外文翻译--发动机轴承设计的发展 英文版.pdf

Developments in engine bearing design F.A. Martin* Some of the important recent developments in engine bearing design tech- niques are highlighted. The availability of increased computing power has enabled more realistic assumptions about bearing conditions to be considered; these include oil feed features, oil film history, non-circular bearings, inertia effects due to journal centre movement, improved prediction of main bearing loads, flexible housings and special bearings. References to these advances are made, together with illustrations of how they affect predicted bearing performance. Experimental evidence is also being obtained, which helps to verify and give confidence in the analytical predictions Keywords: journal bearings, bearings + design, hydrodynamic lubrication, bearing stress, bearing housings, oil grooves Engine bearing performance is dependent upon many factors, from the mechanical configuration of the engine to the hydrodynamics of the oil film. This paper highlights the more important factors to be considered, and relates them to recent advances, both published and unpublished, throughout the world. The review attempts not just to reference these advances, but to illustrate how they extend the areas of performance prediction, experimental verifica- tion and the design of special bearings. Historically, the earliest attempts at the design of dynamic- ally loaded bearings were based on maximum allowable specific load (defined as maximum applied load divided by projected bearing area), and this is still a valuable parameter. With the advent of graphical and numerical techniques capable of solving a hydrodynamic bearing model, albeit still highly simplified, estimates of minimum oil film thick- ness could be made, and used as a comparator to judge the likelihood of problems on new engines. A comprehensive study of those early predictive methods can be found in the 1967 review paper by Campbell et al I ; as a study case this used the big end bearing of a Ruston and Hornsby VEB Mk III 600 hp, 600 r/min diesel engine. Nearly twenty predicted and experimental journal orbits from various sources were discussed in the volume of I. Mech. E. proceedings which contained that paper, and the same study case is still being used by workers in this field today (polar load diagram, Fig 1 (a); complete data, Ref 1). It has been used in this review to illustrate some of the subse- quent advances in prediction capabilities. Many of the major assumptions used in the early prediction methods were certainly not realistic, but were used as expedients to obtain a mathematical model which could be solved with the limited computing capabilities then available. These assumptions included circular rigid bearings and a perfect supply of isoviscous Newtonian oil. In many cases the bearing surface was assumed to be uninterrupted by oil feed features in the developed film pressure regions and, external to the bearing, the calculation of the main bearing loads took no account of the crankshaft and crank- case stiffnesses. Over the last decade increases in computing power have meant that many of those early assumptions are no longer *Department of Applications Engineering, The Glacier Metal Com- pany Limited, Alperton. Wembley, Middlesex HAO 1HD, UK necessary and work has been carried out on bearing shapes 23 elastic connecting rod bearing 4 , oil feed feat- ures s6 , oil film history 7 , and more realistic main bearing load sharing 89 . This is in keeping, although a little late, with the 1967 prophecy from Campbell , which stated that: It is the authors belief that, with the continuing rapid advance in computational methods and with the growing awareness of the powerful design techniques which are A AB a D “- B k ,j b E C ,4 i i aT- C v Fig 1 Polar load diagrams for VEB connecting-rod bearing relative to: (a) connecting rod axis, (b) cylinder axis, (c) crankpin axis TRIBOLOGY international 0301 679X/83/030147 -18 $03.00 1983 Butterworth relating to the VEB big end stud, case use EooKers short oearing Mobility solution. The Mobi!ity coT:co-or :qas been successfully applied over the last t 5 years, ano. .z explained in detail elsewhere u . its great attraction is the way L splits journal movement into two con:onents squeeze and whirl, which enab!e a FulI orbi! to be caicu lated ver)/ rapidly with no reiterative caicuiations a each time step. For completeness the short bearing VEB )er hal centre orbit is included in the new %urvev af orbits in Fig 2a (supplementing those in Ref I“, and the variation fn minimum film thickness at different times tLroughot. the load cycle (defined by crank angle) is shown i: Fig 3. 148 983 Voi !8 N( J J f Experiments Qx o 3;0 Angular extent of oil feed, degrees Fig 10 Overestimate of flow QR using conventional Reynolds boundary conditions (intermain bearing, 1.8 litre engine) bearing and for a single oil hole. For a partially grooved main bearing an orbit relative to the bearing should be considered, whereas for a crank drilling and plain big end bearing one would consider an orbit relative to the crank pin. For a circumferentially grooved bearing any frame of reference would be suitable. The characteristics of feed pressure flow Qp, from equation 6 (Appendix 1) for the VEB bearing with a circumferential groove, are represented by the inset diagram in Fig 9(b). This shows the orbit superimposed on the lines representing values of constant flow. The predicted feed pressure flow is given in the main part of Fig 9(b). Actual flows from the 1.8 litre engine intermain bearing 6 with various oil feed arrangements (a single oil hole, a 180 groove and a full circumferential groove) all show that the predicted feed pressure flow (averaged over the operating cycle) gives a reasonable estimate of total flow. Similar conclusions were drawn by the author after he was privileged to have a preview of some National Engineering Laboratory reports on recent experimental work conducted by W L Cooke (See Experimental Support section). Total flow predicted from rigorous methods Improved predictive techniques and more rigorous programs are being developed and used. In many cases full 2-D solutions are being developed which take into account the groove shape, its size and position together with a dimensionless supply pressure parameter generally of the form: (Pffi7 co) (Cr/R ) 2 Such feed conditions are included in the two finite differ- ence solutions developed at Glacier, one using simple Reynolds boundary conditions and the other considering oil film history. These solutions give total flows defined as QR and QF respectively. The predicted total flow (QR) generally overestimates the flow, particularly for a single hole feed case. This is illustrated by the 1.8 litre engine results shown in Fig 10. The oil film history study of Jones 7 relating to the same 1.8 litre engine, with various bearing grooving arrangements, shows that the film history flow (QF) averaged over the load cycle gives excellent agreement with the measured flows from that engine. These rigorous solutions have also been applied to the VEB study case and the predicted total flows QR (conventional Reynolds boundary condition) and QF (with film history) are shown in Fig 1 1. It is of interest to see how QR gives an overestimate of flow, compared to QF, especially over the first 200 of crank angle position. Flows averaged through- 0.3- Conventional I O F Film history finite bearing / flow flow QR Ill (Pf =0) =0.193 v “ I A I i o , : 0 14t , i “, / t I Average 0 180 360 540 720 Crank angle, degrees Fig Comparison of predicted flows (VEB) TR IBOLOGY international 153 Martin - E,qg/ne bearhE design out the operating cycle (including those using rapid solu- tions, ie Q! and Qp) are shown on the right hand side of this figure. The idea developed so far, that the average feed pressure flow Qp (rapid solution), wtt give a good guide to the Tim history flow QF (rigorous solution) is supported by the closeness of these points (Fig i !); both of these solutions, in terms of average flows are generally consistenz with experimental trends, as will be seen later. Heat balance and friction in engine bearings The prediction of friction in dynamically loaded bearings is important for two reasons. Firstly, when coupled with the oil flow, it forms the reiterative heat balance for dete mining the operating viscosity or viscosities in the bearing. Secondly the prediction of friction (and therefore power loss) is important in its own right when looking for minimum energy loss. A comprehensive text showing the development of frictior: and power loss equations for dynamicaty loaded bearings is given in the appendix of a paper by Booker, Goenka and van Leeuwen 9 . It is very general and considers a free body analysis of the lubricant film. The equation for friction power (the rate of work done on the film) involved three terms: Power loss = (Jr :qR3 L/C) A,oAoo- e x Fo d0 + F (3) The last term is often negligible; it dominates where there is I a. 5Oi , I, t- - z5 i! /“ i f J 15 I o “ IO - 5 Constant viscosity l Viscosity calculated from 0.5 P,ex 0 Viscosity clculted from Pme ,I o-41 O.5. I o.,! 0 90 180 2_70 :560 450 540 6.30 720 Crenk angle 82 ,degrees Fig t2 Predicted performance considering pressure viscosity effects (VEB) (Pmax is the instantaneous maximum film pressure) little relative rotatmn, (eg squeeze fiim bearings). The first zerm generally predominates m ergine bearings and J( Tbds term is quoted extensively as part of the power loss equation, tt shotid be noted however, that for a fim exterlt (such as the short bearing Mobility method uses tiis verm is not simply halved, since for dynamical loaded bearings the load carrying (active) par of the film rare!;r extends from exactly hmax to the ,min positiotas. The heat balance is often used co predic a stogie efi?ctive: viscosity, found by considering the global effect of total heat generated by friction which is removed by 5e toal oil flow. A refinement on this, particularly for circumfbrentialiy grooved bearings, is to consider two v),scosities One toe- trois oii flow: which will be mostly from the coole thick film region, and the other controls load capacity and fric tion toss, which are meaniy inflenced by t29.e hotter thin lm region Other refinements involve the emperacure variaor throughou the bearing 202 and Jm pressure effects on yrs. cosity “-2 . This latter effect can be very significant, as skow for the VEB study case in Fig 12; for tMs exerdse the bear-. ing temperature was assumed cor.szan. Another importan= aspect, with the introductio of ron-Newtonian muRigrade oils, is the effect of shear rae on viscosity (also influerced by temperature) =a . (it is interesting to note hat the VEB study case *s continnally being used independently by others 2 ), fain bearieg load sharing The loads on a big end bearing are reiativeiy simple ,:o calculate, being based on the inertia of the reciprocating and the rotating components and on the gas forces imposed on the piston. The main bearing loads must react agais the big end loads, and traditionatly a staticaRy determinate system has been considered in which the crankshaft is - Static determinate Uneoupled . . . . . Idetermmae coupled z “ . Over I0 bearing arc 0c.= 7zo (sep Io) * n, arcs X j c./h. ),o - “10 Beori CR = Over 360 bearing arc - , / , SO(step I0) o X,o Fig 21 Friction factor - performance comparator for inten- sity of heat generated (FEB) TR I BOLQGY international 157 Martin - Engine bearing design on a horizontaI plane. The upper figure shows the first part of the process for a particular crank angle and journai posi- tion. The heights of the vertical bars shown are proportional to the frictional work done on each 10 arc of bearing surface. This process is repeated throughout the 720 of crankshaft rotation and a!t the friction values (height of vertical bars) are summed and averaged at each l 0 of bearing arc. The lower diagram in Fig 21 shows the resultant Eperimentol a Predicted 600 ._= E o 400- o 200 Gtcier /j experiment 0 C Feed pressure, ber 2 4 6 Fig 22 OrcumferentiaEy grooved bearing (NEL/VEB study ease).“ (a) experimental /ournaI orbit (NEL ), (b ) predicted journal orbit Glacier Metal), c) oit flow versus feed pressure crank pin IS F Orbff Oil hole axis Op 8 Ill c,P o.oool o.ool ( .:,_: predicted to increase let areas marked A). These positions are very dose to the measured negauve pressure regions in Fig 28. A further indication of the tendency towards film rupture and the possibb dove,or merit of these negative oil film pressures is shown i te rupture region map iv_. Fig 30. Again there is a remarkabb correspondence between Fig 28 and ig 30 in a negative pressure region. The third and final NEL report 4e in this series gave detais of measured oil flows into a big end :bearing of a Perkins 4.236 enNne. Measurements of the actual dynamic flow using constant emperature anemometry methods were made. detecting the variations in ow hroughout dan Ioacl cycle. This ogether with the detailed resuts iN tb. other two NEL reports give a valuable aid Jbr assessing predictive procedures. Speci