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关键词：: 校园电动车汽车设计汽车车辆校园电动车的设计汽车车辆设计

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摘要

校园电动车是近几年出现的一种新兴交通工具。校园电动车有着节能，环保，便捷等诸多的优点。因此本课题具有很强的现实意义和接近实际水平的设计要求，尤其是对机械设计部分和电气控制部分的设计。本毕业设计主要进行了电动车的机械部分的分析和设计，包括电动车的减速器、差速器、轴承的选择设计、各主要零部件的强度校核与计算、以及基本的机械传动部分的实现等。通过所设计好机械各部分或者整体部分的结构以及相关尺寸，利用制图软件CAD进行相关的配图的绘制以及其他零件图的绘制。根据电动车前进、后退、制动等基本控制要求，给出电动机的电气控制设计。

关键词: 校园电动车；交通工具；强度校核；机械传动

Abstract

Campus electric vehicle is the emergence of a new transport in recent years. Campus electric car has energy-saving, environmentally friendly, convenient, and many other advantages. Therefore, this issue has a strong practical and realistic level design requirement, especially the part of design for mechanical design and electrical control part.

The graduation project is mainly for the electric vehicle analysis and design of mechanical parts. Including electric reducer, differential, bearing choice design, the major components Strength check and calculation, and basic mechanical transmission part of the implementation and so on. Designed by a good mechanical parts or the whole part of the structure and related dimensions, Use graphics software related with CAD drawing and other component drawing.

According to electric cars forward, reverse, brake control and other basic requirements, electrical motor control design can be given.

Keywords: campus electric vehicles; transport; strength check; mechanical transmission

1绪论1

1.1概述1

1.2电动车的优势与发展2

1.3本设计的主要任务3

2 机械部分设计4

2.1概述4

2.1.1基本要求4

2.1.2 基本数据4

2.2 传动部分设计5

2.2.1减速器传动比计算5

2.2.2 齿型选择6

2.2.3 载荷计算6

2.2.4 齿轮材料选择7

2.2.5 齿轮强度计算8

2.3 差速器设计12

2.3.1 对称式圆锥行星齿轮差速器原理12

2.3.2 对称式圆锥行星齿轮差速器结构14

2.3.3 差速器齿轮基本参数选择14

2.3.4 差速器齿轮强度计算17

2.4 轴承选择与校核18

2.4.1 概述18

2.4.2 滚动轴承类型及代号19

2.4.3 滚动轴承选择21

2.4.4 滚动轴承约束设计22

3电动车电气控制设计24

3.1主电路24

3.1.1 H型双极模式PWM控制24

3.1.2控制电路25

3.1.3 SG3525的内部电路和参数26

3.2电动车电池设计方案27

3.2.1 电池槽27

结论28

致谢29

参考文献30

1绪论

1.1概述

校园电动车是近几年出现的并且不断发展日益增多的小范围内使用的交通工具,它的出现和广泛应用为校园内的师生提供了更为便利的交通，还可以作为校区的旅游和观光工具，目前在各大景区已经广泛使用。它有着诸多的优点，例如：首先，环保，电动车行驶零排放，不污染大气，是节能、环保的典范；第二，需求量大，一辆电动自行车一次充电能行驶30-50公里，有较大的市场需求；第三，操作简单，车速不高，每小时20公里左右，不会对其他人力自行车和行人构成威胁和安全问题；第四，维修简单；第五，用户白天使用，夜晚充电，续航能力很强，也不影响日常的工作和生活。该设计集机械和电力电子技术于一体，充分体现了节能、环保和方便实用等特点。通过对其的设计，能够使自身综合能力与设计创新的思维得到很好的锻炼。

电池电动车的历史。世界上第一辆电动汽车于1881年诞生，发明人为法国工程师古斯塔夫·特鲁夫，这是一辆用铅酸电池为动力的三轮车，而在1873年，由英国人罗伯特·戴维森用一次电池作动力发明的电动汽车，并没有列入国际的确认范围。后来就出现了铅酸、镍镉、镍氢电池，锂离子电池，燃料电池作为电力。电动车-行业前景电动车行业在中国崛起仅仅几年时间，在这短短的几年内，电动车行业由无到有，由零星分布到大范围普及，取得了高速的发展和长足的进步。由于不需要核心技术，进入门槛低，赢利空间大，短时间内大量企业将目光锁住电动车这个新兴行业。电动车产业的发展具有较强的地域性，一方面表现在生产，一方面表现在消费领域，而且这也是一个渐进的过程。经过十余年的发展，中国电动车行业从小到大，已经形成一个规模庞大的产业群，尤其是进入二十一世纪以后，整个产业呈现高速发展态势。2004年，中国电动车行业已有1000多家生产厂，年产量达675万辆。2005年，中国的电动车年产量达960万辆，市场保有量在1500万辆以上。2006年国内电动车产量达到近2000万辆，比上年增幅60%以上。2010年，中国轻型电动车的产销量将可能达到3000万辆，出口量将可能达500万-600万辆，实现工业产值700亿元，包括上下游带动产值的产业总体规模，将达1300亿元。我们在为这个行业快速发展而欣喜的同时也应看到，目前电动车行业的整体发展质量并不高，主要表现在厂家虽多但质量不佳。具有自主研发能力、上规模的大品牌很少，而大多数是一些靠模仿拼装、以低价运作的厂家，有些小厂甚至几个人、几把螺丝刀就能组装销售。

参考文献

[1]赵明生.机械工程手册.专用机械[M].机械工业出版社.1995: 77-89

[2]刘永波主编.电力电子技术[M].北京：机械工业出版社.2005: 34-42

[3]郑堤、唐可洪主编.机电一体化设计基础[M]. 北京：机械工业出版社，2006: 109-116

[4]周开勤. 机械零件手册第五版[M]. 北京：高等教育出版社 .2001:50-63

[5]成大先.机械设计手册·机械传动[M].北京：化学工业出版社.2004:23-42

[6]成大先.机械设计手册·机械制图·极限与配合[M].北京：化学工业出版社.2004: 98-67

[7]席伟光、杨光、李波主编.机械设计课程设计[M].北京：高等教育出版社.2003:88-100

[8]章宏甲、黄谊主编．液压传动[M]．机械工业出版社.2000.9:111 -132

[9]Jongsoo Lee. Passivity-based control of synchronous motors in mine hoist systems [J].Journal of Coal Science and Engineering.2001, 001(01):20-33

[10]ZHOU Xiuhua .Study in New Type Signaling Device for Mine Hoists [J].COAL MINE AUTOMATION, 2000,001(01):111-149

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内容简介：: 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 (coupled) solutions, differed significantly as shown in Fig 13. Welsh 26 has recently improved the crankshaft modelling, taking advan- tage of the fact that some of the substructure cheek ele- ments (ie, crank web and half of the adjacent crankpin and shaft) have common shapes along the crankshaft. Therefore only a few substructures are used as fundamental /Main journal face - / / d e numbers 1-58 inclusive) . /)k ,de 38 (centre of foce) /-Node 374 (centre of face Crank pin face (node numbers 337-374 inclusive) Fig 14 Finite element model of a basic superelement com- prising 48 solid elements (Perkins Engines Ltd) Fig 15 Cylinder block - basic substructure (Perkins Engines Ltd) Martin - Engine bearing design building units. This is further described and applied to a six cylinder engine in a paper by Welsh and Booker 8 . The author has had many discussions with Perkins Engines Ltd, who have expressed the following views on load cal- culations: Methods of main bearing load calculation in widespread use embody unrealistic assumptions and therefore cannot be viewed with confidence as a design aid. Many of the methods of dynamic load analysis are statically determinate and therefore neglect the effects of crank- shaft and engine stiffness. Alternative indeterminate methods reported in recent years 242728 do not employ realistic crankshaft or engine stiffness models. In a PhD thesis and forthcoming CIMAC paper Law 9 describes work undertaken at Perkins Engines to generate a computer based technique in which the structural and hydrodynamic equations which describe engine main bearings are sequentially solved. The structural equations are formulated in terms of influence coefficients derived from substructured finite element models of the crankshaft and cylinder block (Figs 14 and 15). The hydrodynamic operation of the main bearings is currently modelled by the Mobility method. The assum- ptions governing this method restrict its use to aligned, circumferentially symmetric, cylindrical bearings. To validate the program as a tool for the calculation of engine loadings, predicted crankshaft strains have been compared to those measured in a six cylinder high speed diesel engine. Typical predicted and measured strains are shown in Fig 16, a general agreement of 10% in strain range was achieved. 63.5 (NOM) t ; I00 o 700 _,oo., 1 / ran ,-zoo. o t/ Measured -500. Perkins indeterminate method . Statically determinate -400. -500- -600- i Fig 16 Predicted crankshaft strains compared to measure- ments; strain at web 9 (Perkins Engines Ltd) TRIBOLOGY international 155 Martin - Engine bearing design Load diagram 250r I Measured from engine 200 - .1_ x Rigid bearing theory / 15o .L= I %x-, Short Finite / / I00- / I / / X / 1 1 50 k z 0 60 120 180 240 300 360 Bearing angle , degrees Fig 1 7 Film thickness around bearing at the instant when the maximum inertia toad occurs ( VEB) Fig 18 Typical dynamic distortion of clearance boundao estimated from film thickness measurements (VEB) Oil film thickness at crank 0 1 2 I 4 5 0 -7 i Probe stations 5 Oil film thickness ot Instantaneous position : :. crank angle 120 of journal centre IoadCe . 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 ( ii .5 1 .,rence specs Flow mop o.os Kk_I_ / _ _ / , o.oJ :5l;:V/ (av) = 0.65 014 t.5 c Fig 23 Effect of oil hole position on predicted feed pressure flow (for NELl VEB data) overall friction factors, and gives a comparison oi: the accumulated frictional work done as particular positio.s around the bearing E xperimentN support A few years ago the National Engineering Laboratorms we:e very active m a joint program of research with Perkins Engines with the prhdpat aim of providing hproved design methods for engine bearings. W L Cooke formerly at NELo supervised the experimental work here, which was suppo; ted by the UK Departmem of Industry. One of de objecf5 of the joint project was e measuremen$ of dynamical1) loaded bearing performance and three NE L repor-s, writte,. by Cooke, are pending publication 4-4s _ All three reports are pertinent to the deve!oprnen of new predictive tech niques as they hetp to form an experimentai ramewerk on which to assess the merits of the various theories. The first report examines the effects of varying the ear- ing geometry and oil supply conditions. The work was carried ou on the NEL Engine Bearing Shnuiator. This consists essentially of a 2.5 inch diameter test bearing, carried in a massive yoke (to minimize bearing distortions) and loaded by four hydrau!ic rams The kydrau!ic rams are activated by electrohydrautic serve valves and the e!ectricai. input to these valves depends on the arrangement of diode pros in two 120 x 51 hole matrix boards, representmgthe vertical and horizontal load components. A manuaF feed- back loop from strain gauges on he rams permits fine tuning on the matrix boards to ensure at correct bearing loads are applied. Engine bearing conditions were simiated on te ng by matching the same values of (IN,/Pn) (R/C) (the length :o diameter ratio was not exactly simuiated) Polar load dia- grams t)om a variety of engines were used in the ess These included: Ruston 6VEB-X Mk III Ministry of Defence ASR i Perkins 6.354, and Mirflees KV!

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