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Experimental study of the lubrication mechanism for thrust slide bearings in scroll compressors.

INTRODUCTION

Scroll compressors are widely used in room air-conditioners. In these compressors, a thrust bearing is used to maintain the stable orbiting motion of the orbiting thrust plate as it is firmly pressed against the fixed thrust plate. There are several types of thrust bearings, including the slide-bearing type, the ball-bearing type, etc. Among these, the thrust slide bearing is used most commonly because of its low mechanical friction losses and low noise generation. In a thrust slide bearing, the orbiting flat thrust plate is firmly pressed against the fixed thrust plate. The thrust slide bearing supports a large thrust force and does not require high pressure lubrication from a dedicated high-pressure oil pump. Despite the lack of a high-pressure oil pump, the thrust slide bearing generally does not exhibit any significant lubrication problems that result in seizure of the sliding surfaces. The thrust slide bearing, in fact, exhibits better performance than might be expected.

In the thrust slide bearing, when an oil particle becomes trapped in the sliding space, it can, in theory, never escape due to the moving scroll's orbital motion. The trapped oil particle is then heated by the shear stress. This trapping of the oil particles coarsely explains why the thrust slide bearing does not experience seizure. However, this explanation is insufficient to explain in detail why thrust slide bearings exhibit reduced mechanical friction loss. Kulkarni (1990a, 1990b) presented theoretical studies of thrust slide bearings in scroll compressors while Nishiwaki et al. (1996) presented an experimental study. Neither of these approaches, however, offered a viable explanation for the good performance of thrust slide bearings. The performance of scroll compressors has been analyzed through computer-simulations by Ishii et al. (1992, 1994, 1996a, 1996b, 2000a, 2000b, 2002a, 2002b) and Hiwata et al. (2002) to determine the optimal design for efficiency. In these studies, the fundamentals of the refrigerant leakage flow were examined in detail, but insight into the frictional state of the thrust slide bearing has not been successfully documented. Understanding of the lubrication mechanism for thrust slide bearings would clearly result in improved scroll compressor performance.

This study contributes lubrication test results for the thrust slide bearing of a scroll compressor to determine the key factor producing the good lubrication performance. The thrust slide bearing was replaced by a simple model composed of a cylindrical thrust plate representing the orbiting scroll and a flat thrust plate representing the fixed scroll (Ishii et al. 2004). The present study focuses on especially the pressure difference across the thrust slide bearing. The outside region is pressurized at an intermediate pressure between the discharge and suction pressures, while the inside region is at the suction pressure. Due to this pressure difference, the oil at the thrust slide bearing flows from the outside region into the inside region. Further, a wedge of fluid is formed between the friction surfaces. In order to examine these pressure difference effects on the lubrication performance, the thrust slide bearing model was submerged in refrigerant oil pressurized in a closed vessel with R-22, and the inner region of the cylindrical thrust plate was open to atmospheric pressure. The lubrication tests were conducted with pressure differences between 0 and 1.0 MPa. In addition, the thrust force and the orbital speed were varied up to maximum of 9200 N and 3600 rpm, respectively. First, the frictional force and the temperature at the sliding surface were carefully measured. Second, after lubrication tests, the friction surfaces of the thrust slide bearing were carefully examined to determine the state of wear. Subsequently, finite element method (FEM) analysis of the fixed thrust plate was undertaken to determine its elastic deformation as input into the calculation of the wedge angle characteristics as a function of the applied pressure difference. Finally, measurement of the strain in the fixed thrust plate were conducted to validate the FEM analysis.

SCROLL COMPRESSOR AND THE THRUST SLIDE BEARING

A cross-sectional view of a high-pressure type scroll compressor is shown in Figure 1a. The compression mechanism is located in the upper portion of the closed vessel and is driven by the motor in the center portion of the closed vessel. As shown in Figure 1b, the compressed gas is discharged upward through the center port of the compression mechanism. After one downward pass through the motor rotor, the compressed gas then flows upward and passes over the outside of the motor stator. The refrigerant oil collected at the bottom of the closed vessel is pumped upward by an oil pump attached to the bottom end of the crankshaft through a vertical channel inside the crankshaft.

[FIGURE 1 OMITTED]

As shown in Figure 1b, the oil pumped to the top end of the crankshaft lubricates the eccentric bearing and forces the orbiting thrust plate and the tip seal up against the orbiting wrap. Furthermore, the oil passes through a needle valve into the intermediate pressure space where the pressure is adjusted to an optimal intermediate pressure by the control valve that releases the lubricant to the suction chamber. The intermediate pressure presses the orbiting thrust plate up against the fixed thrust plate. Thus, the orbiting thrust plate slides over the fixed thrust plate and forms a thrust slide bearing, which is indicated by the shaded area in Figure 2a.

[FIGURE 2 OMITTED]

It should be noted here that the outer region of the thrust slide bearing is pressurized at the intermediate pressure, and the inner region is at the lower suction pressure, thereby establishing a pressure difference. This pressure difference produces an oil flow from the outer region into the inner region of the thrust slide bearing. In addition, an oil wedge is formed between the friction surfaces and yields the exceptionally good lubrication performance of the thrust slide bearing. The major purpose of the present study is to unravel the effect of the pressure difference on the lubrication performance of the thrust slide bearing.

TRIBO TESTER FOR LUBRICATION TESTS OF A THRUST SLIDE BEARING

For convenience, a simplified model was used for the lubrication tests of the thrust slide bearing. The orbiting scroll thrust plate was replaced by a cylindrical thrust plate with inner radius [r.sub.i] = 37.85 mm and outer radius [r.sub.o] = 65 mm. In the model, this cylindrical thrust plate was held stationary and mounted above a flat thrust plate that was driven by a motor to produce orbiting motion, as shown in Figure 2b. The test pieces of the thrust slide bearing have the specifications shown in Table 1. The material is aluminum alloy for the fixed cylindrical thrust plate and cast iron for the orbiting thrust plate. The initial surface roughness is [R.sub.a] = 0.7 and 3.0 mm for the fixed and orbiting thrust plates, respectively.
Table 1. Specifications of Thrust Slide Bearing Model

                         Material                   Surface Roughness
                                                    [R.sub.a], [mu]m

Fixed cylindrical  Aluminum alloy (Al 90%, Si 10%)          0.7
plate

Orbiting plate     Cast iron (C 0.25%)                      3.0


To properly simulate the sliding thrust bearing, the external bearing surface must be in pressurized oil and the internal region must be maintained at some preselected lower pressure. In addition, the fixed cylindrical thrust plate should be loaded by an applied axial force that acts in parallel with the pressure thrust force. The specially designed tribo tester, a closed containment vessel for the model bearing, is shown in Figure 3. It permits simulated thrust slide bearing lubrication testing in a carefully controlled environment.

[FIGURE 3 OMITTED]

The thrust slide bearing model is submerged in refrigerant oil VG-56 for refrigerant R-22. The R-22 is stored in a tank outside the closed pressure vessel. The tank is heated and R-22 gas is fed into the closed pressure vessel to adjust the internal ambient pressure. The pressure in the internal space beneath the fixed thrust plate is regulated through a capillary tube and valve vented to atmospheric pressure outside the closed pressure vessel. The control valve at the end of the capillary tube is used to adjust the pressure of the internal bearing region. The fixed thrust plate is axially loaded by the axial load shaft and a spring in the axial load cylinder. This axial spring force, represented by [F.sub.s], can be controlled from outside of the closed pressure vessel. This thrust force, [F.sub.s], was measured with strain gauges (KYOWA: KFG-2-120-C1-11) mounted on the leaf spring and oriented along its axis. The strain gages were connected to a dynamic strain amplifier (KYOWA DPM-6H). In addition, the fixed thrust plate is pressed down by the pressure force, represented by [F.sub.p], due to the pressure difference between the external and internal regions of the thrust slide bearing.

The orbiting thrust plate is driven by a motor located outside the closed pressure vessel. The orbiting thrust plate tries to drag the fixed thrust plate. The drag force exerted on the fixed plate is the frictional force, represented by [F.sub.f], between the fixed and orbiting thrust plates. The fixed thrust plate is connected by a pivot coupling to the bottom end of the axial load shaft. Therefore, the frictional force, [F.sub.f], can be accurately measured by the strain gauges on the axial load shaft. The crankshaft rotation was measured by a rotary encoder. The friction surface temperature, represented by [T.sub.f], was measured by a Type T thermocouple inserted through the fixed bearing plate to be in close proximity to the friction surface, as shown in Figure 3, and was monitored by a digital thermometer (Yokogawa Electric Corporation: TYPE 2572). A connection diagram for the sensors and instrumentation is shown in Figure 4. Measured axial spring force [F.sub.s], friction force [F.sub.f], friction surface temperature [T.sub.f], and crankshaft rotational pulse were recorded using a digital data recorder (TEAC: LX-10) and were monitored using a PC (TOSHIBA: Dynabook G7/U24PDDW).

[FIGURE 4 OMITTED]

LUBRICATION TEST RESULTS

The major parameters for the present lubrication tests are given in Table 2. The internal pressure of the tribo tester was maintained at 1.0 MPa. Initially, the pressure difference control valve was completely closed, which resulted in the internal bearing pressure of 1.0 MPa and, thus, a pressure difference across the bearing (i.e., [DELTA]p) of zero. With this pressure difference and with an orbital radius of 3.0 mm, lubrication tests were conducted with the axial spring force, [F.sub.s], at 600 N and with orbital speeds varying from 300 to 3600 rpm. Representative drag force and rotation pulse data from this first set of lubrication tests at [DELTA]p = 0 MPa with N = 1800 rpm and [F.sub.s] = 600 N are shown in Figures 5a and 5b, where the portion with the dense pulses was recorded during one rotation of the crankshaft. The observed contrast between the drag force histogram and the angular pulse position data indicates that the drag force is synchronized with the orbital motion of the thrust plate. The direction of drag force rotates with the orbital motion of the thrust plate. Therefore, the magnitude of the frictional force, [F.sub.f], can be determined from the spectral peak in the power spectrum at 30 Hz, which results in 25.5 N, as shown in Figure 5c.

[FIGURE 5 OMITTED]
Table 2. Major Parameters for the Lubrication Tests

Pressure difference [DELTA]p, MPa                               0 ~ 1.0
Axial spring force [F.sub.s], N                                   600
Gas thrust force [F.sub.p], N                                  0 ~ 8887
Resultant thrust force [F.sub.t](= [F.sub.s] + [F.sub.p]), N  600 ~ 9487
Orbiting speed N, rpm                                         300 ~ 3600
Orbiting radius, mm                                               3.0
Refrigerant oil                                                 VG-56
Refrigerant                                                      R-22


Representative results of lubrication tests at [DELTA]p = 0 MPa and [F.sub.t] = 600 N are shown by the open circles in Figure 6, where the frictional force, [F.sub.f], the friction coefficient, [mu] (= [F.sub.f]/[F.sub.s]), and the friction surface temperature, [T.sub.f], are presented versus the orbital speed. [F.sub.f] takes on a value of 67 N at 300 rpm and decreases gradually with increasing orbital speed N to a value of about 17 N at 3600 rpm. The data points are linked by a smooth dashed line. In Figure 6b, the corresponding [mu] values range from 0.11 at 300 rpm and to 0.028 at 3600 rpm with a similar decrease with increasing orbital speed. This result suggests that the oil film thrust force at the sliding surface increases with increasing orbital speed and provides a better lubrication state. In contrast, the friction surface temperature, [T.sub.f], shown in Figure 6c, increases from 40[degrees]C to 64[degrees]C with increasing orbital speed. The shape of the curves in Figure 6c suggests that a sufficiently thick oil film was built up for the higher orbital speeds, resulting in no substantial increase in heat due to friction.

[FIGURE 6 OMITTED]

Second, the pressure difference control valve of the capillary tube was adjusted step-by-step so that the inside space pressure decreased from 1.0 to 0 MPa, that is, the pressure difference [DELTA]p increased from 0 to 1.0 MPa. Under these conditions, similar lubrication tests were conducted, which resulted in the data shown by the additional data points in Figure 6. The maximum gas thrust force, [F.sub.p], was 8887 N in addition to the axial spring force, [F.sub.s], of 600 N, which yielded a maximum resultant thrust force, [F.sub.t], of 9487 N. Including the gas pressure, the nominal bearing pressure of the friction surface ranged from 169.8 to 1081.4 kPa. In calculating the pressure thrust force, [F.sub.p], it was assumed that the pressure acting on the friction surface varied linearly from 1.0 MPa at the periphery to 0 MPa at the inner circumference, as shown in Figure 7. Since the friction surfaces of the fixed and orbiting plates were initially ground, flat plates of high accuracy, this assumed pressure distribution along the friction surface should be correct. The upward-acting pressure force was deducted from the downward-acting pressure force on the surfaces. As a result, the friction coefficient, [mu], at the thrust slide bearing can be calculated by dividing the frictional force, [F.sub.f], by the resultant thrust force, [F.sub.t] (i.e., [mu] = [F.sub.f]/[F.sub.t]).

[FIGURE 7 OMITTED]

With increased pressure difference, the resultant thrust force, [F.sub.t], increases significantly, producing an overall increase in the frictional force, [F.sub.f], as shown in Figure 6a. At the maximum pressure difference of [DELTA]p = 1.0 MPa, [F.sub.f] takes on a value of 650 N at 300 rpm and 200 N at 3600 rpm. For all pressure differences, the frictional force decreases with increasing orbital speed.

The friction coefficient, [mu], as shown in Figure 6b, decreases significantly with just a small pressure difference of [DELTA]p = 0.2 MPa. With further increasing pressure difference, [DELTA]p, the friction coefficient decreases in the lower orbital speed range (for speeds less than 600 rpm) and increases in the higher orbital speed range (greater than 600 rpm). In order to clearly display these characteristics, the same data were rearranged as a function of the pressure difference, [DELTA]p, as shown in Figure 8. Figure 8a shows the frictional force and Figure 8b shows the friction coefficient. At all orbital speeds, the friction coefficient decreases rapidly with increasing pressure difference to 0.2 MPa. With further pressure increases, the friction coefficient decreases gradually at 300 rpm, while it remains constant for speeds from 600 to 2100 rpm and gradually increases for orbital speeds beyond 2100 rpm. However, the friction coefficient never reaches its zero pressure difference value and, even when the friction coefficient increases with increasing pressure difference, its maximum value is about 70% of the value at [DELTA]p = 0 MPa.

[FIGURE 8 OMITTED]

The reasons for these friction coefficient characteristics can be understood as follows: in the beginning, with a small pressure difference, a geometrical oil wedge is formed, which increases the hydrodynamic lifting force, which, in turn, significantly decreases the friction coefficient. However, with further increases in pressure difference beyond a certain level, the oil-film thickness decreases due to increased resultant thrust force and, hence, the friction coefficient increases instead and cancels the wedge formation effect. The effectiveness of the wedge cancellation condition varies with the orbital speed. The geometrical oil wedge formation will be further examined and explained in detail in the next section.

As shown in Figure 6c, the friction surface temperature, [T.sub.f], increases continually with increasing pressure difference [DELTA]p. However, even with the largest pressure difference of [DELTA]p = 1.0 MPa, the temperature is between 50[degrees]C and 80[degrees]C. The increase over the temperature with zero pressure difference (40[degrees]C to 64[degrees]C at [DELTA]p = 0 MPa) is relatively small, which suggests that even with higher pressures a surprisingly good lubrication state is maintained.

WEAR STATE OF FRICTION SURFACE AND WEDGE FORMATION

The friction surface of the fixed cylindrical thrust plate (aluminum alloy) after lubrication-testing for about three hours at a pressure difference of 1.0 MPa and orbital speeds ranging from 300 to 3600 rpm is shown in Figure 9. Special attention should be paid to the clear differences in the wear state along the radial direction. As seen in the enlarged photo on the right of Figure 9, the inside area indicated by A is a mirror-like plane, while the middle area indicated by B shows severe abrasive scratches describing circular patterns from the orbital motion. In the outer region, C, the circular pattern disappears. Based on these observations, the mean surface roughness of these three areas was measured, as shown in Figure 10 and Table 3, where the roughness, [R.sub.a], distinctively decreases from 0.27 [micro]m at the outside area, C, to 0.056 mm at the inside area A. These results of surface roughness suggest that a wedge was formed between the friction surfaces of the thrust slide bearing so that the inside area, A, was often rubbed and, in contrast, the outside C became a floating surface.

[FIGURE 9 OMITTED]

[FIGURE 10 OMITTED]
Table 3. Surface Roughness [R.sub.a] of Fixed Cylindrical Thrust Plate

Location                              A      B      C

Surface Roughness [R.sub.a], [mu]m  0.056  0.128  0.270


FEM SIMULATION OF WEDGE ANGLE

The wedge formation between the friction surfaces can be investigated using FEM analysis for the elastic deformation of the fixed thrust plate using Pro/MECHANICA Wildfire Release 24.8 software (PTC 2006) on a PC (Dell DIMENSION 8300; CPU: Pentium 4, 3.0 GHz; RAM: 1.0 GB; HDD: 40 GB; OS: Windows XP). The finite element model of the cylindrical thrust plate was constructed using tetrahedron solid elements. The orbiting thrust plate was modeled as a rigid plate element. In the FEM simulations, the pivot bearing port at the center of the thrust plate was constrained to permit only movements in the horizontal direction, and a contact condition was imposed between the friction surfaces. In addition, the thrust spring force, [F.sub.s], of 600 N was loaded on the pivot surface and the gas pressure difference, [DELTA]p, from 0 to 1.0 MPa (shown in Figure 7), was loaded on the thrust plate.

A representative result of FEM simulations at [F.sub.s] = 600 N and [DELTA]p = 0.3 MPa is shown in Figure 11 with a contour view. From the upper view, showing an enlarged periphery of the thrust plate, it is clear that a linear wedge is formed at the periphery of thrust plate from which a wedge angle [alpha] of 52.5x[10.sup.-6] rad was identified. Similar FEM simulations were made for a variety of pressure differences to determine the wedge angle [alpha] dependence on the pressure difference, [DELTA]p. The results are shown in Figure 12a. The wedge angle [alpha] increases linearly with the increasing pressure difference, [DELTA]p. The wedge angle varied from 20 x [10.sup.-6] rad at [DELTA]p = 0 MPa to 120 x [10.sup.-6] rad at [DELTA]p = 1.0 MPa. With the thrust bearing radial dimension of 27.15 mm, these wedge angles result in vertical clearances at the thrust plate periphery of 0.6 to 3.2 [micro]m.

[FIGURE 11 OMITTED]

[FIGURE 12 OMITTED]

In order to validate the FEM simulations, strain measurements of the cylindrical fixed thrust plate were conducted. As shown in Figure 11, the largest strain appears at position A for tension and at position B for compression. Based on these calculations, single-axis strain gauges (KYOWA: KFG-2-120-C1-11) were mounted at positions A and B in the radial direction, as shown in Figure 13. Strain measurements were made with a spring thrust force of [F.sub.s] = 600 N, pressure differences of [DELTA]p = 0 to 0.3 MPa, and orbital speeds of N = 600, 1800, and 3600 rpm. Representative real-time waveforms measured at [DELTA]p = 0.3 MPa and N = 600 rpm are shown in Figure 14, where the time histories of the thrust force, the friction force, and the strains at A and B are presented. Initially, the test was in steady state operation. Then, at about 3 s, the pressure valve was closed to release the pressure difference. The thrust spring force began to unload at about 4.2 s. The crankshaft rotation speed began to decrease at about 3 s and completely stopped at about 12 s. As shown in Figures 14c and 14d, the strain [epsilon] in the steady operation was 2.0x[10.sup.-5] at position A and -9.0x[10.sup.-5] at position B. Similar strain measurements were made for a variety of pressure differences. Measured data are plotted in Figure 12b, where the solid lines represent FEM-simulated results of the radial strains. The two data sets show good agreement with the FEM results, which suggests that the FEM-simulated wedge angle determinations were sufficiently accurate.

[FIGURE 13 OMITTED]

[FIGURE 14 OMITTED]

Thus, one can conclude that the existence of a geometrical oil wedge in the peripheral portion of the cylindrical thrust plate due to its elastic deformation has been quantitatively verified. It can be concluded that this wedge formation produced an effective hydrodynamic lifting force, which contributed directly to the observed excellent lubrication performance of thrust slide bearings.

CONCLUSIONS

Lubrication tests of the thrust slide bearing of scroll compressors were conducted in a closed vessel pressurized with refrigerant R-22 gas, focusing on the effect that pressure differences between the outside and inside spaces have upon the lubrication performance. Factors such as frictional forces, the friction coefficients, and the friction surface temperature were measured for a variety of pressure differences from 0 to 1.0 MPa and for orbital speeds from 300 to 3600 rpm. In addition, the wear state of the friction surface was carefully examined after lubrication-testing. Elastic deformation and strain estimates were obtained for the cylindrical thrust plate using FEM analysis. As a result, the outstanding lubrication performance of thrust slide bearings due to pressure differences was confirmed with the following characteristics:

1. With a small increase in pressure difference of 0.2 MPa, the friction coefficient decreases significantly despite the increase in thrust load at all orbital speeds.

2. The post-testing observation of wear state of the friction surface suggested that improvement of lubrication performance with increased pressure was caused by an oil wedge formation between the friction surfaces due to elastic deformation of the thrust plate, which produced a hydrodynamic lifting force that effectively reduced the friction force. The wedge angle increased with increasing thrust load, which resulted in improved lubrication performance, even under higher thrust load.

3. When the orbital speed was at a high value of 3600 rpm, the friction coefficient gradually increased from 0.01 to 0.02 with increasing pressure difference above 0.2 MPa. This was considered to result from decreased oil film thickness canceling the improvement due to wedge formation.

4. As the orbital speed increased, the friction coefficient significantly decreased from 0.065 to 0.105 at 300 rpm to 0.01 to 0.03 at 3600 rpm.

5. The wedge angle between the friction planes--a result of the elastic deformation due mainly to pressure loads on the orbiting scroll made of aluminum alloy--was calculated from FEM analysis to be about 120x[10.sup.-6] rad with a pressure difference of 1.0 MPa, which corresponded to a clearance of 3.2 [micro]m at the periphery of the thrust plate.

The wedge formation at the friction planes appears to be a key factor in the outstanding improvement of lubrication performance of the thrust slide bearing in scroll compressors. In addition, the quantitative results concerning the wedge angle, which were determined in this study with a simplified thrust slide bearing model, can be applied to operational scroll compressors. In order to verify the essence of the lubrication mechanism further, a theoretical analysis of lubrication of thrust slide bearings should be developed. With the establishment of such a theoretical analysis (see Oku et al. [2004] for an initial attempt), the development of an optimal performance design method will become possible for the thrust slide bearing.

NOMENCLATURE

[F.sub.f] = Friction force, N

[F.sub.p] = Gas thrust force, N

[F.sub.s] = Thrust shaft load, N

[F.sub.t] = Resultant thrust force, N

N = Orbital speed, rpm

[R.sub.a] = Surface roughness, [mu]m

[T.sub.f] = Friction surface temperature, [degrees]C

[alpha] = Wedge angle, rad

[epsilon] = Strain, dimensionless

[DELTA]p = Pressure difference of friction surface, MPa

[mu] = Friction coefficient, dimensionless

REFERENCES

Hiwata, A., Iida, N., Futagami, Y., Sawai, K., Ishii, N. 2002. Performance Investigation with Oil-injection to Compression Chambers on [CO.sub.2] Scroll Compressor, International Compressor Engineering Conference at Purdue, C18-4.

Ishii, N., Yamamoto, S., Muramatsu, S., Yamamura, M. and Takahashi, M. 1992. Optimum Combination of Parameters for High Mechanical Efficiency of a Scroll Compressor, Proceedings of International Compressor Engineering Conference at Purdue, West Lafayette, Indiana, USA, pp. 118a1-118a8.

Ishii, N., Yamamura, M., Muramatsu, S., Yamada, S. and Takahashi, M. 1994. A Study on High Mechanical Efficiency of a Scroll Compressor with Fixed Cylinder Diameter, Proceedings of International Compressor Engineering Conference at Purdue, West Lafayette, Indiana, USA, Vol.2, pp. 677-682.

Ishii, N., Sakai, M., Sano, K., Yamamoto S. and Otokura T. 1996a. A Fundamental Optimum Design for High Mechanical and Volume Efficiency of Compact Scroll Compressors, Proceedings of International Compressor Engineering Conference at Purdue, West Lafayette, Indiana, USA, Vol. II, pp.639-644.

Ishii, N., Bird, K., Sano K., Oono, M., Iwamura S. and Otokura T. 1996b. Refrigerant Leakage Flow Evaluation for Scroll Compressors, Proceedings of International Compressor Engineering Conference at Purdue, West Lafayette, Indiana, USA, Vol.II, pp. 633-638.

Ishii, N., Morita, N., Kurimoto, M., Yamamoto, S. and Sano, K. 2000a. Calculations for Compression Efficiency Caused by Heat Transfer in Compact Rotary Compressors, Proceedings of International Compressor Engineering Conference at Purdue, West Lafayette, Indiana, USA, Vol.I, pp. 467-474.

Ishii, N., Morita, N., Ono, M., Yamamoto, S. and Sano, K. 2000b. Net Efficiency Simulations of Compact Rotary Compressors for Its Optimal Performance, Proceedings of International Compressor Engineering Conference at Purdue, West Lafayette, Indiana, USA, Vol.I, pp. 475-482.

Ishii, N., Kawamura, S., Yamamoto, S., Sawai, K., Hiwata, A., Kawano, H., Ting, K.S. 2002a. Efficiency Simulations with Consideration of Heat Losses of a R410A Compact Scroll Compressor for Its Optimal Performance, International Compressor Engineering Conference at Purdue, C22-2.

Ishii, N., Kawamura, S., Yamamoto, S., Sawai, K., Hiwata, A., Kawano, H., Ting, K.S. 2002b. Efficiency Simulations of a Compact [CO.sub.2] Scroll Compressor and its Comparison with Same Cooling Capacity R410A Scroll Compressor, International Compressor Engineering Conference at Purdue, C22-3.

Ishii, N., Oku, T., Anami, K. and Fukuda, A. 2004. Lubrication Mechanism at Thrust slide bearing of Scroll Compressors (Experimental Study). International Compressor Engineering Conference at Purdue. C103.

Kulkarni, S. S. 1990a. Thrust Bearing Design Under Laminar Conditions, International Compressor Engineering Conference at Purdue, pp.327-332.

Kulkarni, S. S. 1990b. Thrust Bearing Design with Rigid Body Dynamics of The Runner Plate, International Compressor Engineering Conference at Purdue, pp.333-344.

Nishiwaki, F., Hasegawa, H., Ikoma, M., Matsuzaki, R. and Muramatsu, S. 1996. Mechanical Loss Reduction at Thrust Bearings of Scroll Compressors Using R407C, International Compressor Engineering Conference at Purdue.

Oku, T., Anami, K., Ishii, N. and Sano, K. 2004. Lubrication Mechanism at Thrust slide bearing of Scroll Compressors (Theoretical Study). International Compressor Engineering Conference at Purdue. C104.

PTC. 2006. Pro/MECHANICA Wildfire, Release 24.8. Needham, MA: Parametric Technology Corporation.

Noriaki Ishii, PhD, PE

Tatsuya Oku, PhD

Keiko Anami, PhD

Charles W. Knisely, PhD

Kiyoshi Sawai

Takashi Morimoto

Noboru Iida

Received February 12, 2007; accepted December 12, 2007

Noriaki Ishii is a professor in the Department of Mechanical Engineering, Osaka Electro-Communication University, Osaka, Japan. Tatsuya Oku is an engineer in the R&D Center, Mayekawa Manufacturing Company, Ltd., Ibaraki, Japan. Keiko Anami is an assistant professor in the Department of Mechanical Engineering, Ashikaga Institute of Technology, Tochigi, Japan. Charles W. Knisely is a professor in the department of mechanical engineering, Bucknell University, Lewisburg, PA. Kiyoshi Sawai is general manager and Takashi Morimoto and Noboru Iida are managers of the Refrigeration, Air-Conditioning & Heating Research Lab., Matsushita Electric Industrial Company, Ltd., Shiga, Japan.
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No portion of this article can be reproduced without the express written permission from the copyright holder.
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Author:Ishii, Noriaki; Oku, Tatsuya; Anami, Keiko; Knisely, Charles W.; Sawai, Kiyoshi; Morimoto, Takashi;
Publication:HVAC & R Research
Geographic Code:1USA
Date:May 1, 2008
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