design and development of a test setup for online wear...

Tribology Transactions, 52: 47-58, 2009Copyright C© Society of Tribologists and Lubrication EngineersISSN: 1040-2004 print / 1547-397X onlineDOI: 10.1080/10402000802163017

Design and Development of a Test Setup for Online WearMonitoring of Mechanical Face Seals Using a Torque

SensorSHASHIKANT S. GOILKAR and HARISH HIRANI

Depatment of Mechanical EngineeringIIT Bombay

Powai, Mumbai - 400076, India

Online condition monitoring technique helps to detect the

root cause of failure of any machine. The present article de-

scribes a design of the online failure monitoring facility for me-

chanical 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 en-

vironment. 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.

KEY WORDS

Conical Face Seal; Dry Lubrication of Stainless; Seal Misalign-ment; Carbon Graphite Seals

INTRODUCTION

A mechanical face seal is an important component of a vari-ety of rotary joints and pumps, which are used in chemical, textile,petrochemical, and process industries. Two typical mechanical faceseal arrangements in rotary joint are shown in Fig. 1. The mainfunction of these rotary joints (sketched in Fig. 1) is to supply steamfrom a stationary pipe to the rotating drum used in the paper indus-try. The main difference between Figs. 1(a) and 1(b) is the siphonarrangement, which is required to take out the condensate fromthe drying drum. In X-assembly (Fig. 1a) only one pair of the con-ical shaft-seal interface is used, while in Y-assembly (Fig. 1b) twopairs of such interfaces are used. In both the configurations theface seals are fixed to the housing, and the steam-leakage from

Manuscript received December 14, 2007Manuscript accepted April 22, 2008

Review led by Jim Netzel

the housing to the atmosphere is prevented by mechanical con-tact between the tapered rotating shaft and the mating seal face.The continuous mechanical contact is retained with the help of aspring. In both the rotary joint arrangements (Figs. 1(a) and 1(b))a secondary seal is employed to guide the rotating shaft and bearits weight.

Mechanical face seals are often designed considering the hy-drostatic (Lipschitz (1)), the hydrodynamic (Etsion and Pascovici(2)), or the squeeze (Etsion and Michael (3)) lubrication mech-anism. Such lubrication mechanisms, if properly achieved, mayprovide infinite seal life. However, the expected seal life is gener-ally in the range of two to ten years, considering the uncertaintyin the strength of the seal materials, unforeseen operating condi-tions, etc. However, unpredictable seal failures, with the seal lifeequal to 2 days to 2 months, have been observed in industries.Such seal failures cause direct (leakage of fluid, loss of preparedmaterial in paper industry, etc.) as well as indirect (downtime cost,maintenance cost, reputation of company, etc.) losses. Two typesof seal failures, observed within two months after the installationof new seal rings, are shown in Fig. 2. On a few occasions failureswere repetitive and required the replacement of seal rings every15 days. Such unpredictable life (two days, two months, or twoyears) of the same seal may occur due to improper design, wrongassembly, or incomplete operating instructions.

In the present study, an experimental test facility has been de-veloped to investigate the effect of the rotational speed (40 rpmto 800 rpm), the steam pressure (0 to 12 bars), and the angularmisalignment on the seal life. Online failure of the seal has beenperceived by continuous monitoring of frictional torque exertedby the seals on the rotating shafts. The main aim of the devel-oped setup is to find out the root cause of the seal failure so thatcorrective actions can be suggested.

DESIGN OF ONLINE FAILURE MONITORING FACILITY

To simulate the industry environment in a laboratory, the testsetup requires:

� A variable-speed motor, which can run continuously for hours.The control of the rotational speed is essential to acceleratethe wear of the seals. Generally, accelerated wear depends onthe lubrication mechanism. If the seal interface is dry, then

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48 S. S. GOILKAR AND H. HIRANI

Fig. 1—Construction of rotary joints.

increasing the rotational speed will accelerate the wear rate.However, if the mechanism is hydrodynamic lubrication, thendecreasing the rotational speed will enhance the wear rate. A3-phase 2-HP AC 750 rated rpm motor with a forced coolingsystem has been selected to run the experiment continuouslyfor hours. A variable-frequency (1 to 200 Hz) drive system(Toshiba, VF-A7, Japan) to manage the motor speed has beenchosen. With the help of the frequency drive, the motor shaftcan 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 offluid to be sealed is essential to simulate the extreme industryconditions. For example, seals of a rotary joint used in the paperindustry operate under 4–12 bar steam pressure. Similarly, theoperating temperature may vary from 110 to 160 degrees centi-grade. A steam boiler, having control of the flow and pressureof steam, has been selected for the present study.

� Sensors to measure uniform wear, misalignment, the temper-ature, the pressure, the acceleration, and the dynamic frictionforce. Anderson, et al. (4) used a piezoelectric shear transducerto diagnose the contact condition of the seal interface. To detectthe position of the rotating mating ring relative to the station-ary primary ring, Sehnal, et al. (5) used eddy current proximity

sensors. Choy, et al. (6) used acoustic sensors to diagnose thehealth of mechanical seals. To measure the film thickness andthermal distortions, Doust and Parmar (7) used miniature ca-pacitance probes and thermocouples. In the present study, twoeddy current proximity sensors (Bentley Nevada, Model 3300XL NSv) have been chosen to indicate the variation in the ax-ial position of the shaft relative to the stationary seal ring. Acharge amplifier unit (B & K type 4366 and type 2635, respec-tively; Denmark) has been selected to measure the change inthe vibration level, which can provide an indication of initi-ation of seal failure. To measure the number of rotations andthe rotational shaft speed, a cycle counter (Selectron RC102-A;Mumbai, India) has been engaged. In addition to these regularsensors, a torque sensor (Lorenz, Model DR-2513) is selectedto detect the instantaneous values of torques at various angu-lar positions of the shaft running in mechanical contact of theseal face. In the authors’ view, seal failures shown in Fig. 2 mayhave happened due to the misalignment. To assure chances ofmisalignment, micrographs of cracked primary seal, as shownin Fig. 3, were taken through an electron microscope. Thesemicrographs (Fig. 3) show nonuniform wear of the seal surfaceinterfaced with the shaft. Major wear near the crack indicatesvery 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 49

for two reasons. One reason is the excessive radial clearancebetween the primary seal and the shaft, and the other reasonis the angular misalignment between the seal and the shaftaxes. Tolerance of the seal and the shaft did not indicate achance of radial clearance between them; therefore, it was de-cided to assume misalignment as a key failure mechanism. Themisalignment in rotating parts varies the frictional resistance

between the contacting cylindrical surfaces at different angu-lar positions of complete rotation. A non-contact-type torquesensor may be the best to register such dynamic variations inthe coefficient of friction. Further, any breakage and/or crackon the seal ring will change the loading pattern exerted on theseal surface. On cracking, the seal surface is subjected to ex-tra bending and direct shear loads. In addition, the stiffness of

Fig. 3—Photomicrographs of cracked primary seal indicating maximum wear near crack, moderate wear at 90 degrees, and minimum wear at 180 degreesphase from of crack.



Fig. 4—Model of test rig.

the seal near the crack will be lower and that portion will pro-vide smaller resistance to the shaft movement. The change instiffness and the load pattern immediately will alter the torqueresistance offered by the seals. Therefore, the torque measure-ment may be a very reliable diagnostic tool to detect the sealfailure. To measure the misalignment, a dial gauge setup hasbeen arranged.

� A data acquisition system to convert analog data to digitaldata with required resolution. For the present study, a mutli-channel 1-MHz NI-DAQ-7 data acquisition system has beenchosen.

� Computer hardware and software to access and process thedata. Suitable software to process the collected data is essen-tial. In the present study a Matlab code has been written toperform the synchronous cycle averaging to reduce the noiseand provide meaningful data.

EXPERIMENTAL SETUP

The experimental test facility has been developed to operatethe 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 prioridea about the probable difficulties that may arise during fabri-cation of the test rig, a 3-D solid-model (Fig. 4) of the setup wasmade. Figure 4 shows that the base plate along with the support-

ing structure provides a main common reference platform for therotary joint and the driving motor. To create the misalignmentcondition, slots are provided along the radial direction in the baseplate. In addition, misalignment can be introduced by tilting theaxis of the shaft by inserting the packing (thin metallic sheets ofknown thickness) between the base plate and the rotary joint asshown in Fig. 5.

To test the performance of seals in the steam and the water en-vironment, the rotary joint has been connected to the boiler usinga piston valve, the pressure-regulating valve, the moisture separa-tor, etc., as shown in Fig. 6. This figure illustrates that to control thesteam pressure, a pressure-regulating valve, along with two pres-sure gauges, has been used. Such arrangement of pressure gauges,one before the supply and another after the pressure-regulatingvalve, helps to maintain the exact pressure required during thetesting. To regulate the dryness of steam a moisture separatormodule consisting of a sight glass, a thermodynamic trap, and adiffuser has been used. To extract the condensate from the rotaryjoint, a separate condensate removing system, which utilizes thehelp of a float trap, has been used.

A schematic of the total setup is shown in Fig. 7. Photographof the developed setup is shown in Figs. 8 and 9. In the presentexperimental study, carbon-graphite seals having a conical (coneangle = 45◦) mating face geometry, as shown in Fig. 1, with anouter and an inner diameter of 82 mm and 63 mm, respectively,have been used. The corrosion-resistant stainless steel shaft worksas one of the seal rings. The hardness of the stainless steel shaft is

Fig. 5—Procedure to create misalignment in the rotary joint.



Fig. 6—Steam connection diagram for test rig.

slightly higher (78 HRB), which makes the shaft more resistant toabrasive wear compared to the carbon-graphite (C G) seal ring.Therefore, the C G seal ring acts as a sacrificial element. EDAXelement analysis indicates that a carbon-graphite seal containsC = 59.195%, O = 4.625 weight percentage, and Sb = 36.18%.The use of antimony (∼ 36 weight percentage) in carbon-graphiteseals increases the heat transfer rate and decreases the changes ofthermal cracking of the graphite seals.

EXPERIMENTAL PROCEDURE

The operating conditions in the paper industry change accord-ing to the process requirement, which in turn imposes differentlubrication conditions—i.e., full film, mixed, or dry running—onthe seal interface. For example, the seal interface experiences dryrunning for about an hour during startup. Then the steam is passedinside the dryer. The steam pressure gradually increases and thenreaches the steady condition, which leads to hydrostatic lubrica-tion of the seal interface. During shutdown once again for almostone hour dry running is experienced by the seal interface. Thispractice is generally being followed in paper industries. In ad-dition, the assembly of a rotary joint may induce misalignment,which in turn does not allow the hydrostatic lubrication of theseal interface even in full pressure steady steam condition. There-fore, to simulate the real industry environment for the seal in-terface, experiments were planned for Assembly-X (Fig. 1a) andAssembly-Y (Fig. 1b) in the sequence depicted in Fig. 10.

The output data from the torque sensor, the displacement sen-sor, and the accelerometer for ten rotations of the shaft under

Fig. 7—Schematic of test setup.

Fig. 8—Developed setup.

each operating condition have been recorded using a computer-ized data acquisition system. Two hundred and fifty-six data pointsfrom each sensor for each rotation have been acquired.

ANALYSIS OF EXPERIMENTAL RESULTS

Figure 11 provides the comparison between torque resistedby seals of X- and Y-assemblies. This figure illustrates the rel-atively constant frictional torque exerted by X-assembly, whilefluctuations in torque is resisted by Y-assembly. This comparisonpoints toward an inherent misaligned tendency of Y-assembly, dueto which torque fluctuations occur. The clearance (10 microns to140 microns) fit between the two shafts, as shown in Fig. 12, al-lows misalignment to happen in the Y-assembly. Further, Fig. 11shows a slight increase in the torque with an increase in the ro-tational speed from 60 rpm to 120 rpm. The average torque fromX-assembly is 4.3 N.m when the shaft rotates at 60 rpm, while4.4 N.m average torque resistance occurs when the shaft of X op-erates at 120 rpm. The average value of the torque resisted byY-assembly is 9.4 and 11.0 N.m when the shaft rotates at 60 rpmand 120 rpm, respectively.

To validate the measured torque results, it is better to derive thecoefficient of friction and compare it with the established results.The expression for the coefficient of friction between the matingfaces of the shaft and the seal in the rotary joint, as shown in Fig.13, can be expressed as (Shigley and Mischke (8)),

µ = 32

TW

sin αr2

1 − r22

r31 − r3

2

[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 presentstudy) is the outer radius of the seal, and r2 (= 31.5 mm in thepresent study) is the inner radius of the seal. Substituting the valuesof the geometric parameters,

µ = 19.4TW

[2]



Fig. 9—Arrangement of sensors.

Fig. 10—Sequence for experiments.



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 thepre-compression. In the present study, a close and ground endedhelical spring, made of spring material SS302, having a wire di-ameter (d) = 6 mm, and a total number of turns (N) = 5, hasbeen 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 equalto 426 N on the seal interface. In the case of the X-assembly themean diameter of spring = 100 mm, free length = 132, and pre-compression = 86 mm has been used. Therefore, the spring applies321 N of the axial load to close the shaft on the seal surface. Onsubstituting the values of W and T, the coefficient of the frictionfor 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 ofthe Y-assembly; therefore, the coefficient of friction for the sin-gle seal surface will halve. In other words, µ60Y for single surface= 0.215, and µ120Y for single surface will be 0.25. These valuesof the coefficient of friction for carbon-graphite versus stainlesssteel under dry running conditions are reasonable (Jones (9)).Hence, the torque sensor provides a reliable torque resisted by sealinterfaces.

Figure 14 illustrates the torque variation in the X-assemblymisaligned by 0.014 radians. The values of the maximum, av-erage, and minimum torques equal to 7.2 N.m, 6.0 N.m, and5.1 N.m, respectively, have been observed in the misaligned X-assembly operating at 60 rpm. To compare these results andimplement a trend analysis technique, a synchronous time av-eraging technique may be useful. To employ this technique inthe present study, data obtained from fiber-optic displacementsensor (Philtec, Type RC-90) have been used. Figure 15 demon-strates the displacement and torque data. Sharp peaks of displace-ment readings indicate the end of the previous cycle and start ofthe new cycle. Using these data, a trend analysis can easily bemade.

The trend analyses for X- and Y-assemblies have been plottedin Figs. 16 and 17, respectively. Figure 16 illustrates an increase inthe torque fluctuation with an increase in the time duration of themisaligned dry condition. This torque fluctuation is an indicationof seal wear. Figure 17 notifies the torque fluctuations of the Y-assembly. One major difference between the torque behavior of Xand Y observed from Figs. 16 and 17 is the progressive flattening ofthe peak torque of Y-assembly. This indicates the progressive wearof the seal. Finally, the curve for t = 55 min in Fig. 17 demonstrates



Fig. 12—Clearance fit between shaft 1 and shaft 2 of Y-assembly.

the major variation in torque behavior, which is an indication ofseal failure.

DISCUSSION AND RECOMMENDATIONS

In the present article, the main emphasis has been given toseal failure caused by misalignment. To online monitor the fail-ure of the mechanical seal, three types of sensors—accelerometer,torque sensor, and displacement sensors—have been used. Theobserved accelerometer readings do not provide any conclusivestudy. One possible reason for such failure is the location ofthe sensor. The accelerometer was mounted on the stationaryhousing at one particular location, while the Y-assembly has four(two primary and two secondary) seal interfaces. Displacementreadings from proximity sensors also could not provide conclusive

Fig. 13—Axially loaded mechanical seal.

data. One possible reason is the mixing of circumferential and ax-ial displacements. Due to misalignment, the displacement sensorobserves the cyclic variation in the data. In addition, axial wear ofthe seal affects the displacement readings. Superposition of axialand circumferential displacement makes it difficult to conclude theseal failure. However, data from the torque sensor show a definitetrend, which can be related to seal wear. The main advantage ofthe torque sensor is that it can be connected to the rotating parts,and its magnetic pick-up provides reliable data without usage ofany slip-ring arrangement. In the present study, the torque sensorconnects the motor with the rotary joint and any geometric vari-ation at the seal interface is immediately reflected in the torquedata. Based on the torque data, shown in Figs. 16 and 17, it can bestated that X-assembly experiences mild wear even in misaligneddry conditions, while seals of Y-assemblies undergo thorough se-vere wear and breakage under a dry misaligned condition. Withsuch intuitions, the assemblies of Y and X were opened. The pho-tographs of seals are shown in Figs. 18 and 19. A photograph of theshaft in Fig. 18 indicates polishing wear, which probably occurreddue to wear debris of carbon graphite seals. The breakage of thesecondary seal and the pitting of the primary seal is also shownin Fig. 19. These photographs validate the seal failure estimatedfrom torque data of Fig. 17. Based on the photographs shown inFig. 18, it can be said that the present study illustrates the possi-bility of seal failure of Y-assemblies within 3 h of operation. Suchfailures generally occur under dry and misaligned conditions. Thedry conditions are enforced at the start and at the shutdown of thedryer by process industries. Therefore, avoiding the dry conditionis almost impossible; however, the time duration of such dry con-ditions can be minimized to enhance the seal life. Changing sealmaterial (Blau and Martin (10)) (such as silicon nitride or siliconcarbide) in place of carbon graphite may reduce the wear rate indry conditions. However, the present research indicates misalign-ment as one of the factors affecting the wear rate of the seal. There-fore, the major recommendation of the present research work isto minimize the misalignment as far as possible. As the presentdesign of the Y seal assembly uses two shafts having a clearancefit, the misalignment is inherent in the seal assembly. Introducing a



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.



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.



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 correspond-ingly the design of the Y-assembly will provide better seal life.

Figure 19 shows mild wear scars on the primary seal of X-assembly. These wear marks indicate the misalignment operationexperienced by the seal interface. This photograph validates themild progressive wear concluded by the torque data plotted inFig. 16. Quantification of wear loss is performed by measuring theweight of the seal rings using a weighing machine (CP 423 S, Sarto-

TABLE 1—WEAR LOSS DATA FOR PRIMARY AND SECONDARY SEALS

TestNumber

Typeof

Joint

Typeof

Seal

Numberof

Cycles

SlidingDistance

in mm

WeightLossin mg

1 Y PM 30,322 6,906,304.6 101SM 5,691,747.5 —PB 6,906,304.6 97SB 5,691,747.5 —

2 Y PM 30,590 6,967,345.7 76SM 5,742,053.8 21PB 6,967,345.7 77SB 5,742,053.8 20

3 Y PM 30,368 6,916,781.8 69SM 5,700,382.1 19PB 6,916,781.8 70SB 5,700,382.1 22

4 X PM 31,317 7,132,931.2 28SM 5,878,519.1 —

5 X PM 34,755 7,915,988.9 33SM 6,523,866.6 26

6 X PM 36,703 8,359,676 32SM 6,889,526 23

rius, Goettingen, Germany) of 420 g capacity with an accuracy of0.001 g. Wear loss data are listed in Table 1. In this table, PM standsfor “primary seal motor side,” SM stands for “secondary seal mo-tor side,” PB stands for “primary seal boiler side,” and SB standsfor “secondary seal boiler side.” The sliding distance has been cal-culated 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 ex-perience a high wear rate compared to the seals of the X-assembly.

CONCLUSIONS

The present research work is aimed at the development of atest setup that emulates the conditions dealt with by mechanicalface seals. Two commonly used rotary joints containing face sealshave been investigated. The conclusions of the present researchwork are

� A torque sensor, with an appropriate data acquisition and anal-ysis system, predicts the seal failure with authenticity.

� Misalignment is the root cause of failure of face seals used inY-assemblies.

� The developed test setup can predict the failure of seals undervarious rotational speeds, lubrication mechanisms, and springloads.

ACKNOWLEDGEMENT

The authors wish to express their appreciation to Forbes Mar-shall 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 Boil-

ing Interface in a Misaligned, Two-Phase, Mechanical Seal—A QualitativeStudy,” Tribology Transactions, 39(4), pp 922–928.



(3) Etsion, I. and Michael, O. (1994), “Enhancing Sealing and Dynamic Per-formance with Partially Porous Mechanical Face Seals,” Tribology Trans-actions, 37(4), pp 701–710.

(4) Anderson, W. B., Jarzynski, J. and Salant, R.F. (2001), “A Condition Monitorfor Liquid Lubricated Mechanical Seals,” Tribology Transactions, 44(3), pp479–483.

(5) Sehnal, J., Sedy, J., Zobens, A. and Etsion, I. (1983), “Performance of theConed-Face End Seal with Regard to Energy Conservation,” TribologyTransactions, 26(4), pp 415–429.

(6) Choy, F.K., Polyshchuk, V., Braun, M.J. and Lu, M. (1999), “A New Methodin Acoustic Health Monitoring in Mechanical Seals,” Tribology Transac-tions, 42(1), pp 46–52.

(7) Doust, T.G. and Parmar, A. (1986), “An Experimental and TheoreticalStudy of Pressure and Thermal Distortions in a Mechanical Seal,” TribologyTransactions, 29(2), pp 151–159.

(8) Shigley, J.E. and Mischke, C.R. (2003), Mechanical Engineering Design (inSI Units), 6th ed., Tata McGraw Hill, India.

(9) Jones, G.A. (2004), “On the Tribological Behaviour of Mechanical Seal FaceMaterials in Dry Line Contact, Part I. Mechanical Carbon,” Wear, 256, pp415–432.

(10) Blau, P.J. and Martin, R.L. (1994), “Friction and Wear of Carbon-GraphiteMaterials against Metal and Ceramic Counterfaces,” Tribology Interna-tional, 27(6), pp 413–422.


design and development of a test setup for online wear...

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