資源目錄里展示的全都有，所見即所得。下載后全都有,請(qǐng)放心下載。原稿可自行編輯修改=【QQ：401339828 或11970985 有疑問可加】 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paper highlights the more important factors to be considered, and relates them to recent advances, both published and unpublished, throughout the world.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 & Co (Publishers) Ltd 147 Mair - Engine bearing design I i ,/ / I ie aiming for fewer assumptions data presentation for better understem.ding of results better prediction of operating conditions (load sharing, heat balance) experimental verification. Progress in each of these categories is very importam and each section complements the others. With the need to conserve energy and with fuel economy a major issue, many engines are now being designed with higher power to weight ratios The resultant effects on bear- ings are reduction in bearing size, higher specific loads and the use of lower viscosity oils. All these changes bring the Simplified and quick method Many data oresentation techniques shown in this pape; 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 3 A second part of Bookers work was to produce a clearance circle film pressure map 2 giving the ratio of the maximum hydrodynamic pressure to the specific load at any point in the clearance circle. The inset diagram in Fig 4 shows the clearance circle film pressure map with the VEB orbit superimposed. Note that this orbit is not plotted relative to space - the conventional method - but on a clearance map which is effectively being moved in an angular sense throughout the cycle, such that the direction of the applied load is always downwards. This is an important and valuable technique when using the Mobility method. The maximum oil film pressure is obtained from these relationships and Nomenclature Cr radial clearance, m D bearing diameter, m hmi n minimum film thickness, m e eccentricity vector F force vector JlOO f o2 (1 +ecosO) -1 dO 0 L bearing length, m M Mobility, dimensionless Pf oil feed pressure, N m-: Pmax maximum film pressure, N m-2 Pn specific load (W/LD), N m -2 QF oil flow considering film history, m 3 s - (rigorous solution) QH hydrodynamic flow, m 3s-1 (rapid solution) Qp feed pressure flow, m 3 s -1 (rapid solution) QR flow not considering film history, m 3 s - (rigorous solution) Qx flow from experiments, m 3 s-1 R shaft radius, m rl dynamic viscosity, Ns m-2 e eccentricity ratio, dimensionless k friction factor 0 angle of oil hole from centreline CF (see Fig 23) co and co are functions of journal and bearing angular velocity 0.5 0.4- 0,3- G .5 E 0.2- 0.1- F 1.875 ,5: : -)o ;,?., .-. ,2/0,1 0.001 , mL_ o 90 &o s;o 5 ,o Crank angle, degrees Fig 3 Short bearing film thickness ratio (VEB) do 720 Martin - Engine bearing design its variation throughout the load cycle is shown in the main part of Fig 4. At GEC in the UK Ritchie n developed a new semi- analytical method for predicting the journal centre orbit; it uses an easily obtained optimized short bearing solution which has improved accuracy at high eccentricities over the standard short bearing method; the orbit of the VEB big end bearing is shown in Fig 2(b). This looks very simi- lar to a general finite bearing orbit and apparently only took 16 seconds to run on an IBM 370/145 computer (several years ago). The minimum oil film thickness of 0.0033 mm (0.00013 inches) is compared in Table 1 with values from other sources (including the results of a GEC finite bearing program using the stored data approach - see next section). It is seen to be within the scatter band of the more rigorous finite bearing methods, but still main- tains the advantage of a rapid solution. The minimum oil film thickness during a complete cycle of operation is one of the most significant parameters on which to judge bearing performance. It is generally used as a comparator and represents a major factor in relating predicted performance with existing bearing experience on similar type engines. It is difficult to give precise values of minimum film thickness at which bearing damage might occur, as other factors such as high bearing temperature, misalignment, inadequate oil feed arrangements and adverse environmental conditions will all have an effect. Booker ll gives some guidance on danger levels for film thickness in connecting rod bearings (for use with short bearing predic- tion methods). Finite bearing theories Using a finite element method (fern) to solve the finite bearing theory, General Motors Research Laboratories 2 have the ability to consider different shapes of bearing and also to allow for the presence of grooving. For a plain cir- cular bearing GM have successfully curve-fitted basic data from their fem bearing model, and used this to develop a rapid method, typically reducing computational time from hours to seconds. Both methods have been applied to the Prolix 1.667 2 Pn 2.5 40 , 3 ;50 / l/i/: - 25 m = 2o _E 15 .E. E 1 I0 e 5 i i I 1 I I i 0 90 180 270 560 450 540 650 720 Crank angle, degrees Fig 4 Short bearing maximum film pressure (VEB) TR I BOLOGY international 149 MartL, . Engine bearing design Ruston VEB big end, and Figs 2(c) and (d) show the journal centre orbit for the finite element program and curve-fit program respectively. These two orbits look very similar, Nthough there was a remarkable saving in compu- tational time for the curve-fit program. Film thickness ratio and maximum film pressure from the two me,hods are com- pared in Figs 5(a) and (b). Also note that the film pressure from the short bearing theory (Fig 4) is very similar to that from the finite bearing fern theory (Fig 5b)o Many establistments now have finite element or finite difference 2-D solutions capable of allowing for the effect of oil feed features on hydrodynamic pressure generation , The %tandard VEB study case, with its circumferential groove, is not suitable for illustrating such effects, so instead the intermain bearing of a 1.8 itre gasoline engine will be used. The lead diagram is shown in Fig 6 and further dat can be found in References 6 and 7o The orbits in the torc diagram of Fig 7 show the film thickness reduced locaily as a result of the presence of an oil hole. tt should be noted however, that the smallest film thickness during the cycle may not necessarily be impaired A design method has been developed at the Glacier Metal Co whi.ch altows, in a more complete way, for the effects of feed features in the bearing o It considers these effects to fl into two categories. The first relates to the deh- :nentai effect of the developed pressure region passing over the oil feed region (hole, groove etc) of the bearing The second involves the study of oit transport within the bear- o .4 7! Curv fit program 0.5 Finite element orcgrarr . : ! o.i-, , 40- 90 t80 270 560 450 4, 6.30 720 g_ 50 m o E = 20. Curve fit program Fmffe element program . /m / / I / t/ t/ W ,j 1 r C 90 80 70 360 450 540 630 720 b Cro Ongledegrees Fig 5 General Motors rapid curve fit program compared ro rigorous fern program ( VEB: (a) dimensioMess film thick- ness, (b maximum film pressure ing oii film, and takes into accoum the deleterious effect when the oi1 fi!m extent is depleted due to insufficient eli being available to filI the ioad carrying area of the bear ind. This second category is sometimes referred to as cil fi Nstory. eli fIm history Much of the fundamental work on eli film history and or.; film. boundaries m dynamically loaded bearings was pione.:rc at the UZK National Engineering Laboratory by the iate A.Ao Milne s16 , whose untimely death left a vod in the knowledge of tNs very specialized fbldo Milne% apFroach considered an everchanging an_d me,and mesh )a:tem o mach .:he film boundaries. Arxther method developed at Glacier by Jones considered :,( 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( z 2r film fie. one that is active over the full circmfere,ce ; the bearing) this erm becomes. 2re (rgR3 LooP /C)/( i = uP) 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 . i I o = S i j -.f / Fig t3 Computed loads centre mum bearing, o,r cylinder engine (Booker/Stkkier, 2 982) t54 June 1983 Vol IG No 3 treated as if it were pin jointed at the axial mid-position of each main bearing. Effectively this means that any main bearing can be influenced by big end loads only in immedi- ately adjacent bays. In practice however both crankshaft and crankcase have finite stiffness, so that very complex interactions can be set up throughout the entire engine. Improved crankshaft mode/ling Many researchers have now attempted to take into account engine flexibility, and to couple this with the bearing analy- sis. In recent years work at Cornell University (USA) and Perkins Engines Ltd (UK) has been progressing in this field independently. At Cornell Unviersity, Stickler 24 carried out a feasibility study using simple beam type elements to represent the crankshaft in the structural analysis. Booker and Stickler 2s applied this procedure to a 4 cylinder inline automotive engine using a rigid crankcase and short bearing theory. The computed centre main bearing loads, using the static deter- minate (uncoupled) and indeterminate