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C6132
改装
双工
数控
缸体
钻床
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C6132改装成双工位数控缸体孔钻床,C6132,改装,双工,数控,缸体,钻床
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Experimental investigation of laser surface textured parallel thrustbearingsI. Etsion*, G. Halperin, V. Brizmer and Y. KligermanDept. of Mechanical Engineering Technion, Haifa 32000, IsraelReceived 14 September 2003; accepted 29 December 2003Performance enhancements by laser surface texturing (LST) of parallel-thrust bearings is experimentally investigated. Testresults are compared with a theoretical model and good correlation is found over the relevant operating conditions. A compari-son of the performance of unidirectional and bi-directional partial-LST bearings with that of a baseline, untextured bearing ispresented showing the benefits of LST in terms of increased clearance and reduced friction.KEY WORDS: fluid film bearings, slider bearings, surface texturing1. IntroductionThe classical theory of hydrodynamic lubricationyields linear (Couette) velocity distribution with zeropressure gradients between smooth parallel surfacesunder steady-state sliding. This results in an unstablehydrodynamic film that would collapse under anyexternal force acting normal to the surfaces. However,experience shows that stable lubricating films candevelop between parallel sliding surfaces, generallybecause of some mechanism that relaxes one or moreof the assumptions of the classical theory.A stable fluid film with sufficient load-carryingcapacity in parallel sliding surfaces can be obtained,for example, with macro or micro surface structure ofdifferent types. These include waviness 1 and protrud-ing micro-asperities 24. A good literature review onthe subject can be found in Ref. 5. More recently,laser surface texturing (LST) 68, as well as inletroughening by longitudinal or transverse grooves 9were suggested to provide load capacity in parallelsliding. The inlet roughness concept of Tonder 9 isbased on effective clearance reduction in the slidingdirection and in this respect it is identical to the par-tial-LST concept described in ref. 10 for generatinghydrostatic effect in high-pressure mechanical seals.Very recently Wang et al. 11 demonstrated experi-mentally a doubling of the load-carrying capacity forthe surface- texture design by reactive ion etching ofSiC parallel-thrust bearings sliding in water. Thesesimple parallel thrust bearings are usually found inseal-less pumps where the pumped fluid is used as thelubricant for the bearings. Due to the parallel slidingtheir performance is poorer than more sophisticatedtapered or stepped bearings. Brizmer et al. 12 demon-strated the potential of laser surface texturing in theform of regular micro-dimples for providing load-car-rying capacity with parallel-thrust bearings. A modelof a textured parallel slider was developed and theeffect of surface texturing on load-carrying capacitywas analyzed. The optimum parameters of the dimpleswere found in order to obtain maximum load-carryingcapacity. A micro-dimple collective effect was identi-fied that is capable of generating substantial load-car-ryingcapacity,approachingthatofoptimumconventional thrust bearings. The purpose of the pres-ent paper is to investigate experimentally the validityof the model described in Ref. 12 by testing practicalthrust bearings and comparing the performance ofLST bearings with that of the theoretical predictionsand with the performance of standard non-texturedbearings.2. BackgroundA cross section of the basic model that was analyzedin Ref. 12 is shown in figure 1. A slider having awidth B is partially textured over a portion Bp aB ofits width. The textured surface consists of multipledimples with a diameter 2r?p, depth h?pand area densitySp. As a result of the hydrodynamic pressure generatedby the dimples the sliding surfaces will be separated bya clearance h?0depending on the sliding velocity U, thefluid viscosity l and the external load W?. It was foundin Ref. 12 that an optimum ratio exists for the param-eter hp h?p=h?0that provides maximum dimensionlessload-carrying capacity W W?h?02=lULB2where L is*To whom correspondence should be addressed. E-mail: etsiontx.technion.ac.il1023-8883/04/08000295/0 ? 2004 Plenum Publishing CorporationTribology Letters, Vol. 17, No. 2, August 2004 (? 2004)295the bearing length, and this optimum value is hp=1.25. It was further found in Ref. 12 that an optimumvalue exists for the textured portion a depending onthe bearing aspect ratio L/B. This behavior is shown infigure 2 for a bearing with L/B = 0.75 at various val-ues of the area density Sp. As can be seen in the rangeof Spvalues from 0.18 to 0.72 the optimum a valuevaries from 0.7 to 0.55, respectively. It can also be seenfrom figure 2 that for a 0.85 no optimum valueexists for Spand the maximum load W increases withincreasing Sp. Hence, the largest area density that canbe practically obtained with the laser texturing isdesired. It is also interesting to note from figure 2 theadvantage of partial-LST (a 1) over the full LST(a = 1) for bearing applications. At Sp= 0.5, forexample, the load W at a = 0.6 is about three timeshigher than its value at a = 1. A full account of thisbehavior is given in Ref. 12.3. ExperimentalThe tested bearings consist of sintered SiC disks10 mmthick,having85 mmouterdiameterand40 mm inner diameter. Each bearing (see figure 3)comprises a flat rotor (a) and a six-pad stator (b). Thebearings were provided with an original surface finishby lapping to a roughness average Ra= 0.03 lm. Eachpad has an aspect ratio of 0.75 when its width is mea-sured along the mean diameter of the stator. The pho-tographs of two partial-LST stators are shown infigure 4 where the textured areas appear as brightermatt surfaces. The first stator indicated (a) is a unidi-rectional bearing with the partial-LST adjacent to theleading edge of each pad, similar to the model shownin figure 1. The second stator (b) is a bi-directionalversion of a partial-LST bearing having two equal tex-tured portions, a/2, on each of the pad ends. The lasertexturing parameters were the following; dimple depthh?p= 6.5?0.5 lm, dimple diameter 2r?p= 60 ?5 lmand dimple area density Sp= 0.6?0.03. These dimpledimensions were obtained with 4 pulses of 30 ns dura-tion and 4 mJ each using a 5 kHz pulsating Nd:YAGlaser. The textured portion of the unidirectional bear-ing was a= 0.73 and that of the bi-directional bearingwas a= 0.63. As can be seen from figure 2 both thesea values should produce load-carrying capacity varyclose to the maximum theoretical value.The test rig is shown schematically in figure 5. Anelectrical motor turns a spindle to which an upperholder of the rotor is attached. A second lower holderFigure 1. A cross section of a partially laser surface textured parallelslider.00.0050.010.0150.020.0250.030.03500.10.20.30.40.50.60.70.80.91Textured Portion, a Load Capacity, WSp = 0.18 Sp= 0.32Sp = 0.5Sp = 0.72 L/B = 0.75hp= 1.25Figure 2.Dimensionless load-carrying capacity, W, vs. textured portion, a, for various densities, Sp.296I. Etsion et al./Investigation of LST parallel thrust bearingsof the stator is fixed to a housing, which rests on ajournal bearing and an axial loading mechanism thatcan freely move in the axial direction. An arm thatpresses against a load cell and thereby permits frictiontorque measurements prevents the free rotation of thishousing. Axial loading is provided by means of deadweights on a lever and is measured with a second loadcell. A proximity probe that is attached to the lowerholder of the stator allows on-line measurements ofthe clearance change between rotor and stator as thehydrodynamic effects cause axial movement of thehousing to which the stator holder is fixed. Tap wateris supplied by gravity from a large tank to the centerof the bearing and the leakage from the bearing is col-lected and re-circulated. A thermocouple adjacent tothe outer diameter of the bearing allows monitoring ofthe water temperature as the water exit the bearing. APC is used to collect and process data on-line. Hence,the instantaneous clearance, friction coefficient, bear-ing speed and exit water temperature can be monitoredconstantly.The test protocol includes identifying a referencezero point for the clearance measurements by firstloading and then unloading a stationary bearing overthe full load range. Then the lowest axial load isapplied, the water supply valve is opened and themotor turned on. Axial loading is increased by stepsof 40 N and each load step is maintained for 5 minfollowing the stabilization of the friction coefficient ata steady-state value. The bearing speed and water tem-perature are monitored throughout the test for anyirregularities. The test ends when a maximum axialload of 460 N is reached or if the friction coefficientexceeds a value of 0.35. At the end of the last loadstep the motor and water supply are turned offandthereferencefortheclearancemeasurementsisrechecked. Tests are performed at two speeds of 1500and 3000 rpm corresponding to average sliding veloci-ties of 4.9 and 9.8 m/s, respectively and each test isrepeated at least three times.4. Results and discussionAs a first step the validity of the theoretical modelin Ref. 12 was examined by comparing the theoreti-cal and experimental results of bearing clearance ver-susbearingloadforaunidirectionalpartial-LSTbearing. The results are shown in figure 6 for the twospeeds of 1500 and 3000 rpm where the solid anddashed lines correspond to the model and experiment,respectively. As can be seen, the agreement betweenthe model and the experiment is good, with differ-ences of less than 10%, as long as the load is above150 N. At lower loads the measured experimentalclearances are much larger than the model predic-tions, particularly at the higher speed of 3000 rpmwhere at 120 N the measured clearance is 20 lm,which is about 60% higher than the predicted value.It turns out that the combination of such large clear-ances and relatively low viscosity of the water mayresult in turbulent fluid film. Hence, the assumptionof laminar flow on which the solution of the Rey-nolds equation in Ref. 12 is based may be violatedFigure 3. A photograph of the parallel-thrust bearing showing theflat rotor (a) and the six pads stator (b).Figure 4. A photograph of two textured stators showing a unidirec-tional (a) and a bi-directional (b) versions of the partial-LST thrustbearing.MotorSpindleWater supplyVelocityGaugeProximityprobeBearingrotorClearanceThermocoupleLoad cellsLoadBearingstatorFigure 5. A schematic description of the test rig.Etsion et al./Experimental investigation of LST bearings297making the model invalid especially at the higherspeed and lowest load. In order to be consistent withthe model of Ref. 12 it was decided to limit furthercomparisons to loads above 150 N.It should be noted here that the first attempts to testthe baseline untextured bearing with the original sur-face finish of Ra= 0.03 lm on both the stator androtor failed due to extremely high friction even at thelower loads. On the other hand the partial-LST bearingran smoothly throughout the load range. It was foundthat the post-LST lapping to completely remove about2 lm height bulges, which are formed during texturingaround the rims of the dimples, resulted in a slightlyrougher surface with Ra= 0.04 lm. Hence, the baselineuntextured stator was also lapped to the same rough-ness of the partial-LST stator and all subsequent testswere performed with the same Ravalue of 0.04 lm forall the tested stators. The rotor surface roughnessremained, the original one namely, 0.03 lm. Figure 7presents the experimental results for the clearance as afunction of the load for a partial-LST unidirectionalbearing(seestatorinfigure 4(a)anda baselineuntextured bearing. The comparison is made at the twospeeds of 1500 and 3000 rpm. The area density of thedimples in the partial-LST bearing is Sp= 0.6 and thetextured portion is a 0:734. The load range extends57.51012.51517.5200 50100150200250300350400450500Load, NClearance, mSp = 0.6 = 0.73n = 3000 rpmn = 1500 rpm 5.010.015.020.0Figure 6. A comparison of the theoretical (solid lines) and experimental (dashed lines) results of the bearing clearance versus the load at 1500and 3000 rpm.02468101214150200250300350400450500Load, NClearance,mSp = 0.600 = 0.734Untexturedn = 1500 rpm n = 3000 rpm Figure 7. A comparison of clearance vs. load performance of partial-LST unidirectional bearing and a baseline untextured bearing.298I. Etsion et al./Investigation of LST parallel thrust bearingsfrom 160 to 460 N. The upper load was determined bythe test-rig limitation that did not permit higher load-ing. It is clear from figure 7 that the partial-LST bear-ing operates at substantially larger clearances than theuntextured bearing. At the maximum load of 460 Nand speed of 1500 rpm the partial-LST bearing has aclearance of 6 lm while the untextured bearing clear-ance is only 1.7 lm. At 3000 rpm the clearances are 6.6and 2.2 lm for the LST and untextured bearings,respectively. As can be seen from figure 7 this ratio ofabout 3 in favor of the partial-LST bearing is main-tained over the entire load range.Figure 8 presents the results for the bi-directionalbearing (see stator in figure 4(b). In this case the LSTparameters are Sp 0:614 and a 0:633. The clear-ances of the bi-directional partial-LST bearing arelower compared to these of the unidirectional bearingat the same load. At 460 N load the clearance for the1500 rpm is 4.1 lm and for the 3000 rpm it is 6 lm.These values represent a reduction of clearance between33 and 10% compared to the unidirectional case. How-ever, as can be seen from figure 8 the performance ofthe partial-LST bi-directional bearing is still substan-tially better than that of the untextured bearing.The friction coefficient of partial-LST unidirectionaland bi-directional bearings was compared with that ofthe untextured bearing in figures 9 and 10 for the twospeeds of 1500 and 3000 rpm, respectively. As can beseen the friction coefficient of the two partial-LSTbearings is very similar with slightly lower values inthe case of the more efficient unidirectional bearing.The friction coefficient of the untextured bearing ismuch larger compared to that of the LST bearings. At1500 rpm (figure 9) and the highest load of 460 N thefriction coefficient of the untextured bearing is about0.025 compared to about 0.01 for the LST bearings.Figure 8. A comparison of clearance vs. load performance of partial-LST and a baseline untextured bi-directional bearings.0.0050.0150.0250.0350.0450.0550.065150200250300350400450500Load, NFriction coefficientUntexturedBi-directionalUnidirectionalFigure 9. A comparison of friction coefficient of the partial-LST bearings and a baseline untextured bearing at 1500 rpm.Etsion et al./Experimental investigation of LST bearings299At the lowest load of 160 N the values are about 0.06for the untextured bearing and around 0.02 for theLST bearings. Hence, the friction values of the untex-tured bearing are between 2.5 and 3 times higher thanthe corresponding values for the partial-LST bearingsovertheentireloadrange.Similarresultswereobtained at the velocity of 3000 rpm (figure 10) butthe level of the friction coefficients is somewhat higherdue to the higher speed. The much higher friction ofthe untextured bearing is due to the much smallerclearances of this bearing (see figures 7 and 8) thatresult in higher viscous shear.5. ConclusionThe idea of partial-LST to enhance performance ofthe parallel thrust bearing was evaluated experimentally.Good correlation was found with a theoretical model aslong as the basic assumption of laminar flow in the fluidfilm is valid. At low loads with relatively large clear-ances, where turbulence may occur, the experimentalclearance is larger than the prediction of the model.The performance of both unidirectional and bi-directional partial-LST bearings in terms of clearanceand friction coefficient was compared with that of abaseline untextured bearing over a load range in whichthe theoretical model is valid. A dramatic increase, ofabout three times, in the clearance of the partial-LSTbearings compared to that of the untextured bearingwas obtained over the entire load range. Consequent
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