外文翻譯--葉片泵磨損理論和實(shí)踐方面【優(yōu)秀】

1、<p><b>　　英文原文</b></p><p>　　wear 181-183 (1995) 868-875</p><p>　　Case Study</p><p>　　Theoretical and practical aspects of the wear of vane pumps</p><p>

2、　　Part B. Analysis of wear behaviour in the Vickers vane pump</p><p><b>　　test</b></p><p>　　A. Kunz a, R. Gellrich b, G. Beckmann c, E. Broszeit a</p><p>　　a Institute o

3、f Material Science, Technical University Darmstadt, P.O. Box 11 1452, 64229 Darmstadt,Gcmb University for Technol08y, Economy and Social Science Zittau/Goditz, Facuky of Maihematics, P.O. Box 264, 02763 Zutau</p>

4、<p>　　cPetersiliensrr. 2d, 03044 Cottbus, Received 16 August 1994; accepted l November 1994</p><p><b>　　Abstract</b></p><p>　　The wear behaviour of the vane pump used in the s

5、tandard method for indicating the wear characteristics of hydraulicfluids (ASTM D 2882/DIN 51 389) has been examined by comparison of the calculated wear and experimental data using alubricant without any additives. In a

6、ddition to the test series according to DIN 51 389, temperature profiles from the pump have been analysed using the bulk temperatures of the contacting components and the temperature in the lubrication gap as input data

7、for the wea</p><p>　　Keywords: Vane pumps; Hydraulic fluids; Wear prediction; Vickers vane pump test</p><p>　　1. Introduction</p><p>　　Efforts to develop a mathematical tool for wea

8、rprediction will not be successful without considering wear and its phenomena. The task of Part B of this study is to describe the analysis of the wear behaviour in the tribo system investigated and how the knowledge ach

9、ieved influences the calculations. Input data are derived from the measurement of mechanical and geometrical quantities, such as the hardness, stylus profilometry, fluid properties and contact radii. Thermal quantities a

10、re also essent</p><p>　　2. Experiments</p><p>　　AlI Vickers vane pump tests described were run with the same fiuid. It is a reference oil of the German Rcscarch Association for Transmission Tech

11、nique (FVA), and is a mineral oil without any additives (FVA3). Thus the disturbing influences of additives can be excluded.</p><p>　　2./.Input data for calculation</p><p>　　Fig. 1 lists the inp

12、ut and output quantities of the calculations. Most of the input parameters were derived surface profiles contact force and contact velocity dynamic viscosity contact radiihar

13、dness values Youngs moduli, Poisson numbersand lubrication gapspecific shear energy densities* pressure exponentc,f viscosity; tlubrication gap temperature</p><p>　　Rough

14、 surfuce ←→ shaar energy hypot ←→ elasto liubiction</p><p><b>　　↓</b></p><p><b>　　Wm=f（t）</b></p><p>　　Wf =f（ɑ）</p>&

15、lt;p>　　Fig. 1. Input parameters and output quantities of the mathematicalmodel of Part A.</p><p>　　Fig. 2. Cartridge V 104 C: bushing, rotor, ring, bushing (abcwe),single vane, pin (below).experimentally

16、from all the components involved beforeand after use in the vane pump tests. The mechanical components, which must be renewed for each test run,are shown in Fig. 2. Such a cartridge kit consists of a rotor, ring, 12 vane

17、s, bushings and pin.</p><p>　　Stylus profilometry was performed on the inner surface of the ring and on the tips of two vanes of the cartridge before and after each test run. Earlier investigations have show

18、n that ten parallel sections in the sliding direction on each body are sufficient to describe the surface topography in a statistically satisfactory manner as a two-dimensionalisotropic gaussianfield according to Ref. [1

19、]. Only the high pass filtered components of the profile (sampling length, 1.5 mm; cut o五 0.25 mm) were</p><p>　　The contact force and contact velocity were calculated with different fluid pressures and dyna

20、mic forces acting on the vanes, revolution number and ring radu, whereas the change in contact radius was documented with a profile projector. Because the ring radii are much larger thar) the radii of the vanes in the co

21、ntact zone, the vanes can be assumed to be hertzian cylinders sliding</p><p>　　along a plane surface and the contact radii are simply the radii of the vane tips. Each vane tip was twice drawn up at magnifica

22、tions of 100 : 1 and the contact radii and contact locations were measured with a stenciLMean values of the contact radii were transferred to the calculation, which is based (similar to the surfaceprofiles) on vanes in b

23、oth conditions.</p><p>　　The Vickers hardness HVlO was measured on thering and three vanes of each cartridge. This hardnessleads to a better reproducibility than microhardness values, but due to the large in

24、denter load, it couldonly be taken after the test runs. Therefore changes in hardness values could not be registered.</p><p>　　The Young's moduli, Poisson numbers and densities of the ring (AISI 52100) a

25、nd vane materials (M2 reg C) are the first input parameters in the shear energy hypothesis and were obtained from the literature. The specific shear energy densities (see Part A) are materialspecific constants [2l.</p

26、><p>　　The fluid properties (Fig. 1) were measured, derived from the literature or calculated. To obtain the dynamicviscosity, the densities and kinematic viscosities at 20,40 and 80 0C were measured. Because t

27、he fluid is a reference oil of FVA, the pressure exponent of the viscosity is given [3]. The temperature in the lubrication gap between the ring and vanes was approx:imated by measurements and calculations described belo

28、w.</p><p>　　2.2. Temperature profiles</p><p>　　Temperature measurement was performed to obtain information on how a heatable tribometer must be controlled to simulate the wear behaviour of the v

29、ane pump. Therefore shortened test runs were carried out until temperatures were stabilized. These 10 h vane pump tests delivered the input data for the approximation of the lubrication gap temperature in the ring-vane c

30、ontact, as well as additional wear masses to be compared with the calculated progressiort of wear in time. The sampling principles fo</p><p>　　The temperature of the lubricant in the gap between the ring and

31、 vanes was estimated to be equal to or greater than the bulk temperature on the inner ring surface. Following the first main statement of thermodynamics, the heat flux Q mp into the components of the pump can be derived

32、from with the fluid as the medium for energy transport.Qa,mp can only be transferred to the components shownin Fig. 2. For the same temperature differences and materials, this heat nUX can be divided into single comp<

33、/p><p>　　can be calculated and transferred to the model of elastohydrodynamic lubrication.</p><p>　　All test runs with the Vickers vane pump V 104 C were performed on a test rig according to ASTM D

34、2882/DIN 51 389, which is shown schematically in Fig.</p><p>　　4. These standards describe the procedure for testingthe anti-wear properties of hydraulic fiuids. To start the Vickers vane pump test according

35、 to the German standard, the system pressure must be raised in steps of 2 MPa every 10 min, beginning at 2 MPa, until a final pressure of 14 MPa is reached. At this stage, the fluid temperature measurcd bcfore the pump (

36、see Fig.4) must be controlled to guarantee a kinematic viscosity of 13 mm2 S-i at the inlet for every :tluid tested. These conditions mu</p><p>　　For performing the tests safely with the fluid FVA3, it was p

37、reheated t0 40 0C and circulated in a pressurefree way. The damage which may occur during the critical first hour of the runs can be avoided using TiNcoated bushings [4]. For comparison with the results derived from comp

38、utation, the wear produced in these runs must be documented as amounts, both locally and temporally.</p><p>　　The wear masses were derived from the weight differences of the ring and vanes before and after e

39、ach run. They were obtained from a sequence of four 250 h test runs and tw0 10 h runs for temperature measurement. The local linear amount of wear was documented by the differences in the inner ring radii perdegree of re

40、volution, which were measured by surface digitization along the inner ring surface at three different positions of the ring width before and after the tesi runs.</p><p>　　In earlier investigations [5], the w

41、ear progression over time of the vanes was measured under identical testing conditions, except for a lower fluid temperature. For this experiment, the radiotracer technique was used. Two vane tips in the set of 12 vanes

42、of each cartridge were radiologically activated by bombardment with protons. A detector close to the pump body allowed thedecrease in radiological activity to be monitored continuously, which was found to be reciprocally

43、 proportional to the lin</p><p>　　behaviour of the activated zone of the vane tips could be excluded.</p><p>　　Phyd+Pfric-Qcomp-Qfluid=0 (1) </p><

44、;p>　　Qfluid=mcfluid△Tfluid (2) </p><p>　　Fig. 4. Hydraulic circuit of the test rig.</p><p><b>　　3 result</b></p><p>　　lines the statistical r

45、eliability of surface modelling as a two-dimensional isotropic gaussian field. Although only the filtered profiles scanned in the sliding direction are shown, a distinct change in surface roughness is obvious. A good rep

46、resentation of the wear phenomena (see Part A) by the input data for the wear calculation derived from these profiles can be assumed.</p><p>　　The change in the vane tip shape over the testing period is docu

47、mented in Part A. The hardness values for the rings and vanes varied from 743 t0 769 HVlO (rings) and from 778 t0 816 HVlO (vanes). In all cases, the vanes of one cartridge had higher hardness values than the ring, but t

48、hese differences varied and had a large influence on the wear calculation (see Part A).</p><p>　　The measurement of the fiuid properties led, in combination with the kinematic viscosity prescribed by the Ger

49、man standard, to a fluid temperature of 84-86oC at the pump inlet. Together with the other temperature measurements acquired in the 10 h runs, these temperature profiles are illustrated in Fig. 6.</p><p>　　T

50、est Number t was found that, in about l h, all temperatures were stabilized. It should be noted that all temperatures in or on the pump components are higher than the fluid temperature measured behind the pump. The highe

51、st temperatures were found on the outer ring surface,</p><p>　　with significant differences depending on the location of the thermocouples.</p><p>　　The calculation of the bulk temperatures on t

52、he inner ring surface via the heat flux balance eliminated the infiuence of the different ring thicknesses at the scan locations. Depending on tbese different distances for heat conduction, between 4 and 7 0C must be add

53、ed to the mean values of the component temperatures to obtain the surface temperatures. These values are 20c70 higher than the fluid temperature measured behind the pump, which was used as input data for the wear</p&g

54、t;<p>　　calculation.</p><p>　　During the l h starting phase of the test runs, the stepwise increase in system pressure leads to an immediate effect on the component temperatures, whereas the fluid tem

55、perature increases with a more or less constant gradient, which demonstrates the association of load and frictional heat.</p><p>　　The four 250 h test runs caused a mixture of adhesive and abrasive wear at a

56、 high level (see Part A). The wear results achieved are shown in Fig. 7. Ring wear increased from test I to test 3. Therefore the 12 pm filter normally used was replaced after the third test by a 3 pm filter, and a press

57、ure-free run with an additional cartridge was started as a cleaning procedure. Due to the filter change, the reservoir needed to be refilled by about lOv-/o of its content with fresh fluid before control </p><

58、p>　　on the measurement location with strictly limited areas of high and low wear.</p><p>　　The results of continuous vane wear monitoring are shown in Fig. 8 in addition to the principle of measurement. D

59、egressive wear laps were found, where the stationary level was reached after 100 h.</p><p>　　4. Discussion</p><p>　　Before the wear calculations can be verified by wear data, it must be demonstr

60、ated that the assumptions,measurements and calculations forming the input for the mathematical model correlate with the wear measured.Fig. 9 compares the calculated load n the ring-vane contact, derived from the contact

61、force and changing shapes of the vane tips introduced in Part A, with the measured linear amounts of wear along the inner ring surface and the temperature distribution at the same place. There is qualit</p><p&

62、gt;　　gression of hertzian pressure and the linear amounts of wear, serious mistakes in the collection of input information are probably avoided, so that the verification of the calculated wear results by experimental dat

63、a will show the validity of the mathematical model.</p><p>　　For local amounts of linear ring wear, this verification can be seen in Fig. 10. It should be noted that the calculation and experimental results

64、are placed in the same decade, the progressions show the characteristic leaps similar to the load in Fig. 9 at almost the same degrees and the amounts are directly comparable. The loading zones are adapted to the progres

65、sion of the contact force (see Part A), which the calculated linear wear must follow as well as the hertzian pressure. The</p><p>　　different shapes of the two graphs between 300 and 700 (2100 and 2500) turn

66、 angles remain unsatisfactory, because this shows an uncertainty in the load as sumptions. The fluid pressure in a cell formed by two vanes, rotor and ring was assumed to be segmentally constant. Therefore the contact fo

67、rce was determined to follow these assumptions, which need to be dempressure) in the ring-vane contact, derived from the contact force and changing shapes of the vane tips introduced in Part A, with the me</p><

68、;p>　　For local amounts of linear ring wear, this verification can be seen in Fig. 10. It should be noted that the calculation and experimental results are placed in the same decade, the progressions show the character

69、istic leaps similar to the load in Fig. 9 at almost the same degrees and the amounts are directly comparable. The loading zones are adapted to the progression of the contact force (see Part A), which the calculated linea

70、r wear must follow as well as the hertzian pressure. The</p><p>　　different shapes of the two graphs between 300 and 700 (2100 and 2500) turn angles remain unsatisfactory, because this shows an uncertainty i

71、n the load assumptions. The fluid pressure in a cell formed by two vanes, rotor and ring was assumed to be segmentally constant. Therefore the contact force was determined to follow these assumptions, which need to be de

72、m-</p><p>　　onstrated by corresponding experiments. Comparedwiththe measurement, the contact force in loading zone I was assumed to be too high and caused values above the experimental data. This was due to

73、difficulties in modelling the large variety of vane tip geometries which can appear in one cartridge and strongly determine the contact force in this region. More information about</p><p>　　the reliability o

74、f load assumptions could have been obtained from a knowledge of the bulk surface temperatures, which were not measured or calculated.Despite other deviations of the two progressions, the areas below each graph are compar

75、able, so that a good correlation between calculated and measured wear masses can be expected.</p><p>　　For vane wear masses after 250 h of testing, the expectations have been fulfilled in a satisfactory mann

76、er, If the progression of the vane wear mass in time (see Part A) is verified by the measured amounts of linear wear (Fig. 8), good correlation of the progressions can be found. For both values, the stationary phase is r

77、eached after 100 h. The differences in the gradients during the stationary phase may be caused by the different tcmperatures of the two test series. The dif- ferences in the ri</p><p>　　of the wear masses fo

78、r that time.</p><p>　　conclusion</p><p>　　IThe following conclusions can be drawn.</p><p>　　(1) For the wear system, Vickers vane pump V 104C/lubricant FVA 3, a good correlation bet

79、ween</p><p>　　load and wear location on the ring was found, which is associated with a corresponding temperature distribution.</p><p>　　(2) The load assumptions are widely confirmed.</p>

80、<p>　　(3) The mathematical model introduced in Part A, withinput information based on these assumptions, delivers wear masses, progressions of wear masses in time and local amounts of linear wear which correlate w

81、ith corresponding experiments.</p><p>　　(4) This mathematical model based on the shear energy hypothesis is a qualified instrument for retracing thewear behaviour in friction regimes with boundary lubricati

82、on, with the exclusion of additives.</p><p>　　(5) Large efforts are necessary to obtain qualified input data.</p><p>　　(6) Wear prediction is not possible, because several parameters derived fr

83、om investigations on components in their final condition need to be used as input data.</p><p>　　Future investigations are required.</p><p>　　(1) To improve the assumptions on the structure of t

84、he fiuid pressure in the pump.</p><p>　　(2) To develop a method to obtain all input data from components in the new condition to allow real wear prediction.</p><p>　　(3) To enlarge the theory w

85、ith an empirical statement describing the influence of additives. Experiments and investigations similar to those in this paper have been performed with the same fluid containing additives and with a commercially availab

86、le HM fluid.</p><p><b>　　中文譯文</b></p><p>　　葉片泵磨損理論和實(shí)踐方面</p><p>　　第二部分：關(guān)于維克斯公司葉片泵實(shí)驗(yàn)?zāi)p情況分析</p><p>　　Kunza，R.Gellrichb，G.Beckmannc，E.Broszeita</p><p

87、>　　a材料科學(xué)研究所，達(dá)姆城工業(yè)大學(xué)，P.O.Box 11 14 52，64229 達(dá)姆城，德國</p><p><b>　　b齊陶</b></p><p><b>　　摘要</b></p><p>　　葉片泵的磨損狀況標(biāo)準(zhǔn)方法是指示水力的失效流體（美國材料試驗(yàn)學(xué)會 D 2882/德國工業(yè)標(biāo)準(zhǔn) 51389)已經(jīng)被通

88、過用沒有任何添加劑的潤滑劑得到的失效計(jì)算和審查實(shí)驗(yàn)數(shù)據(jù)審查。除了依照德國工業(yè)標(biāo)準(zhǔn)得到的檢驗(yàn)系列之外，泵的剖面溫度已經(jīng)用來自絕大部分聯(lián)系原件和間縫潤滑之間的溫度作為失效計(jì)算的原始數(shù)據(jù)。根據(jù)德國標(biāo)準(zhǔn)檢驗(yàn)的卷筒已經(jīng)被前前后后嚴(yán)格的測試為了獲得精確模型的原始數(shù)據(jù)和確定磨損位置。執(zhí)行流體的性能分析和在液壓環(huán)路中粒子磨損的調(diào)查。實(shí)驗(yàn)結(jié)果和預(yù)測的相比較，預(yù)測的是由協(xié)議核實(shí)負(fù)荷條件，時(shí)間磨損過程和當(dāng)?shù)氐哪p證實(shí)的。已經(jīng)得出關(guān)于合理的載荷消耗和失效校核的

89、結(jié)論，就像這種方法在添加劑存在的適用性范圍。</p><p><b>　　1.說明</b></p><p>　　在沒有考慮到失效磨損的一些現(xiàn)象時(shí)，努力去開發(fā)一種精確工具去預(yù)測磨損失效是不會成功的。本研究第二部分的目的就是為了描述磨損行為在調(diào)查的tribo系統(tǒng)中的分析和怎樣運(yùn)用知識完成影響計(jì)算。初始數(shù)據(jù)來源于機(jī)械的測量和幾何量，比如硬度，針式輪廓，流體特性和接觸半徑。熱

90、量對模型潤滑也是必不可少的的數(shù)據(jù)量。</p><p><b>　　2.實(shí)驗(yàn)</b></p><p>　　所有的維克斯葉片泵的實(shí)驗(yàn)都是用同種的流體。它是德國一個(gè)研究協(xié)會為傳輸技術(shù)涉及的一種油FVA，它是一種沒有任何添加劑的礦物質(zhì)油FVA3。因此可以排除添加劑所引起的后果。</p><p><b>　　2.1計(jì)算原始數(shù)據(jù)</b>

91、;</p><p>　　數(shù)據(jù)1.列出輸入和輸出的計(jì)算量。大部分參數(shù)來源于：</p><p>　　粗糙度 流體性質(zhì)</p><p>　　平面度 接觸力和接觸速度 動態(tài)粘度</p><p>　　接觸半徑 粘性的壓力指數(shù)</p><p>　　硬度標(biāo)

92、準(zhǔn) 間隙潤滑溫度</p><p><b>　　泊松數(shù)和密度</b></p><p><b>　　實(shí)際單位剪切力</b></p><p>　　隨即模擬的粗糙表面 ←→ 剪切力假說 ←→ 彈性液壓潤滑</p><p><b>　　↓</b><

93、/p><p><b>　　Wm=f（t）</b></p><p>　　Wf =f（ɑ）</p><p>　　圖1.數(shù)學(xué)模型的第一部分的原始參數(shù)和實(shí)際工程量</p><p>　　圖2.卷筒V104C：套管，轉(zhuǎn)子，定子，上套管，單一葉片，釘</p><p>　　所有涉及實(shí)驗(yàn)前后的原件在葉片泵測試都用

94、到了。在每個(gè)實(shí)驗(yàn)中的機(jī)械原件都不一樣，比如在圖2中卷筒由轉(zhuǎn)子，定子，12個(gè)輪葉，套管和釘組成。</p><p>　　在每一個(gè)實(shí)驗(yàn)前后古老的輪廓測定法會在卷筒的環(huán)的內(nèi)表面和兩個(gè)葉片的頂端測定。根據(jù)專家所說，較早的研究已經(jīng)指出十個(gè)類似的部分在每個(gè)部分的不同方向由統(tǒng)計(jì)學(xué)來描述表面度是足夠的。只有輪廓中重要的濾過部件（采樣長度1.5mm，剪下0.25mm）用于測定光譜時(shí)刻m0，m2，m4和粗糙度ɑ。依據(jù)不同承載位置的接觸

95、力的劃分，新表面的地形圖數(shù)據(jù)被用于平面Ⅳ（低負(fù)載，參考part A）。對于另外一些存在高載荷的平面，最后一個(gè)條件的表面的輪廓被用了，在試運(yùn)轉(zhuǎn)之后證明外表內(nèi)環(huán)符合要求。</p><p>　　盡管接觸半徑的變化被記錄在剖面投光器，接觸力和接觸速度還是根據(jù)葉片上不同的流體壓力，動力，旋轉(zhuǎn)量和環(huán)半徑計(jì)算得出。因?yàn)槎ㄗ拥陌霃竭h(yuǎn)遠(yuǎn)大于在接觸位置葉片的半徑，葉輪能被假定變得赫茲圓柱體滑動向前一個(gè)平的表面和接觸半徑只是葉輪的半徑

96、。每一個(gè)葉片的尖端的損耗是100:1的兩倍，并且接觸半徑和接觸為定位由模板測量。接觸半徑的平均值由計(jì)算得到，而計(jì)算是根據(jù)兩種不同的條件。</p><p>　　測得定子和三個(gè)葉片的硬度為10HV，這個(gè)硬度值決定了它比微硬度值有更好的彈性，但是由于存在大的切應(yīng)力，所以它只能在試驗(yàn)之后得到。所以硬度標(biāo)準(zhǔn)不能被注冊。</p><p>　　泊松數(shù)，模數(shù)，定子的密度和葉片原料是從文獻(xiàn)中得到的剪切力假說

97、中最基本的參數(shù)。實(shí)際的單位剪切力是不變的。</p><p>　　數(shù)據(jù)1.中的流體性質(zhì)是由計(jì)算和文獻(xiàn)中得到的。密度和運(yùn)動粘性分別在20℃、40℃和80℃測量而得到動態(tài)粘性參數(shù)。粘性的壓力指數(shù)由德國傳輸工程動力研究協(xié)會給出。在定子和葉片間隙間的潤滑溫度接近于測定和計(jì)算得到的。</p><p><b>　　2.2 溫度分布</b></p><p>　

98、　測量溫度是為了獲得需要多少加熱量能使得接近葉片泵的磨損現(xiàn)象。因此要不斷縮短實(shí)驗(yàn)期直到溫度穩(wěn)定為止。這些 10 h 輪葉泵檢驗(yàn)為近似值遞送輸入數(shù)據(jù)潤滑間隙溫度在這定子與輪葉的接觸，連同另外磨耗集合被與磨耗的有計(jì)劃級數(shù)相較及時(shí)。葉片泵的溫度分布在數(shù)據(jù)3.通過抽樣原理論證。</p><p>　　在定子和葉片間隙中的潤滑溫度估計(jì)會等于或稍高于內(nèi)部定子主題表便的溫度。其次主要的熱力學(xué)報(bào)表，熱流量Qcomp可由一下得<

99、;/p><p>　　數(shù)據(jù)3.溫度測量原理</p><p>　　Phyd+Pfric-Qcomp-Qfluid=0 (1)</p><p><b>　　和</b></p><p>　　Qfluid=mcfluid△Tfluid (2)

100、 </p><p><b>　　圖3.</b></p><p>　　流體作為能量運(yùn)輸?shù)拿浇?，熱量通量可以在圖2.中體現(xiàn)出來。同樣的溫度差異和材料的熱通量可分為單根據(jù)構(gòu)件的關(guān)系fhrxes腫塊。推導(dǎo)過程中產(chǎn)生的熱量通量向是流動的在一段時(shí)間的徑向通過定子。與已知的溫度在外環(huán)線表面上的溫度，大部分的內(nèi)圈的表面能計(jì)算和轉(zhuǎn)移到模型。&l

101、t;/p><p><b>　　2.3.資料比較</b></p><p>　　所有的測試運(yùn)行與維氏的葉片泵V 104 C試驗(yàn)臺進(jìn)行了按照ASTM(美國材料試驗(yàn)協(xié)會)D嗎2882 / DIN 51 389位，這體現(xiàn)schematically圖4。這些標(biāo)準(zhǔn)描述程序進(jìn)行測試抗磨液壓流體的性質(zhì)。開始葉片泵試驗(yàn)的維氏據(jù)德國標(biāo)準(zhǔn)、系統(tǒng)壓力必須得到提高的腳步每隔10分鐘的2兆帕，開始在2