1、668 2009 ASHRAEABSTRACTThis experimental investigation is focused on improving the frictional characteristics of a scroll compressor thrust slide bearing undergoing an orbital motion. An experimental test rig was designed and constructed to accommodate actual parts of a scroll compressor and measure
2、 friction between a scroll compressor crankcase and its orbiting scroll. It was envisaged that in an orbiting situation where couette velocities in both tangential and radial directions are present, circular pockets would be able to provide the geometric wedge in both radial and tangential direction
3、 and thus enhance the lubrication performance by shifting the lubrication regime of the thrust slide bearing from boundary to full film. The lubrication performance of the surfaces with circular pockets was compared to that of unmodified specimens which were used as base line specimens. It was obser
4、ved that in orbiting situation macro circular pockets with large diameters viz. 0.3 to 0.36 inch (7.62 to 9.14 mm) and very shallow depth viz. 0.002 to 0.003 inch (0.5 to 0.076 mm) significantly improved the fric-tional performance between the scroll crankcase and orbiting scroll.INTRODUCTIONA thrus
5、t 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
6、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
7、 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.
8、radial, spiral, or circumferential) on the surface of the rotational thrust bearing to facilitate hydrody-namic 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 pr
9、oper size and arrangement, provide an effec-tive lubrication and load support mechanism. In this experi-mental 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 wil
10、l 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 g
11、rooves. Yu and Sadeghi (2001) showed analytically an effective groove method of load support to liquid-lubri-cated 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 geometr
12、y 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 Sadeghis Experimental Investigation of Orbiting Thrust Bearing Using Wide and Sh
13、allow Circular PocketsAmit Vaidya Farshid Sadeghi, PhDAmit Vaidya is a mechanical engineer, Syracuse, NY. Farshid Sadeghi is a professor in the School of Mechanical Engineering, Purdue Univer-sity, West Lafayette, IN. LO-09-063 2009, American Society of Heating, Refrigerating and Air-Conditioning En
14、gineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions 2009, vol. 115, part 2. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAEs prior written permission.ASHRAE Transactions 669computational model.
15、 They found that grooved thrust wash-ers 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 hydrody-namic performance of wet clutches. T
16、ian et al. (1989) studied boundary lubrication considering ploughing effects; they demonstrated that undulated titanium surfaces would consid-erably reduce friction coefficient and wear. Andriy et al. (2004) showed that use of dimples expands speed load param-eters for hydrodynamic lubrication. The
17、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 e
18、t 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
19、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 bear-ing. Wang et al. (2002) studie
20、d the effect of micro pores formed by laser texturing on a SIC surface.Computational models for smooth surfaces have demon-strated that by properly sizing the circular pockets on the bear-ing surface, the shift in the operation of thrust slide bearing (thrust bearing undergoing orbiting motion) from
21、 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 fric-tion, in a liquid-lubricated orbiting thrust slide bearing under a constant loading condition. This study
22、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.EXPERIMENTAL
23、TEST RIG AND SPECIMENSSeveral orbiting thrust washer specimens were investi-gated 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 interf
24、ace of orbiting scroll and crankcase under fully flooded conditions. OTWT was used to evaluate the effects of various surface features (circular pock-ets, grooves, etc.) on friction in orbiting thrust bearing. The design of OTWT is based on stacked plate concept, wherein various sections are support
25、ed 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 ar
26、rangement 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
27、 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.Fig
28、ure 1 depicts the OTWT. A 30 Amps 230 V AC contin-uous current variable speed and torque servo motor is used in Figure 1 Orbiting thrust washer test rig. Figure 2 Testing chamber.670 ASHRAE Transactionsvelocity control mode to operate the test rig. The motor is operated by a 200 Hz digital drive tha
29、t 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
30、 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 b
31、ottom. 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 dynami-cally balanced using counterweights such that its mass center is aligned with the central axis o
32、f 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
33、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
34、 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)
35、 located directly below the test chamber is capable of simultaneously measuring three orthog-onal force components along X, Y and Z axes and three moments. Data is acquired through six channel digital ampli-fier at a rate of 16000 samples per second for each channel. Load cell is mounted on a vertic
36、ally 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
37、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 show
38、s 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 hydro-dynamic load except through relatively weak and unstable mechanisms (e.g. Thermal d
39、istortion) 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 whe
40、ther the groove designs for rotational washers can be extended to orbiting thrust bearing circular grooves as shown in Figure 4 and 5 were experimentally eval-uated to determine the lubrication characteristics (regimes).In order to obtain the optimum size, shape and pattern of the circular pockets o
41、n the thrust surface, parametric study was performed on circular pockets by varying number of groove/pockets in the radial direction (Np), number of grooves/pock-ets (Ng), groove/pocket width (Gw) and groove/pocket depth in inches (Gd). Table 1 provides a comprehensive list of spec-imens used in thi
42、s investigation (D through K). Figure 6 illus-trates 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 circu-lar poc
43、kets is 0.00003 inch (0.00762 mm) with round Figure 3 Thrust surfaces of a scroll compressor.Figure 4 Intersecting groove pattern.ASHRAE Transactions 671circular milling marks and average surface finish on the land in between the pockets is 0.00003 inch (0.00762 mm) char-acterized by cross grinding
44、marks. Figure 7 depicts specimens (F, G and H) with varying width of the pockets (Gw). Further circular pockets with varying depth (Gd) (not shown but refer-enced as I, J and K in Table 1) were tested to obtain optimum depth that would enhance the lubrication at the thrust inter-face. All the specim
45、ens 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 thermocou-ple bead for temperature measurement.Tests were also carried out by changing surface finish (special grinding operation was performed on scrol
46、ls 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 PROCEDURETest specimen is mounted in the test chamber and ther-mocouple bead (K-Type) is inserted in a hol
47、e drilled just below the stationary thrust surface. The load cell is set to zero using hardware zero function on the amplifier. Performing hardware Table 1. Summary of Feature Depth (Gd) and Width (Gw) for Thrust SpecimensSpecimen GeometriesGw (in) Gd 0.0005 (in)(A) 0.04 0.02(B) 0.02 0.01(C) 0.984 0
48、.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 0.002Figure 5 Uniformly distributed radial circular grooves.Figure 6 Circular pockets with varying Np.Figure 7 Circular pockets with varying width (Gw).672 ASHRAE Transactionszero nullifies the
49、 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
