Conical Face Seal; Dry Lubrication of Stainless; Seal Misalign- ment; Carbon ... spring. In both the rotary joint arrangements (Figs. 1(a) and 1(b)) a secondary ...
Tribology Transactions, 52: 47-58, 2009 C Society of Tribologists and Lubrication Engineers Copyright ISSN: 1040-2004 print / 1547-397X online DOI: 10.1080/10402000802163017
Design and Development of a Test Setup for Online Wear Monitoring of Mechanical Face Seals Using a Torque Sensor
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SHASHIKANT S. GOILKAR and HARISH HIRANI Depatment of Mechanical Engineering IIT Bombay Powai, Mumbai - 400076, India
Online condition monitoring technique helps to detect the root cause of failure of any machine. The present article describes a design of the online failure monitoring facility for mechanical face seals. The experimental test facility is to operate the seal (i.e., carbon-graphite) in real conditions of fluid pressure, temperature, and misalignment, which occur in an industrial environment. The developed test setup consists of two proximity displacement sensors (accuracy 2 µm), one fiber-optic sensor (accuracy 10 µm), one accelerometer (3.97 mV/ms−2 ), and one non-contact torque sensor (accuracy 0.05 N.m). To validate the test facility, a typical conical carbon graphite (C = 59.195%, O = 4.625%, and Sb = 36.18%.) mechanical face seal (outer dia = 82 mm and inner dia = 63 mm) for a rotary joint used in steam/hot water was selected. The root cause of failure of such seals has been identified. Finally, recommendations have been made that provide some assistance to design the mechanical face seal.
the housing to the atmosphere is prevented by mechanical contact between the tapered rotating shaft and the mating seal face. The continuous mechanical contact is retained with the help of a spring. In both the rotary joint arrangements (Figs. 1(a) and 1(b)) a secondary seal is employed to guide the rotating shaft and bear its weight. Mechanical face seals are often designed considering the hydrostatic (Lipschitz (1)), the hydrodynamic (Etsion and Pascovici (2)), or the squeeze (Etsion and Michael (3)) lubrication mechanism. Such lubrication mechanisms, if properly achieved, may provide infinite seal life. However, the expected seal life is generally in the range of two to ten years, considering the uncertainty in the strength of the seal materials, unforeseen operating conditions, etc. However, unpredictable seal failures, with the seal life equal to 2 days to 2 months, have been observed in industries. Such seal failures cause direct (leakage of fluid, loss of prepared material in paper industry, etc.) as well as indirect (downtime cost, maintenance cost, reputation of company, etc.) losses. Two types of seal failures, observed within two months after the installation of new seal rings, are shown in Fig. 2. On a few occasions failures were repetitive and required the replacement of seal rings every 15 days. Such unpredictable life (two days, two months, or two years) of the same seal may occur due to improper design, wrong assembly, or incomplete operating instructions. In the present study, an experimental test facility has been developed to investigate the effect of the rotational speed (40 rpm to 800 rpm), the steam pressure (0 to 12 bars), and the angular misalignment on the seal life. Online failure of the seal has been perceived by continuous monitoring of frictional torque exerted by the seals on the rotating shafts. The main aim of the developed setup is to find out the root cause of the seal failure so that corrective actions can be suggested.
KEY WORDS Conical Face Seal; Dry Lubrication of Stainless; Seal Misalignment; Carbon Graphite Seals
INTRODUCTION A mechanical face seal is an important component of a variety of rotary joints and pumps, which are used in chemical, textile, petrochemical, and process industries. Two typical mechanical face seal arrangements in rotary joint are shown in Fig. 1. The main function of these rotary joints (sketched in Fig. 1) is to supply steam from a stationary pipe to the rotating drum used in the paper industry. The main difference between Figs. 1(a) and 1(b) is the siphon arrangement, which is required to take out the condensate from the drying drum. In X-assembly (Fig. 1a) only one pair of the conical shaft-seal interface is used, while in Y-assembly (Fig. 1b) two pairs of such interfaces are used. In both the configurations the face seals are fixed to the housing, and the steam-leakage from
DESIGN OF ONLINE FAILURE MONITORING FACILITY To simulate the industry environment in a laboratory, the test setup requires:
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Manuscript received December 14, 2007 Manuscript accepted April 22, 2008 Review led by Jim Netzel
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A variable-speed motor, which can run continuously for hours. The control of the rotational speed is essential to accelerate the wear of the seals. Generally, accelerated wear depends on the lubrication mechanism. If the seal interface is dry, then
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Fig. 1—Construction of rotary joints.
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increasing the rotational speed will accelerate the wear rate. However, if the mechanism is hydrodynamic lubrication, then decreasing the rotational speed will enhance the wear rate. A 3-phase 2-HP AC 750 rated rpm motor with a forced cooling system has been selected to run the experiment continuously for hours. A variable-frequency (1 to 200 Hz) drive system (Toshiba, VF-A7, Japan) to manage the motor speed has been chosen. With the help of the frequency drive, the motor shaft can be operated in the range of 30 rpm to 3000 rpm. A supply source having control of the flow rate, the pressure, and the temperature. The regulation of supply conditions of fluid to be sealed is essential to simulate the extreme industry conditions. For example, seals of a rotary joint used in the paper industry operate under 4–12 bar steam pressure. Similarly, the operating temperature may vary from 110 to 160 degrees centigrade. A steam boiler, having control of the flow and pressure of steam, has been selected for the present study. Sensors to measure uniform wear, misalignment, the temperature, the pressure, the acceleration, and the dynamic friction force. Anderson, et al. (4) used a piezoelectric shear transducer to diagnose the contact condition of the seal interface. To detect the position of the rotating mating ring relative to the stationary primary ring, Sehnal, et al. (5) used eddy current proximity
sensors. Choy, et al. (6) used acoustic sensors to diagnose the health of mechanical seals. To measure the film thickness and thermal distortions, Doust and Parmar (7) used miniature capacitance probes and thermocouples. In the present study, two eddy current proximity sensors (Bentley Nevada, Model 3300 XL NSv) have been chosen to indicate the variation in the axial position of the shaft relative to the stationary seal ring. A charge amplifier unit (B & K type 4366 and type 2635, respectively; Denmark) has been selected to measure the change in the vibration level, which can provide an indication of initiation of seal failure. To measure the number of rotations and the rotational shaft speed, a cycle counter (Selectron RC102-A; Mumbai, India) has been engaged. In addition to these regular sensors, a torque sensor (Lorenz, Model DR-2513) is selected to detect the instantaneous values of torques at various angular positions of the shaft running in mechanical contact of the seal face. In the authors’ view, seal failures shown in Fig. 2 may have happened due to the misalignment. To assure chances of misalignment, micrographs of cracked primary seal, as shown in Fig. 3, were taken through an electron microscope. These micrographs (Fig. 3) show nonuniform wear of the seal surface interfaced with the shaft. Major wear near the crack indicates very high localized loading. Such nonuniform loading occurs
Fig. 2—Seal failures of primary and secondary seals of rotary joints used in the paper industry.
Online Condition Monitoring Technique
between the contacting cylindrical surfaces at different angular positions of complete rotation. A non-contact-type torque sensor may be the best to register such dynamic variations in the coefficient of friction. Further, any breakage and/or crack on the seal ring will change the loading pattern exerted on the seal surface. On cracking, the seal surface is subjected to extra bending and direct shear loads. In addition, the stiffness of
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for two reasons. One reason is the excessive radial clearance between the primary seal and the shaft, and the other reason is the angular misalignment between the seal and the shaft axes. Tolerance of the seal and the shaft did not indicate a chance of radial clearance between them; therefore, it was decided to assume misalignment as a key failure mechanism. The misalignment in rotating parts varies the frictional resistance
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Fig. 3—Photomicrographs of cracked primary seal indicating maximum wear near crack, moderate wear at 90 degrees, and minimum wear at 180 degrees phase from of crack.
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Fig. 4—Model of test rig.
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the seal near the crack will be lower and that portion will provide smaller resistance to the shaft movement. The change in stiffness and the load pattern immediately will alter the torque resistance offered by the seals. Therefore, the torque measurement may be a very reliable diagnostic tool to detect the seal failure. To measure the misalignment, a dial gauge setup has been arranged. A data acquisition system to convert analog data to digital data with required resolution. For the present study, a mutlichannel 1-MHz NI-DAQ-7 data acquisition system has been chosen. Computer hardware and software to access and process the data. Suitable software to process the collected data is essential. In the present study a Matlab code has been written to perform the synchronous cycle averaging to reduce the noise and provide meaningful data.
EXPERIMENTAL SETUP The experimental test facility has been developed to operate the seal (i.e., carbon-graphite) in real conditions of fluid pressure, temperature, and misalignment (by tilting the axis of rotary joint), which occur in the industrial environment. In order to get a prior idea about the probable difficulties that may arise during fabrication of the test rig, a 3-D solid-model (Fig. 4) of the setup was made. Figure 4 shows that the base plate along with the support-
Fig. 5—Procedure to create misalignment in the rotary joint.
ing structure provides a main common reference platform for the rotary joint and the driving motor. To create the misalignment condition, slots are provided along the radial direction in the base plate. In addition, misalignment can be introduced by tilting the axis of the shaft by inserting the packing (thin metallic sheets of known thickness) between the base plate and the rotary joint as shown in Fig. 5. To test the performance of seals in the steam and the water environment, the rotary joint has been connected to the boiler using a piston valve, the pressure-regulating valve, the moisture separator, etc., as shown in Fig. 6. This figure illustrates that to control the steam pressure, a pressure-regulating valve, along with two pressure gauges, has been used. Such arrangement of pressure gauges, one before the supply and another after the pressure-regulating valve, helps to maintain the exact pressure required during the testing. To regulate the dryness of steam a moisture separator module consisting of a sight glass, a thermodynamic trap, and a diffuser has been used. To extract the condensate from the rotary joint, a separate condensate removing system, which utilizes the help of a float trap, has been used. A schematic of the total setup is shown in Fig. 7. Photograph of the developed setup is shown in Figs. 8 and 9. In the present experimental study, carbon-graphite seals having a conical (cone angle = 45◦ ) mating face geometry, as shown in Fig. 1, with an outer and an inner diameter of 82 mm and 63 mm, respectively, have been used. The corrosion-resistant stainless steel shaft works as one of the seal rings. The hardness of the stainless steel shaft is
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Fig. 6—Steam connection diagram for test rig.
slightly higher (78 HRB), which makes the shaft more resistant to abrasive wear compared to the carbon-graphite (C G) seal ring. Therefore, the C G seal ring acts as a sacrificial element. EDAX element analysis indicates that a carbon-graphite seal contains C = 59.195%, O = 4.625 weight percentage, and Sb = 36.18%. The use of antimony (∼ 36 weight percentage) in carbon-graphite seals increases the heat transfer rate and decreases the changes of thermal cracking of the graphite seals.
Fig. 8—Developed setup.
each operating condition have been recorded using a computerized data acquisition system. Two hundred and fifty-six data points from each sensor for each rotation have been acquired.
EXPERIMENTAL PROCEDURE The operating conditions in the paper industry change according to the process requirement, which in turn imposes different lubrication conditions—i.e., full film, mixed, or dry running—on the seal interface. For example, the seal interface experiences dry running for about an hour during startup. Then the steam is passed inside the dryer. The steam pressure gradually increases and then reaches the steady condition, which leads to hydrostatic lubrication of the seal interface. During shutdown once again for almost one hour dry running is experienced by the seal interface. This practice is generally being followed in paper industries. In addition, the assembly of a rotary joint may induce misalignment, which in turn does not allow the hydrostatic lubrication of the seal interface even in full pressure steady steam condition. Therefore, to simulate the real industry environment for the seal interface, experiments were planned for Assembly-X (Fig. 1a) and Assembly-Y (Fig. 1b) in the sequence depicted in Fig. 10. The output data from the torque sensor, the displacement sensor, and the accelerometer for ten rotations of the shaft under
ANALYSIS OF EXPERIMENTAL RESULTS Figure 11 provides the comparison between torque resisted by seals of X- and Y-assemblies. This figure illustrates the relatively constant frictional torque exerted by X-assembly, while fluctuations in torque is resisted by Y-assembly. This comparison points toward an inherent misaligned tendency of Y-assembly, due to which torque fluctuations occur. The clearance (10 microns to 140 microns) fit between the two shafts, as shown in Fig. 12, allows misalignment to happen in the Y-assembly. Further, Fig. 11 shows a slight increase in the torque with an increase in the rotational speed from 60 rpm to 120 rpm. The average torque from X-assembly is 4.3 N.m when the shaft rotates at 60 rpm, while 4.4 N.m average torque resistance occurs when the shaft of X operates at 120 rpm. The average value of the torque resisted by Y-assembly is 9.4 and 11.0 N.m when the shaft rotates at 60 rpm and 120 rpm, respectively. To validate the measured torque results, it is better to derive the coefficient of friction and compare it with the established results. The expression for the coefficient of friction between the mating faces of the shaft and the seal in the rotary joint, as shown in Fig. 13, can be expressed as (Shigley and Mischke (8)), µ=
r2 − r22 3 T sin α 13 2W r1 − r32
[1]
where T is the experimentally determined frictional torque in N.m, W is the spring force exerted on the seal surfaces in N, α (= 45◦ in the present study) is the cone angle, r1 (= 41 mm in the present study) is the outer radius of the seal, and r2 (= 31.5 mm in the present study) is the inner radius of the seal. Substituting the values of the geometric parameters, Fig. 7—Schematic of test setup.
µ = 19.4
T W
[2]
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Fig. 9—Arrangement of sensors.
Fig. 10—Sequence for experiments.
S. S. GOILKAR AND H. HIRANI
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Fig. 11—Comparison of torque measured during operation of X (srj) and Y (rrj) assemblies under aligned conditions.
The spring force depends on the geometry of the spring and the pre-compression. In the present study, a close and ground ended helical spring, made of spring material SS302, having a wire diameter (d) = 6 mm, and a total number of turns (N) = 5, has been used. In the case of the Y-assembly, mean coil diameter (D) = 82 mm, free length = 123 mm, and pre-compression = 63 mm has been used. This spring imposes an axial load equal to 426 N on the seal interface. In the case of the X-assembly the mean diameter of spring = 100 mm, free length = 132, and precompression = 86 mm has been used. Therefore, the spring applies 321 N of the axial load to close the shaft on the seal surface. On substituting the values of W and T, the coefficient of the friction for the X-assembly at 60 rpm is µ60X = 0.26. Similarly, µ120X = 0.27, µ60Y = 0.43, and µ120Y = 0.50 can be obtained using Eq. [2]. Figure 1b shows that there are two seal interfaces in the case of the Y-assembly; therefore, the coefficient of friction for the single seal surface will halve. In other words, µ60Y for single surface = 0.215, and µ120Y for single surface will be 0.25. These values of the coefficient of friction for carbon-graphite versus stainless steel under dry running conditions are reasonable (Jones (9)). Hence, the torque sensor provides a reliable torque resisted by seal interfaces.
Figure 14 illustrates the torque variation in the X-assembly misaligned by 0.014 radians. The values of the maximum, average, and minimum torques equal to 7.2 N.m, 6.0 N.m, and 5.1 N.m, respectively, have been observed in the misaligned Xassembly operating at 60 rpm. To compare these results and implement a trend analysis technique, a synchronous time averaging technique may be useful. To employ this technique in the present study, data obtained from fiber-optic displacement sensor (Philtec, Type RC-90) have been used. Figure 15 demonstrates the displacement and torque data. Sharp peaks of displacement readings indicate the end of the previous cycle and start of the new cycle. Using these data, a trend analysis can easily be made. The trend analyses for X- and Y-assemblies have been plotted in Figs. 16 and 17, respectively. Figure 16 illustrates an increase in the torque fluctuation with an increase in the time duration of the misaligned dry condition. This torque fluctuation is an indication of seal wear. Figure 17 notifies the torque fluctuations of the Yassembly. One major difference between the torque behavior of X and Y observed from Figs. 16 and 17 is the progressive flattening of the peak torque of Y-assembly. This indicates the progressive wear of the seal. Finally, the curve for t = 55 min in Fig. 17 demonstrates
S. S. GOILKAR AND H. HIRANI
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Fig. 13—Axially loaded mechanical seal.
Fig. 12—Clearance fit between shaft 1 and shaft 2 of Y-assembly.
the major variation in torque behavior, which is an indication of seal failure.
DISCUSSION AND RECOMMENDATIONS In the present article, the main emphasis has been given to seal failure caused by misalignment. To online monitor the failure of the mechanical seal, three types of sensors—accelerometer, torque sensor, and displacement sensors—have been used. The observed accelerometer readings do not provide any conclusive study. One possible reason for such failure is the location of the sensor. The accelerometer was mounted on the stationary housing at one particular location, while the Y-assembly has four (two primary and two secondary) seal interfaces. Displacement readings from proximity sensors also could not provide conclusive
data. One possible reason is the mixing of circumferential and axial displacements. Due to misalignment, the displacement sensor observes the cyclic variation in the data. In addition, axial wear of the seal affects the displacement readings. Superposition of axial and circumferential displacement makes it difficult to conclude the seal failure. However, data from the torque sensor show a definite trend, which can be related to seal wear. The main advantage of the torque sensor is that it can be connected to the rotating parts, and its magnetic pick-up provides reliable data without usage of any slip-ring arrangement. In the present study, the torque sensor connects the motor with the rotary joint and any geometric variation at the seal interface is immediately reflected in the torque data. Based on the torque data, shown in Figs. 16 and 17, it can be stated that X-assembly experiences mild wear even in misaligned dry conditions, while seals of Y-assemblies undergo thorough severe wear and breakage under a dry misaligned condition. With such intuitions, the assemblies of Y and X were opened. The photographs of seals are shown in Figs. 18 and 19. A photograph of the shaft in Fig. 18 indicates polishing wear, which probably occurred due to wear debris of carbon graphite seals. The breakage of the secondary seal and the pitting of the primary seal is also shown in Fig. 19. These photographs validate the seal failure estimated from torque data of Fig. 17. Based on the photographs shown in Fig. 18, it can be said that the present study illustrates the possibility of seal failure of Y-assemblies within 3 h of operation. Such failures generally occur under dry and misaligned conditions. The dry conditions are enforced at the start and at the shutdown of the dryer by process industries. Therefore, avoiding the dry condition is almost impossible; however, the time duration of such dry conditions can be minimized to enhance the seal life. Changing seal material (Blau and Martin (10)) (such as silicon nitride or silicon carbide) in place of carbon graphite may reduce the wear rate in dry conditions. However, the present research indicates misalignment as one of the factors affecting the wear rate of the seal. Therefore, the major recommendation of the present research work is to minimize the misalignment as far as possible. As the present design of the Y seal assembly uses two shafts having a clearance fit, the misalignment is inherent in the seal assembly. Introducing a
Online Condition Monitoring Technique
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Fig. 14—Torque characteristics of X assembly operating at 60 rpm under misalignment of 0.014 radians.
Fig. 15—Displacement and torque data for misaligned X assembly operating at 60 rpm.
Fig. 16—Trend analysis of X assembly operating at 240 rpm under dry misaligned condition.
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Fig. 17—Trend analysis of Y assembly operating at 240 rpm under dry misaligned condition.
Fig. 18—Photographs of shaft, primary seal, and secondary seal of Y-assembly after 30,000 operating cycles.
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Fig. 19—Photographs of primary and secondary seals of X-assembly after 30,000 operating cycles.
major single shaft in place of two shafts and modifying correspondingly the design of the Y-assembly will provide better seal life. Figure 19 shows mild wear scars on the primary seal of Xassembly. These wear marks indicate the misalignment operation experienced by the seal interface. This photograph validates the mild progressive wear concluded by the torque data plotted in Fig. 16. Quantification of wear loss is performed by measuring the weight of the seal rings using a weighing machine (CP 423 S, Sarto-
rius, Goettingen, Germany) of 420 g capacity with an accuracy of 0.001 g. Wear loss data are listed in Table 1. In this table, PM stands for “primary seal motor side,” SM stands for “secondary seal motor side,” PB stands for “primary seal boiler side,” and SB stands for “secondary seal boiler side.” The sliding distance has been calculated using the π *D*N formula, where N is the number of cycles. Wear loss data clearly indicate that the seals of the Y-assembly experience a high wear rate compared to the seals of the X-assembly.
CONCLUSIONS TABLE 1—WEAR LOSS DATA FOR PRIMARY AND SECONDARY SEALS Test Number 1
Type of Joint
Type of Seal
Number of Cycles
Sliding Distance in mm
Weight Loss in mg
Y
PM SM PB SB PM SM PB SB PM SM PB SB PM SM PM SM PM SM
30,322
6,906,304.6 5,691,747.5 6,906,304.6 5,691,747.5 6,967,345.7 5,742,053.8 6,967,345.7 5,742,053.8 6,916,781.8 5,700,382.1 6,916,781.8 5,700,382.1 7,132,931.2 5,878,519.1 7,915,988.9 6,523,866.6 8,359,676 6,889,526
101 — 97 — 76 21 77 20 69 19 70 22 28 — 33 26 32 23
2
Y
3
Y
4
X
5
X
6
X
30,590
30,368
31,317 34,755 36,703
The present research work is aimed at the development of a test setup that emulates the conditions dealt with by mechanical face seals. Two commonly used rotary joints containing face seals have been investigated. The conclusions of the present research work are
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A torque sensor, with an appropriate data acquisition and analysis system, predicts the seal failure with authenticity. Misalignment is the root cause of failure of face seals used in Y-assemblies. The developed test setup can predict the failure of seals under various rotational speeds, lubrication mechanisms, and spring loads.
ACKNOWLEDGEMENT The authors wish to express their appreciation to Forbes Marshall Company for supporting this work through a research grant.
REFERENCES (1) Lipschitz, A. (1989), “Dynamic Performance of the Stepped Hydrostatic Circumferential Gas Seal,” Tribology Transactions, 32(2), pp 189–196. (2) Etsion, I. and Pascovici, M.D. (1996), “Hydrodynamic Effects on the Boiling Interface in a Misaligned, Two-Phase, Mechanical Seal—A Qualitative Study,” Tribology Transactions, 39(4), pp 922–928.
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(3) Etsion, I. and Michael, O. (1994), “Enhancing Sealing and Dynamic Performance with Partially Porous Mechanical Face Seals,” Tribology Transactions, 37(4), pp 701–710. (4) Anderson, W. B., Jarzynski, J. and Salant, R.F. (2001), “A Condition Monitor for Liquid Lubricated Mechanical Seals,” Tribology Transactions, 44(3), pp 479–483. (5) Sehnal, J., Sedy, J., Zobens, A. and Etsion, I. (1983), “Performance of the Coned-Face End Seal with Regard to Energy Conservation,” Tribology Transactions, 26(4), pp 415–429. (6) Choy, F.K., Polyshchuk, V., Braun, M.J. and Lu, M. (1999), “A New Method in Acoustic Health Monitoring in Mechanical Seals,” Tribology Transactions, 42(1), pp 46–52.
(7) Doust, T.G. and Parmar, A. (1986), “An Experimental and Theoretical Study of Pressure and Thermal Distortions in a Mechanical Seal,” Tribology Transactions, 29(2), pp 151–159. (8) Shigley, J.E. and Mischke, C.R. (2003), Mechanical Engineering Design (in SI Units), 6th ed., Tata McGraw Hill, India. (9) Jones, G.A. (2004), “On the Tribological Behaviour of Mechanical Seal Face Materials in Dry Line Contact, Part I. Mechanical Carbon,” Wear, 256, pp 415–432. (10) Blau, P.J. and Martin, R.L. (1994), “Friction and Wear of Carbon-Graphite Materials against Metal and Ceramic Counterfaces,” Tribology International, 27(6), pp 413–422.
