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Experimental investigation of orbiting thrust bearing using wide and shallow circular pockets.


A thrust bearing surface undergoing an orbital motion is usually encountered in a scroll compressor. In this case the stationary surface is part of the scroll compressor crank case made of gray cast iron ASTM 35 grade and the mating surface is on the orbiting scroll. Orbiting scrolls can be made grey cast iron, ductile or nodular iron or steel. Steel is expensive and usually offers little strength advantage over ductile iron. Grey iron is usually satisfactory and is lowest in cost. The thrust bearing in a scroll compressor is subjected to an average thrust load of about 1150 lb (5118 N) and speeds of about 3600 rpm (60) Hz in North American market and 3000 rpm (50) Hz in European and Asian continent.

Hydrodynamic thrust bearings are usually annular pad bearings in which one of the moving surfaces rotates relative to the other mating surface. Often there will be some type of groove pattern (e.g. radial, spiral, or circumferential) on the surface of the rotational thrust bearing to facilitate hydrodynamic pressure generation in the lubricant film and aid in lubricant transport through the groove. It has been known for sometime (Razzaque and Kato, 1999) that these grooved features, in the proper size and arrangement, provide an effective lubrication and load support mechanism. In this experimental investigation for a thrust slide bearing undergoing an orbiting motion it is illustrated that wide and shallow circular pockets arranged around the periphery of the fixed thrust surface will shift the lubrication regime from boundary to full film.

Vaidya and Sadeghi (2008) illustrated analytically that in an orbiting thrust bearing usually encountered in scroll compressors, wide and shallow circular pockets improve thrust bearing lubrication performance as compared to radial circular grooves. Yu and Sadeghi (2001) showed analytically an effective groove method of load support to liquid-lubricated thrust washers. They developed a computational model to solve for the polar-coordinate Reynolds equation using the finite volume approach. In the case of a liquid lubricant, the geometry of the groove results in cavitation, a good discussion of which is presented by Broman (2001) as applied to spiral-groove bearings. Thus, in order to ensure mass continuity, cavitation was taken into account in Yu and Sadeghi's computational model. They found that grooved thrust washers can support significant loads for rotational motion.

Razzaque and Kato (1999) showed the effect of groove inclination on the hydrodynamic behavior of wet clutches. They also demonstrated the effect of inertia in the hydrodynamic performance of wet clutches. Tian et al. (1989) studied boundary lubrication considering ploughing effects; they demonstrated that undulated titanium surfaces would considerably reduce friction coefficient and wear. Andriy et al. (2004) showed that use of dimples expands speed load parameters for hydrodynamic lubrication. The dimensions of the dimples quoted in their studies were 0.002-0.03 inches (0.05-0.76 mm) deep and 0.001-0.02 inches (0.0254-0.5) wide.

Kulkarni (1990) proposed an approach to design the inner and outer radii of thrust bearing to take into account axial load as well as any twisting moment. Tatsuya et al. (2004) in their theoretical study provided an explanation for the existing lubrication condition at the thrust surface by taking into account wedge formation between the thrust surfaces caused due to the elastic deformation of thrust plate under large loads. This elastic deformation cannot be controlled and hence cannot be used to improve the existing performance. Noriaki et al (2004) in their experimental study showed improvement in the performance of thrust slide bearing by using the pressure difference between the interior and exterior of the thrust bearing. Wang et al. (2002) studied the effect of micro pores formed by laser texturing on a SIC surface.

Computational models for smooth surfaces have demonstrated that by properly sizing the circular pockets on the bearing surface, the shift in the operation of thrust slide bearing (thrust bearing undergoing orbiting motion) from boundary to hydrodynamic full film lubrication regime can be effectively achieved. In this experimental work, a test rig was designed and developed to enable the direct measurement of the friction, in a liquid-lubricated orbiting thrust slide bearing under a constant loading condition. This study demonstrates that properly dimensioned macro circular pockets of about 0.002 to 0.003 inch (0.5 to 0.076 mm) deep and 0.3 to 0.36 inch (7.62 to 9.14 mm) wide on thrust washers can support large loads (700 lbs/3115 N) at about 3000 rpm with significant reduction in friction coefficient.


Several orbiting thrust washer specimens were investigated in the orbiting thrust washer tester (OTWT) designed and developed for this investigation. Figure 1 and 2 illustrates orbiting thrust washer tester (OTWT) designed and developed to measure friction at the thrust interface of orbiting scroll and crankcase under fully flooded conditions. OTWT was used to evaluate the effects of various surface features (circular pockets, grooves, etc.) on friction in orbiting thrust bearing. The design of OTWT is based on stacked plate concept, wherein various sections are supported by hollow cylindrical posts at each corner and each of the compartments accommodates various sub systems of the test rig. Four threaded rods are inserted through the hollow posts along the entire length of the machine and keep all the shelves aligned and assembled together. This entire stacked arrangement is mounted on a steel base (table) that also serves as the housing for the drive (motor) controller. This arrangement makes the design modular; a set of distinct subassemblies' viz. motor, coupling, test chamber, loading mechanism and drive controller can be developed independently and then put together. The change in any of the compartments does not affect the design as the interfaces between the compartments are simple. Simple interfaces also make it easier to reuse components in different circumstances thus reducing the time and cost for design and developing the test apparatus.



Figure 1 depicts the OTWT. A 30 Amps 230 V AC continuous current variable speed and torque servo motor is used in velocity control mode to operate the test rig. The motor is operated by a 200 Hz digital drive that was interfaced with a computer through a RS232 (serial port) communication port. The controller requires a special software interface provided by the motor manufacturer and needs to be operated on COM 1 port (computer port for communication) with 9600 baud rate. Power is transmitted from the motor to the crankshaft that drives the orbiting scroll through a torsionally flexible crown pin coupling located in the top compartment of the test rig.

The crankshaft (not shown) driving the scroll is kept aligned using a preloaded deep groove ball bearing on the top and a needle roller bearing in the bottom. The needle roller bearing is housed in the compressor crank case. This paired arrangement of bearings isolates the motor from the thrust load applied to the test specimen. The crankshaft is dynamically balanced using counterweights such that its mass center is aligned with the central axis of the crankshaft.

Figure 2 illustrates the test chamber which is a cylindrical vessel open at one end and serves as the lubricant reservoir. It is provided with a glass window in order to monitor the oil level inside the test chamber. The oil level should be sufficient to submerge the entire contact area, thus providing a fully flooded lubrication condition. Test specimen is mounted in the test chamber with the help of four single set screws and entire assembled test fixture is mounted on a force transducer that is used for measuring frictional forces and moments exerted on the test specimen by the orbiting scroll. The location of test specimen below the orbiting scroll ensures that only frictional forces between the scroll and test specimen are recorded by the load transducer and other forces generated in the assembly above the scroll during operation are filtered.

Force transducer (MC3A) located directly below the test chamber is capable of simultaneously measuring three orthogonal force components along X, Y and Z axes and three moments. Data is acquired through six channel digital amplifier at a rate of 16000 samples per second for each channel. Load cell is mounted on a vertically moving platform guided by linear bearing running on shafts located at each corner of the sensor plate. The loading mechanism is a lever arm with a 12:1 leverage. The lever arm is subjected to a maximum static load of 780 lb (3471 N) and is made of 'I' section to resist bending moment. The weight of the test chamber including the weight of the specimen and oil is about 5 lb before the start of test.

The crankcase of the scroll compressor is one of the two mating thrust surfaces; it was modified to fit into the OTWT. This surface is made of a special malleable alloy of cast iron. Figure 3 shows the compressor crankcase and orbiting scroll that were used as base line specimens to bench mark the performance of the other designs. Theoretically the flat unmodified thrust specimens should not support any hydrodynamic load except through relatively weak and unstable mechanisms (e.g. Thermal distortion) and will operate in the boundary lubrication regime. In the case of rotational thrust washers it was shown by Yu and Sadeghi (2001) that wide and shallow uniformly distributed radial grooves (circular and triangular) provide excellent hydrodynamic load support. In order to investigate whether the groove designs for rotational washers can be extended to orbiting thrust bearing circular grooves as shown in Figure 4 and 5 were experimentally evaluated to determine the lubrication characteristics (regimes).




In order to obtain the optimum size, shape and pattern of the circular pockets on the thrust surface, parametric study was performed on circular pockets by varying number of groove/pockets in the radial direction (Np), number of grooves/pockets (Ng), groove/pocket width ([G.sub.w]) and groove/pocket depth in inches ([G.sub.d]). Table 1 provides a comprehensive list of specimens used in this investigation (D through K). Figure 6 illustrates specimens (D and E) with one and two rows of circular pockets along the width of the thrust bearing respectively. In order to manufacture these specimens a custom ball end-mill was used. The average surface finish achieved inside the circular pockets is [+ or -]0.00003 inch ([+ or -]0.00762 mm) with round circular milling marks and average surface finish on the land in between the pockets is [+ or -]0.00003 inch ([+ or -]0.00762 mm) characterized by cross grinding marks. Figure 7 depicts specimens (F, G and H) with varying width of the pockets ([G.sub.w]). Further circular pockets with varying depth ([G.sub.d]) (not shown but referenced as I, J and K in Table 1) were tested to obtain optimum depth that would enhance the lubrication at the thrust interface. All the specimens were drilled with holes of 0.08 inches (2.03 mm) in diameter and 0.2 inches (5.0 mm) deep just below the thrust surface in order to accommodate thermocouple bead for temperature measurement.
Table 1. Summary of Feature Depth ([G.sub.d]) and Width ([G.sub.w]) for
Thrust Specimens

              Specimen Geometries

     [G.sub.w] (in)  [G.sub.d] [+ or -] 0.0005 (in)

(A)      0.04                    0.02

(B)      0.02                    0.01

(C)      0.984                  0.0008

(D)       0.2                    0.002

(E)       0.15                   0.002

(F)       0.1                    0.003

(G)       0.2                    0.003

(H)      0.36                    0.003

(I)       0.3                    0.002

(J)       0.3                    0.002

(K)       0.36                   0.002



Tests were also carried out by changing surface finish (special grinding operation was performed on scrolls as opposed to regular precision turning) of the scrolls (moving surface) in order to investigate the effects of surface finish on the performance of the scroll compressor thrust bearing.


Test specimen is mounted in the test chamber and thermocouple bead (K-Type) is inserted in a hole drilled just below the stationary thrust surface. The load cell is set to zero using hardware zero function on the amplifier. Performing hardware zero nullifies the effect of chamber weight, residual forces induced in the load cell due to tightening of the screws and any misalignment. The hardware zero command in amplifier also takes care of any electrical noise. The lever arm is loaded by attaching dead weights at its end, it should be noted that the maximum normal load carrying capacity of load transducer is 1000 lb therefore maximum normal load applied was limited to this threshold limit. Shake down for every new specimen is achieved by operating it at 1000 rpm and 480 lb (2136 N) of normal load for 10 to 15 minutes. This ensures that initial wear on the thrust surfaces do not affect the experimental results.

In order to eliminate the effect of temperature on lubricant and therefore friction coefficient, data is acquired for one second. The data acquired for one second is sufficient to capture the cyclic variation of friction forces in one rotation of crankshaft because the motor was operated at a frequency of 50 Hz and data was acquired at 16000 samples/sec which results in 320 data sets per cycle. The temperature reference for the test is 76.1 F (24.5 C). The test is started at 74.3 F (23 C) so that the transient squeeze film effects diminish and the data acquisition is started when the reference temperature is reached. Tests were carried out for three different normal loads 480, 660 and 780 lb (2136, 2937 and 3471 N) respectively and speed ranging from 10 to 3000 rpm in steps of 100 rpm. More than three trials were conducted for each specimen in order to establish the scatter in the experimental data. Figures 8 through 11 illustrate stribeck variation of coefficient of friction, [mu] versus speed, N for various specimens for a normal load of 780 lbs (3471 N).






Figure 8 shows the variation of friction coefficient, ([mu]) versus speed for a thrust load of 780 lb (3471 N). Turned specimens (base specimens) in general operated in the boundary lubrication regime (characterized by a constant friction coefficient with varying speed) as expected because parallel surfaces theoretically do not generate hydrodynamic pressure. Ground specimens were also essentially two flat parallel plates with smooth surface finish and theoretically parallel surfaces should not support any load. However these specimens were manufactured by moving grinding wheel spirally on the surface (Due to specific shape of the orbiting scroll regular flat grinder cannot be used). These microscopic spiral marks on the ground surface generates very small geometric wedge effect causing it to operate in inconsistent hydrodynamic regime. Also the sample to sample performance variation of ground specimens was more than 40% which is an indication that control on producing consistent surface patterns generated due to spiral grinding is limited. Of the circular pocket geometries used for this investigation, the single row of pockets (Np = 1) having pocket width ([G.sub.d]) and depth ([G.sub.w]) of 0.3 and 0.003 inches (7.62 and 0.0762 mm) respectively provided the best lubrication performance. The repeatability of the experiments for the pocketed specimens was around [+ or -]5%. On the other hand circular grooves although performed in the hydrodynamic regime, their performance was significantly less as compared to circular pocketed specimens because circular grooves are uniformly distributed on the thrust surface and provide geometric wedge only in the tangential direction. In the case of orbiting thrust bearing couette velocity in the radial direction does not contribute to pressure generation as there in an absence of geometric wedge in radial direction for the circular grooves. On the other hand radial component of couette velocity leads to side leakage through the grooves leading to decrease in load support or increase in friction coefficient. Thus it can be concluded that for orbiting condition circular pockets are more effective in producing load support and shift the lubrication from boundary to full film lubrication.


The effects of variation in pocket depth ([G.sub.d]), width ([G.sub.w]) and number of pockets per width (Np) were also studied. This parametric study was carried out to obtain an optimum configuration of circular pockets on the thrust surface. Figure 9 depicts the effect of number of circular pockets per width on friction coefficient for normal loads of 780 lbs (3471 N). Results presented here are for two specimens with one and two rows of circular pockets along the width. The design of circular pockets is such that the total thrust surface area occupied by the circular pockets is kept constant. The results obtained indicate that one row of pockets per width (Np=1) yields the least friction coefficient in the hydrodynamic regime. This is due to the fact that there is interaction between pressure spikes developed around the edge of individual pockets as number of pockets per width increases as discussed by Vaidya and Sadeghi (2008).


After observing that pockets with Np=1 provides the best lubrication performance, the effect of pocket depth on friction was examined. Figure 10 illustrates the variation of friction coefficient versus speed for various pocket depths ([G.sub.d]). In this study, the pocket width ([G.sub.w]) was held constant at 0.3 inch (7.62 mm) and each specimen was run for three different loads. The depth of pockets was varied from 0.002 to 0.004 inch (0.05 to 0.1 mm). It was observed that pocketed specimens with depth of 0.002 and 0.003 inch (0.05 and 0.076 mm) performed nearly the same until the speed reached 2500 rpm; however with further increase in speed, the friction performance of pockets with 0.003 inch (0.076 mm) depth showed slightly better friction characteristics. However, the difference in the performance was almost within the repeatability of the two specimens. The reason for this type of congruent behavior is that circular pockets were manufactured with ball milling operation having tolerance of [+ or -]0.0005 inches ([+ or -]0.0127 mm) and require dexterous control on the manufacturing process and even then the tolerance is difficult to achieve. It can also be observed from Figure 10 that as the depth of pocket is increased to 0.004 inches (0.1 mm) friction performance starts to deteriorate. Vaidya and Sadeghi (2008) illustrated analytically that friction coefficient; is a function of pressure gradients and as the depth of pockets increases, steep pressure spikes are observed at the edge of the pockets resulting in increased friction. From the above discussion it can be concluded that a depth range of 0.002 to 0.003 inches (0.05 to 0.076 mm) for pockets was optimum for this investigation.


After determining optimum pocket depth, effect of groove width, ([G.sub.w]) on friction was examined. The width of the pockets was varied from 0.1 inches to 0.36 inches (2.54 mm to 9.14 mm) with Np=1 and pocket depth [G.sub.d] = 0.003 inches (0.076 mm). Figure 11 depicts the effect of width on friction coefficient, ([mu]) for different pocket widths. Again the tests were carried out on three load sets. It was observed that the width of the pockets significantly dominated the hydrodynamic lubrication and the larger the width the smaller is the friction due to greater geometric wedge. The upper limit on the width of the pockets is limited by the size of the thrust surface. In our case the best pocket width turns out to be 0.2 to 0.36 inches (5.0 to 9.14 mm) and ball milling or chemical etching are some available machining options.


In this investigation, the effect of various surface modifications (i.e. groove patterns, circular grooves and circular pockets) on the friction performance of scroll compressor thrust bearing was evaluated. The effect surface finish on the orbiting scroll obtained by spiral grinding as compared to surface finish obtained by precision turning was also examined. The frictional performance of modified specimens was compared to that of base specimens (precision turned specimens without any surface features). In order to ensure that performance was consistent and experimental scatter was kept to minimum each test was run three times and it was observed that average maximum percentage error was about 5%. The parametric study revealed that single row of circular pockets of about 0.3 to 0.36 inch (7.62 to 9.14 mm) and depth of about 0.002 to 0.003 inch (0.05 to 0.076 mm) yields the best lubrication performance for the scroll compressor used in this investigation. Though the study has been performed for orbiting motion occurring in a scroll compressor it should be noted that the concept of macro circular pockets on the stationary bearing surfaces can be extended to any bearing application involving one stationary surface and second surface undergoing a complex motion.


The authors would like to express their deepest appreciations to the Tecumseh Company for their support of this research study.


Ng = Number of grooves/pockets.

[G.sub.d] = Pocket/groove depth in inches.

Np = Number of pockets in radial direction.

[G.sub.w] = Pocket/groove width

OTWT = Orbiting thrust washer tester

MC3A = transducer

N = Speed of crankshaft in rpm

[mu] = Coefficient of friction

[eta] = Viscosity of oil

A = Area of thrust surface

[omega] = Angular velocity in rad/sec, 60 2[pi]N/60

[F.sub.z] = Normal thrust load

[eta][omega]A/[F.sub.z] = Sommerfelds number


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Farshid Sadeghi, PhD

Amit Vaidya is a mechanical engineer, Syracuse, NY. Farshid Sadeghi is a professor in the School of Mechanical Engineering, Purdue University, West Lafayette, IN.
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Author:Vaidya, Amit; Sadeghi, Farshid
Publication:ASHRAE Transactions
Article Type:Report
Geographic Code:1USA
Date:Jul 1, 2009
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