchanger can incorporate several different process streams and the unique plate-fin construction allows these to enter/exit the exchanger at intermediate points along the exchanger length rather than just at the ends. Very close temperature approaches between streams (typically 1–3◦ C) can be accommodated leading to operational cost savings. High thermal efficiency, use of aluminum and multistream capability combine to form a compact, low weight structure.

2.2.3

Figure 7. Basic structure of plate-fin heat exchanger (PFHE).

Construction

Operating limits

The versatility of PFHEs, coupled with the ability to manufacture them in a variety of other materials, makes them ideal for a range of process duties outside the cryogenics field.


8


(a)

Plate fin

(b)

Serrated fin

(c)

Perforated fin

(d)

Louvered fin

Figure 8. (a–d) PFHE fin types.

The maximum operating temperature of a PFHE is a function of its construction materials. Aluminum-brazed PFHEs can be used from cryogenic temperatures (−270◦ C) up to 200◦ C, depending on the pipe and header alloys. SS (stainless steel) PFHEs are able to operate at up to 650◦ C, whereas titanium units can tolerate temperatures approaching 550◦ C. The titanium PFHE can be designed for pressures in excess of 200 bar and at temperatures up to 400◦ C. Aluminum-brazed units can operate at up to 120 bar, depending on the physical size and the maximum operating temperature. SS PFHEs are currently limited to 50 bar, with developments expected that will extend the capability to 90 bar. Higher pressures can be tolerated using a diffusionbonded structure (Energy Efficiency office, 2000).

2.3

2 m, as can the exchanger diameter, giving heat transfer areas up to 600 m2 . Gasketed flat covers are fitted to the open side of each channel resulting in easy access and reduced maintenance costs. SPHEs tend to be self-cleaning. The smooth and curved channels result in a lower fouling tendency with difficult fluids. Each fluid has only one channel and any localized fouling will result in a reduction in the channel crosssectional area causing a velocity increase to scour the fouling layer. This self-cleaning effect results in reduced operating costs particularly when the unit is horizontally mounted. Horizontal mounting is essential when handling fibrous, high viscosity, particle-laden, or clogging media as all particles potentially settle to the bottom of the channel curvature.

Spiral plate heat exchangers (SPHEs)

Spiral plate heat exchanger (SPHE) design approaches the ideal in heat transfer equipment by obtaining identical flow characteristics for both media. The classic design of a SPHE is simple; the basic spiral element is constructed of two metal strips rolled around a central core forming two concentric spiral channels. Normally, these channels are alternately welded, ensuring that the hot and cold fluids cannot intermix. The SPHE can be optimized for the process concerned using different channel widths. Channel width is normally in the range 5–30 mm (representing hydraulic diameters of 10–60 mm). Plate width along the exchanger axis may be

2.3.1

Construction

The SPHEs operate in nearly complete counterflow and are assembled from two long strips of plate wrapped to form a pair of concentric spirals, as visible in the schematic in Figure 5a. Alternate edges of the passages are closed so that the fluid streams flow through continuous sealed channels. Studs are normally welded onto one side of each strip for support. Covers are fitted to each end to complete the unit (Kakac and Liu, 2002; Energy Efficiency offices, 2000). It is clear from the single channel spiral form that the typical NTU



(a) Spiral flow-spiral flow

(b)

Cross flow-spiral flow

(c)

Combination flow

Figure 9. (a–c) Basic structure of spiral plate heat exchanger (SPHE). Source: Reproduced with permission from Kakac and Liu, 2002. © Taylor and Francis.

is high, implying “long” thermal duties. They can be tailormade to perform in a wide variety of duties in all metals that can be cold-formed and welded, such as carbon steel, SS and titanium. High grade alloys are routinely used for excellent resistance to corrosion and erosion. In some cases, double spacing may be used, produced by simultaneously winding four strips to form two channels for each fluid. These double channel systems are used when there is a large flow rate or small pressure drop, but should not be used for fouling media or media containing solids. The use of SPHEs is not limited to liquid–liquid services. Variations to the basic design give exchangers that are suitable for liquid–vapor or liquid–gas services. Typically, SPHEs are available in three configurations (as shown in Figure 9). For type I, the hot fluid enters at the center of the unit and flows from the inside outward. The cold fluid enters at the periphery and flows toward the center. Two media are in full counter-current flow. For type II, the medium in crossflow passes through the open channels of the spiral usually in a vertical direction. The service fluid spiral flows through the other channel, welded shut, with side wall inlet and central outlet fed through the side wall as shown in Figure 9b. This design can be used as either a condenser or a vaporizer. However, for Type III, a gas or vapor mixture to liquid exchanger combines the above two designs; the hot stream enters at the top and flows tangentially through the exchanger exiting at the side, as shown in Figure 9c.

2.3.2

Characteristics

Spiral designs have a number of advantages compared to shell and tube heat exchangers: 1. 2. 3. 4.

Optimum flow conditions on both sides of the exchanger. An even velocity distribution, with no dead-spots. An even temperature distribution, with no hot or coldspots. More thermally efficient with higher heat transfer coefficients.

5.

6. 7.

Copes with exit temperature overlap, or crossover, whereas shell and tube units require multi-shells in series to handle temperature crossover. Small hold up times and volumes. Removal of one cover exposes the total surface area of one channel providing easy inspection cleaning and maintenance.

2.3.3

Operating limits

Typically, the maximum design temperature is 400◦ C set by the limits of the gasket material. Special designs without gaskets can operate with temperatures up to 850◦ C. Maximum design pressure is usually 15 bar, with pressures up to 30 bar attainable with special designs.

2.4 2.4.1

Heat pipe heat exchangers (HPHEs) Construction

A heat pipe heat exchanger (HPHE) is similar to a tube-fin exchanger with individually finned tubes or flat (continuous) fins and tubes (Chaudourne, 1992; Vasiliev, 2005; Reay, Kew, and McGlen, 2014). It is a special type of heat exchanger using the excellent thermal properties of the heat pipes. A heat pipe is a tube containing a liquid in equilibrium with its vapor (the working fluid). It does not contain air or any other gas, and it is hermetically closed. The inner surfaces of a heat pipe are usually lined with a capillary wick (a porous lining, screen, or internally grooved wall). The wick is what makes the heat pipe unique and forces condensate to return to the evaporator by the action of capillary force. In a properly designed heat pipe, the wick is saturated with the liquid phase of the working fluid, whereas the remainder of the tube contains the vapor phase. When heat is applied at the evaporator by an external source, the working fluid in the wick in that section vaporizes, the pressure increases, and vapor


10


flows to the condenser section through the central portion of the tube. The vapor condenses in the condenser section of the pipe, releasing the energy of phase change to a heat sink (to a cold fluid, flowing outside the heat pipe). The heat applied at the evaporator section tries to dry the wick surface through evaporation, but as the fluid evaporates, the liquid–vapor interface recedes into the wick surface, causing a capillary pressure to be developed. This pressure is responsible for transporting the condensed liquid back to the evaporator section, thereby completing a cycle (Figure 10). Thus, a properly designed heat pipe can transport the energy of phase change continuously from the evaporator to the condenser without drying out the wick. The condensed liquid may also be pumped back to the evaporator section by the capillary force or by the force of gravity if the heat pipe is inclined and the condensation section is above the evaporator section. If the gravity force is sufficient, no wick may be necessary. As long as there is a temperature difference between the hot and cold gases in an HPHE, the closed-loop evaporationcondensation cycle will be continuous, and the heat pipe will continue functioning. Generally, there is a small temperature difference between the evaporator and condenser section [about 5◦ C or so], and hence the overall thermal resistance of a heat pipe in a heat pipe exchanger is small. Although water is a common heat pipe fluid, other fluids are also used, depending on the operating temperature range. An HPHE, as shown in Figure 11 for a gas-to-gas application, consists of a number of finned heat pipes (similar to an air-cooled condenser coil) mounted in a frame and used in a duct assembly. Fins on the heat pipe increase the surface area to compensate for low heat transfer coefficients with gas flows. The fins can be spirally wrapped around each pipe, or a number of pipes can be expanded into flat plain or augmented fins. The fin density can be varied from side to side, or the pipe may contain no fins at all (liquid applications). The tube bundle may be horizontal or vertical with the evaporator sections below the condenser sections. The tube rows are normally staggered with the number of tube rows typically between 4 and 10. Unit size varies with airflow. Small units have a face size of 0.6 m (length) by 0.3 m (height), and the largest units may have a face size up to 5 m × 3 m. Vessel wall

Vapour flow Evaporation

Heat in

Condensation

Liquid return in wick

Figure 10. Principle of a heat pipe.

Heat out

Spliter plate

Heat transfer

Cold gas with recovered heat Hot gas

Finned heat pipes

Figure 11. Heat pipe heat exchanger (HPHE). Source: Reproduced from Reay et al., 2014. © Elsevier.

2.4.2

Characteristics

The HPHE has several advantages over the classical heat exchangers. The most attractive features of these heat exchangers are the following: 1. 2.

3. 4. 5.

great design flexibility: heat pipes are independent components; high thermal effectiveness: heat pipes have a very high conductance, and it is easy to increase the exchange surface between these heat pipes and the fluid when using finned tubes for heat pipes for example; excellent mechanical isolation between the two fluids: the heat pipes make a double wall between the two fluids; small pressure loss on the two fluids: they flow only along the outside of the heat pipe tubes; possibility to adjust the heat exchange surface temperature according to the choice of the heat exchange areas on the hot side and the cold side (this is useful to avoid corrosion due to acid condensation). This characteristic results from the independence of the hot and cold exchange surfaces.

2.4.3

Operating limits

HPHEs are generally used in gas-to-gas heat transfer applications. They are used primarily in many industrial and


Compact Heat Exchangers in Clean Energy Systems 11 consumer product-oriented waste heat recovery applications. The maximum operating temperature of an HPHE is a function of its construction materials and working fluid. HPHEs can be used from cryogenic temperatures (−60◦ C) up to 1100◦ C. It can operate at up to 120 bar, depending on the physical size and the maximum operating temperature.

described fluid flow rates, inlet temperatures, and allowable pressure drop for an existing heat exchanger. The sizing problem, on the other hand, involves determination of the dimensions of the heat exchanger, that is, selecting an appropriate heat exchanger type and determining the size to meet the requirements of specified inlet and outlet temperatures, flow rates, and pressure drops. The flow char of heat exchanger design methodology is given in Figure 12.

3 DESIGN OF COMPACT HEAT EXCHANGERS

3.1

The tasks in heat exchanger design are rating and sizing. The rating problem is concerned with the determination of the heat transfer rate and the fluid outlet temperatures for

In order to develop relationships among the heat transfer rate, surface area A, fluid terminal temperatures, and flow rates in a heat exchanger, the basic equations used for analysis are the

Overall heat transfer equation

Problem specification

Operating conditions

HE material

HE flow arrangement

Thermo-physical properties of fluids and materials

HE construction

Surface characteristics and geometrical properties

Preliminary thermal design (heat transfer and pressure drop analysis and optimization)

Thermal model

Fluid-dynamic model

Iteration

Mechanical design (including headers, thermal stresses, vibrations and fouling)

Optimized optional solutions (appropriate HE type, dimensions, heat transfer area, pressure drop, etc.)

Evaluation criteria

Evaluation procedure and costing

Optimal solution for the case considered

Figure 12. Heat exchanger design procedure. Handbook of Clean Energy Systems, Online © 2015 John Wiley & Sons, Ltd. This article is © 2015 John Wiley & Sons, Ltd. This article was published in the Handbook of Clean Energy Systems in 2015 by John Wiley & Sons, Ltd. DOI: 10.1002/9781118991978.hces119

Tradeoff factors

12


energy conservation and heat transfer rate equations (Shah and Mueller, 1985). The overall energy balance for any twofluid heat exchanger is given by q = mh cp,h (Th1 − Th2 ) = mc cp,c (Tc2 − Tc1 )

(1)

Tw =

Δtm Ro

Tw = (2)

where Δtm is the true mean temperature difference, which depends on the exchanger flow arrangement and the degree of fluid mixing within each fluid stream mh cp,h = Ch is the capacity rate of the hot fluid mc cp,c = Cc is the capacity rate of the cold fluid Th1 and Th2 are hot fluid terminal temperatures (inlet and outlet) Tc1 and Tc2 are cold fluid terminal temperatures (inlet and outlet).

3.2

Th ∕Rh + Tc ∕Rc (𝜂 hA) T + (𝜂o hA)c Tc = o h h 1∕Rh + 1∕Rc (𝜂o hA)h + (𝜂o hA)c

Ro = Rh + R1 + Rw + R2 + Rc

(3)

where Rh is hot side film convection resistance, 1/(𝜂 o hA)h ; R1 thermal resistance due to fouling on the hot side, Rf,h /(𝜂 o A)h ; Rw thermal resistance of the separating wall, 𝛿/(Aw kw ); R2 the thermal resistance due to fouling on the cold side, Rf,c /(𝜂 o A)c ; and Rc the cold side film convection resistance, 1/(𝜂 o hA)c . Then, the equation can be alternately expressed as 1 1 1 1 1 = + + Rw + + KA (𝜂o hA)h (𝜂o A)h (𝜂o hA)c (𝜂o A)c

(4)

As KA = Kh Ah = Kc Ac , the overall heat transfer coefficient K may be defined optionally in term as of either hot fluid surface area or cold fluid surface area. Thus, the option of Ah or Ac must be specified in evaluating K from the product, KA. The knowledge of wall temperature in a heat exchanger is essential to determine the localized hot spots, freeze points, thermal stresses, local fouling characteristics, or boiling and condensing coefficients. On the basis of thermal circuit of figure, when Rw is negligible, Tw,h = Tw,c = Tw is computed from Shah and Mueller (1985) and Shah (1983) as

(6)

Design methodology

The rating and sizing of heat exchangers are two important problems encountered in the thermal analysis of compact heat exchangers. LMTD and 𝜀-NTU methods have been used to perform a heat exchanger thermal analysis.

3.2.1

The 𝜀-NTU

In the 𝜀-NTU method, the heat transfer rate from the hot fluid to the cold fluid in the exchanger is expressed as q = 𝜀Cmin (Th1 − Tc1 ) = 𝜀Cmin ΔTmax

The inverse of the overall thermal conductance KA is referred to as the overall thermal resistance Ro , and it is made up of component resistances in series

(5)

When R1 = R2 = 0, the above equation further simplifies to

and the heat transfer rate equation is q = KAΔtm =

Th + Tc [(Rh + R1 )∕(Rc + R2 )] 1 + [(Rh + R1 )∕(Rc + R2 )]

(7)

where 𝜀 is the heat exchanger effectiveness and Cmin the minimum of Ch and Cc . The heat effectiveness 𝜀 is a measure of thermal performance of a heat exchanger. It is defined for a given heat exchanger of any flow arrangement as a ratio of the actual heat transfer rate q from the hot fluid to the cold fluid to the maximum possible heat transfer rate qmax thermodynamically permitted: q (8) 𝜀= qmax Consider a counterflow heat exchanger having infinite surface area. An overall energy balance for the two fluid streams is q = Ch (Th1 − Th2 ) = Cc (Tc2 − Tc1 )

(9)

For an infinite area counterflow exchanger, we get qmax as qmax = Cmin (Th1 − Tc2 ) = Cmin ΔTmax { ( ) Ch Th1 − Tc2 for Ch ≤ Cc = Cc (Th1 − Tc2 ) for Ch ≥ Cc

(10)

Thus, 𝜀 can be determined directly from the operating temperatures and heat capacity rates,



𝜀=

C (T − Tc1 ) Ch (Th1 − Th2 ) = c c2 Cmin (Th1 − Tc2 ) Cmin (Th1 − Tc2 )

(11)

potential for heat transfer that can only be obtained in a counterflow exchanger. Some limiting values of ΔTlm are as follows.

An alternative expression of 𝜀 using q from the overall rate equation and qmax from Equation (10) is ( 𝜀=𝜙

) KA Cmin , , flow arrangement Cmin Cmax

ΔTlm

= 𝜙(NTU, C∗ , flow arrangement) where Cmin Cmax

KA Cmin

= NTU is the number of transfer units and C∗ =

the heat capacity ratio. It is noted that 𝜀 is dependent on the NTUs, the heat capacity rate ratio C*, and the flow arrangement for a directtransfer type heat exchanger. The functional relationship 𝜙 is dependent on the flow arrangement. Similar expressions have been developed for heat exchangers having other flow arrangements such as parallelflow and crossflow. Some representative results are summarized in Table 2 (Shah and Sekulic, 2003; Thulukkanam, 2013).

3.2.2

Log-mean temperature difference (LMTD)

⎧ ΔT1 +ΔT2 ⎪ 2 = ⎨ΔT1 = ΔT2 ⎪0 ⎩

for ΔT1 = ΔT for ΔT1 or ΔT2 = 0 (NTU → ∞) (16) The LMTD ΔTlm normalized with respect to the inlet temperature difference ΔTmax = Th1 − Tc2 can be expressed in terms of the temperature effectiveness and the exchanger effectiveness: P1 − P2 ΔTlm = Th1 − Tc1 ln [(1 − P2 )∕(1 − P1 )] =

⎧ ΔTlm = 1 − 𝜀 ΔTlm ⎪ ΔTmax =⎨ Th1 − Tc1 ⎪ ΔTlm =0 ⎩ ΔTmax

(12)

where KA is the exchanger overall thermal conductance and ΔTlm the log-mean temperature difference (LMTD) defined as ΔT1 − ΔT2 (13) LMTD = ΔTlm = ln (ΔT1 ∕ΔT2 ) where ΔT1 and ΔT2 are temperature differences between two fluids at each end of a counterflow or parallelflow exchanger. For a counterflow exchanger, ΔT1 = Th1 − Tc2 , ΔT2 = Th2 − Tc1

(14)

For a parallelflow exchanger, ΔT1 = Th1 − Tc1 , ΔT2 = Th2 − Tc2

(15)

For all other flow arrangements, the heat exchanger is hypothetically considered a counterflow unit operating at the same R value and the same terminal temperatures (or the same effectiveness). Hence, LMTD for all other flow arrangements is evaluated from Equation (1) using ΔT1 and ΔT2 . Note that LMTD represent the maximum temperature

(1 − C∗ )𝜀 ln [(1 − C∗ 𝜀)∕(1 − 𝜀)]

(17)

This relationship is obtained directly from the definitions of ΔTlm , P1 , P2 , 𝜀, and C*, and hence is valid for all flow arrangements. Following are two limiting forms of the above equation.

The heat transfer rate in the heat exchanger is represented by Q = KAFΔTlm

for ΔT1 → ΔT

for C∗ → 1 (18) for 𝜀 → 1

F is referred to as the LMTD correction factor. It is dimensionless. In general, it is dependent on the temperature effectiveness P, the heat capacity rate ratio R, and the flow arrangement. { F=

( ) 𝜙1 P1 , R1 = 𝜙1 (P2 , R2 ) for a stream symmetric exchanger 𝜙1 (P1 , R1 ) = 𝜙2 (P2 , R2 )

for a stream asymmetric exchanger

The explicit relationships are shown in Table 3 (Shah and Sekulic, 2003; Thulukkanam, 2013). The LMTD method is traditionally used to solve sizing problems by the calculation of heat balances. It may also be used for rating problems (performance analysis) for an available heat exchanger. However, it would be tedious, requiring iteration as one of inlet or outlet temperatures are not known. The use of the 𝜀-NTU method is generally preferred in the design of compact heat exchangers where the inlet temperatures of the hot and cold fluids are specified and the heat transfer rates are to be determined. The analysis can be simplified by using 𝜀-NTU method.


∑ C∗ Pn (NTU)


yn+j

𝜀=

NTU 1+NTU

1 C∗ 1−exp(−NTU⋅C∗ )

𝜀 = 1 − exp(−NTU)

+

All heat exchangers with C* = 0

1 1−exp(−NTU)

−

1−exp{−C∗ [1−exp(−NTU)]} C∗

1 NTU

For Cmax mixed, Cmin unmixed, 𝜀 = 𝜀=

All heat exchangers with C* = 1

j!

n ∑ (n + 1 − j)

j=1 [ ] 1−exp(−NTU⋅C∗ ) For Cmin mixed, Cmax unmixed, 𝜀 = 1 − exp − C∗

1 (n+1)!

Both fluids mixed

One fluid mixed, other unmixed

Pn (y) =

n=1

𝜀 = 1 − exp(−NTU) − exp[−NTU(1 + C∗ )]

Both fluids unmixed

1−C∗ 𝜀 1−𝜀

[1−𝜀(1+C∗ )] 1+C∗

ln

for C∗ < 1NTU =

1 1−𝜀

for C∗ = 1

NTU =

𝜀 1−𝜀

NTU = − ln(1 − 𝜀)

/

For Cmin mixed, Cmax unmixed, NTU = − C1∗ ln [1 + C∗ ln (1 − 𝜀)] ] [ For Cmax mixed, Cmin unmixed, NTU = − ln 1 + C1∗ ln (1 − C∗ 𝜀)

/

—

—

Crossflow ∞

NTU = − ln

1 − exp[−NTU(1 + C∗ )] 1 + C∗

𝜀=

Parallel flow

1 1−C∗

NTU =

1 − exp[−NTU(1 − 1 − C∗ exp[−NTU(1 − C∗ )]

𝜀=

NTU

Counterflow

C∗ )]

𝜀

Flow Arrangements

Table 2. 𝜀-NTU formulas for known heat exchanger flow.

14 Energy Efficiency Improvement

Compact Heat Exchangers in Clean Energy Systems 15 Table 3. F as an explicit function of P1 and R1 only for the specific heat exchanger flow arrangements. Flow Arrangements

Formula

Counterflow

F=1

Parallelflow

F=1

Crossflow (single-pass)

—

One fluid unmixed, other fluid mixed, stream asymmetric

F=

One fluid mixed, other fluid unmixed, stream asymmetric

=

ln [(1−R2 P2 )∕(1−P2 )] (1−1∕R2 ) ln [1+R2 ln (1−P2 )]

F=

ln [(1−R1 P1 )∕(1−P1 )] (1−1∕R1 ) ln [1+R1 ln (1−P1 )]

= All exchangers with R1 = 0 or ∞

3.3 3.3.1

ln [(1−R1 P1 )∕(1−P1 )] (R1 −1) ln [1+(1∕R1 ) ln (1−R1 P1 )]

ln [(1−R2 P2 )∕(1−P2 )] (R2 −1) ln [1+(1∕R2 ) ln (1−R2 P2 )]

F=1

Thermal design of compact heat exchangers

where { 16

Plate heat exchangers

One of the most commonly used high performance PHE surfaces has chevron plates with the important geometrical parameters identified in Figure 13. A considerable amount of research has been conducted to determine heat transfer and flow friction characteristics of this geometry. The Nusselt number correlation for this geometry was obtained as follows (Martine, 1996), using the momentum and heat transfer analogy from a generalized Leveque solution in thermal entrance turbulent flow in a circular pipe (Schlunder, 1998). ( )1∕6 hDh 1∕3 𝜇m Nu = = 0.205Pr (f ⋅ Re2 sin 2𝛽)0.374 (19) k 𝜇w It is valid for the corrugation angle 𝛽 within 10–80◦ and is accurate within ±30%, and within ±13% for industrial plates. Note that, the viscosity correction term (𝜇m /𝜇 w )1/6 should be omitted if it is used for gases. A 100% error in f will translate to 30% error in Nu, owing to the exponent 0.374 on f in the above equation. The comprehensive correlations for friction factors were also provided as follows:

(1.56 ln Re − 3.0) { 149.25 Re 9.75 Re0.289

f1 =

1 − cos 𝛽 + √ 3.8f1

+ 0.9625

−2

for Re ≥ 2000

,

for Re < 2000 for Re ≥ 2000

It is valid for the corrugation angle 𝛽 within 0–80◦ and is accurate within −50% and +100%. If the model plate data are eliminated, the correlation is based on industrial plates of PHEs and is within ±40% accuracy. Of course, this correlation can be improved further if the actual detailed geometrical information would be available.

3.3.2

Plate-fin heat exchangers

The three basic fin surfaces, including plain rectangular, offset strip, and perforated fins, are by far the commonest of the PFHE, being used for applications from aerospace airconditioning duties to oil refining. These surfaces have one of the highest heat transfer performances relative to the friction factor. Extensive analytical, numerical, and experimental investigations have been conducted over the past 50 years. The most comprehensive correlations for j and f factors for the offset strip fin geometry are provided by Manglik and Bergles (1995) as follows. ( −0.5403

j = 0.6522Re

cos 𝛽 1 √ = (0.045 tan 𝛽 + 0.09 sin 𝛽 + f0 ∕ cos 𝛽)1∕2 f

for Re < 2000

Re

f0 =

[

sf hf

)−0.1541 ( )0.1499 ( )−0.0678 tf tf l sf (

−5

1.340

× 1 + 5.269 × 10 Re

(20)


sf hf

)0.504 ( )0.456 ( )−1.055 ]0.1 tf tf l sf (21)

16


25%

25%

20%

20%

15%

15%

10%

10%

5%

5% 0%

0%

SS

SS

E-

PH

FE

E-

E

E-

TH

S

(a)

p ra

g

H

PF

SS

h

T -P

H

PF

ite

E-

E-

E-

H

ST

SS

SS

SS

First level

PF

E

E-

TH

S

(b)

p ra

g

H

PF

te

SS

hi

T -P

H

PH

SH

FE

E-

H

E-

ST

SS

E-

SH

Second level

25% 20% 15% 10% 5% 0%

SS

SS FH

P

ite

E

TF

E-

E-

PH

a gr

H

PF

(c)

E-

H

ST

SS

ph

P E-

E-

TH

S

SS

E-

SH

Third level

Figure 13. (a–c) Weight allocation diagram of the available option at different level.

( f = 9.6243Re−0.7422 [

For perforated fins within 400 ≤ Re ≤ 10000,

) ( ) ( ) sf −0.1856 tf 0.3053 tf −0.2659 hf l sf (

−8

4.429

× 1 + 7.669 × 10 Re

sf hf

ln j = −9.544151 × 10−2 (ln Re)3 + 2.137607(ln Re)2

)0.920 ( )3.767 ( )0.236 ]0.1 tf tf l sf (22)

−15.92678(ln Re) + 34.57583

ln f = −6.736098 × 10−2 (ln Re)3 + 1.565191(ln Re)2 −12.31399(ln Re) + 28.79806

The comprehensive correlations for heat transfer and friction factors are provided by ALEX, Japan in 1965, as follows (Qian, 2008). For plain rectangular fins within 400 ≤ Re ≤ 10000, ln j = 0.103109(ln Re)2 − 1.91091(ln Re) + 3.211 (23) ln f = 0.106566(ln Re)2 − 2.12158(ln Re) + 5.82505 (24) For offset strip fins within 300 ≤ Re ≤ 7500, ln j = −2.64136 × 10−2 (ln Re)3 + 0.555843(ln Re)2 −4.09241(ln Re) + 6.21681

(25)

ln f = 0.132856(ln Re)2 − 2.28042(ln Re) + 6.79634 (26)

(27)

(28)

Note that the above correlations are only dependent on the type of fins and cannot be affected by the structural parameters. They have been proved accurate in practical engineering.

3.3.3

Spiral plate heat exchangers

The flow regime dictates the form of expressions used for the determination of pressure drop and for the estimation of the film heat transfer coefficient. The choice of the appropriate expression depends on the values of the Reynolds number and the critical Reynolds number given respectively by ( )0.32 Dh Dh M and Rec = 20, 000 Re = 𝜇Ac Ds


(29)

Compact Heat Exchangers in Clean Energy Systems 17 where Dh is the hydraulic diameter, M the mass flow rate, Ac the free flow area, and 𝜇 the viscosity. The following equations are provided by Minton (1970) for the determination of the heat transfer coefficient and the pressure drop. For Re > Rec , ( ) D h = 0.023Cp Vf Re−0.2 Pr2∕3 1 + 3.54 h (30) Ds ] ) [ 1.3𝜇1∕3 ( )1∕3 H 16 L M 2 + 1.5 + ΔP = 0.001 s bH L (b + 0.125) M (

(

L Dh

)−1∕3 (

𝜇w 𝜇b

(36)

For spiral-fined HPHE, the friction factor can be calculated by the following equation (Gunter and Shaw, 1977) ( ) ( ) ( ) do Gmax −0.316 ST −0.927 ST 0.515 f = 37.86 𝜇t do SL

(37)

where Gmax is maximum mass velocity of the fluid, 𝜇t the fluid viscosity, ST the transverse distance of two adjacent heat pipes, and SL the longitudinal distance of two adjacent heat pipes.

(32)

3.4

)−0.14

( ) L M ΔP = 0.001 s bH [ ] ( )0.17 ( )1∕2 1.035𝜇1∕2 𝜇w H 16 × + 1.5 + (33) M L (b + 0.125) 𝜇b where h is the heat transfer coefficient, 𝜇w the fluid viscosity at the wall temperature, 𝜇 b is the fluid bulk viscosity, L the plate length, Vf the fluid mean velocity, Cp the specific heat capacity of the fluid, Pr the Prandtl number, ΔP the fluid pressure drop, and s the relative density (relative to water at 20◦ C).

3.3.4

f = 1.92Re−0.145

(31) For 100 < Re < Rec , h = 1.86Cp Vf Re−2∕3 Pr−2∕3

The following equations are provided by experiments for the determination of the friction factor. For the circular-fined HPHE, the friction factor can be estimated using Equation (36) (Robison and Briggs, 1966).

Heat pipe heat exchanger

The Nusselt number correlation for this geometry was obtained as follows (Briggs and Young, 1962) ( )0.2 ( )0.1134 s s l 𝛿 for 0.125 < s∕l < 0.610 and 45 < s∕𝛿 < 80 (34)

Nu = 0.134Re0.681 Pr1∕3

where do is outer diameter of heat pipe, 𝜆 heat conductivity coefficient of fluid, s/l the ratio of fin spacing to fin height, and s/𝛿 the ratio of fin spacing to fin thickness. Some experiments were further made for the hot fluid operated at 240–380◦ C (Zhuang, Xu, and Shi, 1989), and the comprehensive Nusselt correlation equation is then obtained as follows. Nu = 0.137Re0.6338 Pr1∕3 for 6000 < Re < 14, 000 (35)

Optimization of compact heat exchangers

In order to design the high efficient compact heat exchanger, the optimization of compact heat exchangers with the different objective functions should be performed, such as minimization of costs, minimization of the number of entropy generation units (Yekoladio, Bello-Ochende, and Meyer, 2013), minimum total pressure loss (Joda, Tahouni, and Panjeshahi, 2013; Kotcioglu, Cansiz and Khalaji, 2013), minimum effectiveness (Sanaye and Hajabdollahi, 2013; Sepehr and Masoud, 2011), minimum total cost (Hadidi and Nazari, 2013; Sun, Alwi, and Manan, 2013), and minimum weight. As for mathematical method, in addition to using traditional mathematical methods (Reneaume, Pingaud, and Niclout, 2000) and artificial neural network (Peng and Ling, 2008), a genetic algorithm (GA) (Allen and Gosselin, 2008), particle swarm optimization (PSO) algorithm (Rao and Patel, 2010), global sensitivity analysis (GSA), and harmony search algorithm (HSA) (Fesanghary, Damangir, and Soleimani, 2009), and so on can be successfully applied for thermodynamic optimization of a compact heat exchanger. The weighting factor of different targets (Chung, Lee, and Kim, 2002) can also be introduced to optimize the geometry of heat exchangers using computer simulation (Jassim and Mohammed, 2003). In addition, in engineering design, technical-economic analysis method (Zhou, Wu, and Tu, 2008), simple additive weighting (SAW) (Chou, Chang, and Shen, 2008), technique for preference by similarity to the ideal solution (TOPSIS) (Cavallaro, 2010), analytical hierarchy process (AHP) (Naghadehi, Mikaeil, and Ataei, 2009), and nonstructural fuzzy decision-making method (NSFDMM) (Zhou, Wu, and Tu, 2014) are used to select the best alternative in the early stage of design process.


18


Table 4. Technological parameters of the available options. Item

Plate HE (PHE)

Material C1-Volume, V (m3 ) C2-Weight, m (kg) C3-Service temperature, T (K) C4-Heat transfer area, A (m2 ) C5-Pressure drop, ΔP (kPa) C6-Life expectancy (year) C7-Manufacture cost ($) C8-Maintenance cost ($) C9-Operating cost ($)

SS 0.23 315.54 523 29.19 12.28 4 9981 100 7184

Plate-Fin HE (PFHE) 316SS 0.38 513.96 923 154.95 0.50 6 16959 165 10,474

PTFE 0.55 178.97 673 140.24 0.60 8 8157 82 754

A case study is given to demonstrate the optimal selection of compact heat exchangers (Zhou, Wu, and Tu, 2014). For a given working conditions, there are three kinds of candidate materials, including SS, graphite, and newly developed polytetrafluoroethylene (PTFE) composite. The available types are PHE, PFHE made of 316SS, STHE (shell and tube heat exchanger), and SPHE. In order to promote energy efficiency and conservation, the minimum value of total pressure drop has been chose to be the ultimate objective for the optimum design. The final technological parameters of eight options are obtained using the above methods, as shown in Table 4. According to the NSFDMM method procedure, the volume (C1), weight (C2), service temperature (C3), heat transfer area (C4), pressure drop (C5), life expectancy (C6), manufacture cost (C7), maintenance cost (C8), and operating cost (C9) are selected as the decision criteria. The three levels fuzzy comprehensive evaluation results are obtained as shown in Figure 13. From the point of comprehensive performance, the PFHE made of PTFE composite is feasible and economically optimal.

4 CONCLUDING REMARKS Compact heat exchangers play an important role in energy saving in clean energy systems. In the coming years, increasing demand for heat exchangers complying with the principles of ecological and economics sustainability will certainly further expand their industrial applications. The success of the new generation of clean energy systems will depend in part on the correct selection of the compact heat exchanger technologies. On the other hand, the long-term and safe operation of the compact heat exchangers in clean energy systems is also a critical issue to enable the technology. The present article has been mainly concerned with the heat transfer and thermal-hydraulic performance of the compact heat exchangers. It is noted that few efforts have been made to study the reliability of the equipment. In

Shell and Tube HE (STHE) Graphite 12.08 3875.88 443 764.20 2.55 7 8774 82 1220

SS 8.28 14645.16 773 673.10 5.46 5 19,386 178 3606

Spiral Plate HE (SPHE) SS 1.18 932.48 573 41.25 84.2 5 27392 219 43543

recent years, the effect of residual stresses and the thermal distortion during the fabrication of the compact structures have been investigated (Jiang et al., 2008a, b). We extended the elastic-media theory to calculate the time-dependent deformation and stress of compact structures of the heat exchangers (Zhou and Tu, 2007; Zhou et al., 2011). Aiming at the creep design of the compact heat exchangers for high temperature applications, the creep behavior of the plate-fin structures and brazed joints of high temperature heat exchangers were studied (Tu and Zhou, 2009). With more and more compact heat exchangers used in the clean energy system, a balanced investigation will be required on both thermal performance and structural integrity.

ENDNOTES 1.

World Energy Needs and Nuclear Power. http:// www.world-nuclear.org/info/Current-and-FutureGeneration/World-Energy-Needs-and-Nuclear-Power/ (accessed 5 November 2014).

RELATED ARTICLES Introduction: Renewable Energy Steam Power Generation Gas Turbines for Power and Propulsion Advanced Heat Exchangers for Clean and Sustainable Technology Thermal Energy Efficiency in Industrial Processes Thermal Performance Prediction and Optimization of “Heat Exchangers” by Artificial Intelligence Techniques

REFERENCES Albanakis, C., Yakinthos, K., Kritikos, K., et al. (2009) The effect of heat transfer on the pressure drop through a heat exchanger for aero engine applications. Applied Thermal Engineering, 29, 634–644.


Compact Heat Exchangers in Clean Energy Systems 19 Allen, B. and Gosselin, L. (2008) Optimal geometry and flow arrangement for minimizing the cost of shell and tube condensers. International Journal of Energy Research, 32 (10), 958–969. Boggia, S. and Rüd, K. (2005) Intercooled recuperated aero engine. DGLR Paper, 179. Brey, H.L. (2001) Historical Background and Future Development of the High Temperature Gas Cooled Reactor. Proceedings of Seminar on HTGR Application and Development, p. 1. Briggs, D.E. and Young, E.H. (1962) AIChE Preprint, No. 1, ASME AIChE National Heat Transfer Conference, Houston Tennes, p. 8. Cavallaro, F. (2010) Fuzzy TOPSIS approach for assessing thermalenergy storage in concentrated solar power (CSP) systems. Applied Energy, 87 (2), 496–503. Chaudourne, S. (1992) The heat pipe heat exchangers: design, technology and applications. Design and Operation of Heat Exchangers, EUROTHERM Seminars, 18, 386–396. Chou, S.Y., Chang, Y.H., and Shen, C.Y. (2008) A fuzzy simple additive weighting system under group decision-making for facility location selection with objective/subjective attributes. European Journal of Operational Research, 189 (1), 132–145. Chung, K., Lee, K.S., and Kim, W.S. (2002) Optimization of the design factors for thermal performance of a parallel-flow heat exchanger. International Journal of Heat and Mass Transfer, 45, 4773–4780. Energy Efficiency Office (2000) Compact heat exchangers a training package for engineers, Energy Efficiency Office, London. Fesanghary, M., Damangir, E., and Soleimani, I. (2009) Design optimization of shell and tube heat exchangers using global sensitivity analysis and harmony search algorithm. Applied Thermal Engineering, 29 (5–6), 1026–1031. Goddard Institute for Space Studies (2009) Datasets and Images, http://data.giss.nasa.gov/gistemp/graphs (accessed 20 August 2009). Gunter, A.Y. and Shaw, W.S. (1977) ASME, Paper, 77-WA/HT-5 for meet. Hadidi, A. and Nazari, A. (2013) Design and economic optimization of shell-and-tube heat exchangers using biogeographybased (BBO) algorithm. Applied Thermal Engineering, 51 (1–2), 1263–1272. Ito, Y. and Nagasaki, T. (2011) Suggestion of Intercooled and Recuperated Jet Engine Using Already Equipped Components as Heat Exchangers. 47th AIAA/ASME/SAE/ASEE Joint Propulsion Conference and Exhibit 31 July–03 August 2011, San Diego, CA. Jassim, R.K. and Mohammed, A.A.K. (2003) Computer simulation of thermoeconomic optimization of periodic-flow heat exchangers. Journal of Power and Energy, 217, 559–570. Jiang, W.C., Gong, J.M., Chen, H., and Tu, S.T. (2008a) The effect of filler metal thickness on residual stress and creep for stainless-steel plate–fin structure. International Journal of Pressure Vessels and Piping, 85 (8), 569–574. Jiang, W.C., Gong, J.M., Tu, S.T., and Chen, H. (2008b) Effect of geometric conditions on residual stress of brazed stainless steel plate-fin structure. Nuclear Engineering and Design, 238, 1497–1502. Joda, F., Tahouni, N., and Panjeshahi, M.H. (2013) Application of genetic algorithms in design and optimisation of multi-stream plate fin heat exchangers. Canadian Journal of Chemical Engineering, 91 (5), 870–881.

Jogi, N.G. and Lawankar, S.M. (2012) Heat transfer analysis of corrugated plate heat exchanger of different plate geometry: a review. International Journal of Emerging Technology and Advanced Engineering, 2 (10), 110–115. Kakac, S. and Liu, S. (2002) Heat exchangers: Selection, Rating and Thermal Design, 2nd edn, CRC Press, New York. Kotcioglu, I., Cansiz, A., and Khalaji, M.N. (2013) Experimental investigation for optimization of design parameters in a rectangular duct with plate-fins heat exchanger by Taguchi method. Applied Thermal Engineering, 50 (1), 604–613. Li, Q., Flamant, G., Yuan, X.P., and Luo, L. (2011) Compact heat exchangers: a review and future applications for a new generation of Neveu high temperature solar receivers. Renewable and Sustainable Energy Reviews, 15, 4855–4875. Manglik, R.M. and Bergles, A.E. (1995) Heat transfer and pressure drop correlations for the rectangular offset-strip-fin compact heat exchanger. Experimental Thermal and Fluid Science, 10, 171–180. Martine, H. (1996) A theoretical approach to predict the performance of chevron-type plate heat exchangers. Chemical Engineering and Processing, 35, 301–310. McDonald, C.F., Massardo, A.F., Rodgers, C., and Stone, A. (2008) Recuperated gas turbine aeroengines. Part III: engine concepts for reduced emissions, lower fuel consumption, and noise abatement. Aircraft Engineering and Aerospace Technology: An International Journal, 80 (4), 408–426. Mehendale, S.S., Jacobi, A.M. and Shah, R.K. (1999) Fluid flow and heat transfer at micro- and meso-scales with application to heat exchanger design. Applied Mechanics Reviews, 53 (7), 175–193. Min, J.K., Jeong, J.H., Ha, M.Y., and Kim, K.S. (2009) High temperature heat exchanger studies for applications to gas turbines. Heat and Mass Transfer, 46, 175–186. Minton, P.E. (1970) Designing spiral plate heat exchangers. Chemical Engineering, 77, 103–112. Naghadehi, M.Z., Mikaeil, R., and Ataei, M. (2009) The application of fuzzy analytic hierarchy process (FAHP) approach to selection of optimum underground mining method for Jajarm Bauxite Mine, Iran. Expert Systems with Applications, 36 (4), 8218–8226. Peng, H. and Ling, X. (2008) Optimal design approach for the platefin heat exchangers using neural networks cooperated with genetic algorithms. Applied Thermal Engineering, 28 (5–6), 642–650. Pra, F., Tochon, P., Mauget, C., et al. (2008) Promising designs of compact heat exchangers for modular HTRs using the Brayton cycle. Nuclear Engineering and Design, 238, 3160–3173. Qian, S.W. (2008) Heat Exchanger Design Handbook, Chemical Industry Press, Beijing (In Chinese). Gupta, R.B. and Demirbas, A. (2010) Gasoline, Diesel and Ethanol Biofuels from Grasses and Plants, Cambridge University Press, New York. Rao, R.V. and Patel, V.K. (2010) Thermodynamic optimization of cross flow plate-fin heat exchanger using a particle swarm optimization algorithm. International Journal of Thermal Science, 47 (9), 1712–1721. Reay, D.A. (2002) Compact heat exchangers, enhancement and heat pumps. International Journal of Refrigeration, 25 (4), 460–470. Reay, D.A., Kew, P.A., and McGlen, R.J. (2014) Heat Pipes—Theory, Design and Applications, 6th edn, Elsevier Ltd., Oxford.


20


Reneaume, J.M., Pingaud, H., and Niclout, N. (2000) Optimization of plate fin heat exchangers: a continuous formulation. Chemical Engineering Research and Design, 78 (6), 849–859. Robison, K.K. and Briggs, D.E. (1966) Engineering Progress Symposium Series, 62, 177–184. Sanaye, S. and Hajabdollahi, H. (2013) Thermal-economic multiobjective optimization of plate fin heat exchanger using genetic algorithm. Applied Energy, 87 (6), 1893–1902. Schleicher, R.W., Raffray, A.R., and Wong, C.P. (2000) An Assessment of the Brayton Cycle for High performance Power Plants. 14th Topical Meeting on the Technology of Fusion Energy, GA PROJECT 04437, October, 2000. Schlunder, E.U. (1998) Analogy between heat and momentum transfer. Chemical Engineering and Processing, 37, 103–107. Schönenborn, H., Ebert, E., Simon, B., and Storm, P. (2006) Thermal Mechanical Design of a Heat Exchanger for a Recuperative Aero Engine. Proceedings of ASME Turbo Expo 2004, Power for Land, Sea, and Air, June 14–17, 2004, Vienna. Sepehr, S. and Masoud, D. (2011) Modeling and multi-objective optimization of parallel flow condenser using evolutionary algorithm. Applied Energy, 88 (5), 1568–1577. Shah, R.K. (1981) Classification of heat exchangers, in Heat Exchangers: Thermal-Hydraulic Fundamentals and Design (eds S. Kakac, A.E. Bergles, and F. Mayinger), Hemisphere Publishing, Washington, pp. 9–46. Shah, R.K. (1983) Heat exchanger basic design methods, in Low Reynolds Number Flow Heat Exchangers (eds S. Kakac, R.K. Shah, and A.E. Bergles), Hemisphere, Washington, pp. 21–71. Shah, R.K. (1991) Compact heat exchanger technology and applications, in Heat Exchanger Engineering, Volume 2, Compact Heat Exchangers: Techniques for Size Reduction (eds E.A. Foumeny and P.J. Heggs), Ellis Horwood, London, pp. 1–29. Shah, R.K. and Mueller, A.C. (1985) Heat exchanger basic thermal design methods, in Handbook of Heat Transfer Applications, 2nd edn (eds W.M. Rohsenow, J.P. Hartnett, and E.N. Ganic), McGrawHill, New York. Shah, R.K. and Sekulic, D.P. (2003) Fundamentals of Heat Exchanger Design, John Wiley & Sons, Inc., New York. Sun, K.N., Alwi, S.R.W., and Manan, Z.A. (2013) Heat exchanger network cost optimization considering multiple utilities and different types of heat exchangers. Computers & Chemical Engineering, 49, 194–204.

Thonon, B. and Breuil, E. (2001) Compact Heat Exchangers Technologies for HTRs Recuperator Application. Proceedings of the Technical Committee Meeting on Gas Turbine Power Conversion Systems for Modular HTGRs-Palo Alto, California, IAEATECDOC, November 14–16, 2000, pp. 1–11. Thulukkanam, K. (2013) Heat Exchanger Design Handbook, 2nd edn, Taylor & Francis Group, London. Tu, S.T. and Zhou, G.Y. (2009) Creep of brazed plate-fin structures in high temperature compact heat exchangers. Frontiers of Mechanical Engineering in China, 4 (4), 355–362. Vasiliev, L.L. (2005) Heat pipes in modern heat exchangers. Applied Thermal Engineering, 25 (1), 1–19. Wang, L., Sundén, B., and Manglik, R.M. (2007) Plate Heat Exchangers: Design, Applications and Performance. Wilfert, G., Kriegl, B., Scheugenpflug, H., et al. (2005) CLEANValidation of a High Efficient Low NOx Core, a GTF High Speed Turbine and an Integration of a Recuperator in an Environmental Friendly Engine Concept. 41st AIAA/ASME/ASEE Joint Propulsion Conference, Tucson, AZ, July 10–13, AIAA-2005-4195. Wilfert, G., Sieber, J., Rolt, A., et al. (2007) New environmental friendly aero engine core concepts. ISABE-2007-1120, Beijing, September, 2007. Yekoladio, P.J., Bello-Ochende, T., and Meyer, J.P. (2013) Design and optimization of a downhole coaxial heat exchanger for an enhanced geothermal system (EGS). Renewable Energy, 55, 128–137. Zhou, G.Y. and Tu, S.T. (2007) Viscoelastic analysis of rectangular passage of microchanneled plates subjected to internal pressure. International Journal of Solids and Structures, 44 (21), 6791–6804. Zhou, G.Y., Wu, E., and Tu, S.T. (2008) Techno-economic study on compact heat exchangers. International Journal of Energy Research, 3212, 1119–1127. Zhou, G.Y., Tu, S.T., Xuan, F.Z., and Wang, Z.D. (2011) Viscoelastic model to describe mechanical response of compact heat exchangers with plate-foam structure. International Journal of Mechanical Sciences, 53, 1069–1076. Zhou, G.Y., Wu, E., and Tu, S.T. (2014) Optimum selection of compact heat exchangers using non-structural fuzzy decision method. Applied Energy, 113, 1801–1809. Zhuang, J., Xu, T.M., and Shi, S.C. (1989) Heat Pipe and Heat Pipe Heat Exchanger, Shanghai Jiaotong University Press, Shanghai.
