机械专业毕业设计--外文资料翻译

毕业设计(论文)外文资料翻译 外文出处：MitsubishiHeavyIndustries,Ltd.TechnicalReviewVol.39No.1(Feb.202)附 件： 1.外文资料翻译译文；2.外文原文。 (用外文写)基于三维 Elasto 水力润滑理论的曲轴设计TakeroMakinoToshimitsuKoga长崎研究和发展中心，技术总部 通用机械和特种机车总部 高效率的要求造成了大量柴油机引擎曲轴的设计困难。当轴承油膜厚度不到几微米 时，由于轴承负荷而产生的变形量也仅为几毫米。本论文详细叙述了三维 Elasto 水力润 滑理论理论在 4 冲程柴油机引擎的曲轴设计上的应用。这些理论包括曲轴的变形和曲轴间 隙中油膜的产生原因。 绪论 近一个时期以来，内燃机的出口量有所增加，但其比重却在下降。这是因为，轴承在 恶劣的的环境下使用，大式轴承和主要的轴承连杆的机架变形在轴承的特征上产生重大影 响。为解决这一问题，三菱重工业有限公司(以下简称MHI)为这些动态轴承负荷开发了一 种应用elastohydrodynamic lubrication(EHL)原理的轴承特性预报方法，并且使用这种方法 来对MHI公司的大负载柴油机引擎进行设计和评估。 EHL 技术分析轴承表面弹性形变导致的油膜压 力，假设轴承刚体机构，既考虑轴承局部表面变形的 影响，同时又准确预测特征相对于传统的分析 此外，在这些年里, 三菱重工引进 EHL 技术分 析研究由于油膜压力而产生的轴承变形的油膜历史 记录，同时追踪轴承清除根据时间历史记录的油填充 比例来改善评估的准确性。 这份报告介绍了这一技术在大型连杆轴承上的 应用实例和对三菱重工的四冲程柴油发动机的主要 影响。理论 2.1 基本公式 图 2 显示了这份论文中采用坐标系统。 影响油膜压力的参数p 可用方程(1)来表示。)()(2(123 thhUphe p + - - r q r r a (1) 当方程(1)和下面的力平衡组合成一个相对于时间 t 的联立方程组，这样一来就可以 得到轴中心局部和槽油膜厚度的信息。0)cos( = - W - W xWdp q (2) 图 1 坐标系统0)sin( = - W - W yWdp q (3) 由于开始的几何间隙，轴的偏心率和弹性形变，所以公式(4) 这样来表示油膜厚度h。 （4） 其中： ：粘度压力系数xa：x轴的偏心率ya：y轴的偏心率rc：轴承半径间隙xe：X方向的离心率ye：Y方向的离心率h：油膜厚度L：变形N：引擎速度 ：面积p：油膜压力 ：油填充比例 ：轴承圆周坐标t：时间U：滑动速度xW：X方向负荷yW：Y方向负荷X：X轴方向坐标Y：Y轴方向坐标Z：Z轴方向坐标 2.2 分析技术 2.2.1 考虑油膜历史记录曲线的EHL技术分析 我们开发了一个基于JONES提出的油膜历史记录曲线概念的EHL分析技术，来考虑 在轴承间隙中的油的运动。三菱重工的常规EHL技术分析，计算假设在整个轴承表面覆盖 润滑油的情况下的压力分布，替代由于计算周围压力获得的负压力区域，并且把油膜断裂 边界视为分界线。在以这个边界为条件下，油膜断裂区域的流动连续性不能被满足。另一 方面，EHL分析技术研究随着油填充比率和时间的推移而变化的油膜历史记录曲线，则显 示流动连续性满足。LpZaeZaech xyyxr + + + + + = q q sin)(cos)(至于受到波动负荷的轴承,如发动机轴承，在轴承上实际油膜压力增长受限制的区域， 在下文中EHL分析可以得出比常规EHL分析更高的压力结果。这是主要用于检验三菱重 工的大型船用柴油机引擎的实际尺寸的。EHL技术分析油膜压力历史记录曲线是作为一个有 益的分析工具来用于设计和评价的。 2.2.2 计算方法 油膜压力P和轴偏心率 xe 、ye 的结果可以从 联立方程(1)到(4)中获得。由于方程(1)和油膜 压力的非线性关系，我们用“牛顿-拉斐尔”方 法 来 确 定 它 们 。 我 们 用 有 限 元 方 法 (FEM)(Galerkin方 法 ) 进 行 数 字 计 算 ， four-point等参数原理被看做是原理内容和线 性方程系统的数字化解决方法。为了测定油膜 断裂边界，我们改进并使用了适用于油膜历史 记录曲线的有限元运算方法的Kumar技术。图2 显示的是流量计算。 3个案研究 3.1 连杆头轴承 例如 S3l 发动机的大型连杆头轴承，我们 比较刚体分析和 EHL 技术分析、 并且比较了 EHL 技术分析的油膜历史记录曲线和传统的 EHL 技 术分析。 此外，我们测算了连杆头轴承在轴承特性 上对曲柄销外形的影响。 表格 1 所示为轴承规格，图 3 显示了影响 轴承和用于计算执行 3DFEM 模型的负载向量。 表格 1 S3l 引擎大型连接头轴承的尺寸参数 图 2 流量计算 （a）轴承负载 （b）连杆头 FEM 模型 图 3 轴承工作状况的计算当考虑到弹性形变，则计算油膜压力的减少 量和油膜厚度的增加量。如果考虑到油膜历史曲 线，则轴承上油膜压力的实际受力面积的发展受 限制，并且压力最大值会变的更大。 3.1.1 技术分析对比 在图 4(a)里显示了油膜压力随着时间改变而 产生的最大变化量。几乎在所有的时间点，取决 轴承直径 轴承宽度 径向间隙 杆长 冲程 引擎速度 润滑油粘度 48mm 21mm 0.025mm 145mm 78.5mm 3600rpm 10cP180200160140120100806040200 180 360 540分析EHL考虑油膜历史的 分析EHL的 常规 分析 刚体 曲柄角度(deg) (a)最大变化，油膜压力的时间变化曲线 (b) 轴心轨迹 轴承的圆周坐标 (deg) 轴承的圆周坐标(deg) (c)轴承中心部分的油膜压力(曲轴角度10度) (d)轴承中心部分的油膜压力(曲轴角度10度) 图 4 大端轴承特性分析结果于刚体分析的油膜压力高于由 EHL 分析获得的数值。这显然表明，比如说，相对于瞬时时 间，曲轴角度大约在 10 度左右的地方负载相当大。当取决于刚体分析的油膜压力是 180 （MPa）时，取决于 EHL 分析的压力是 133（MPa）。图 4 (c)和(d)显示的是曲柄角度在 10 度时轴承中心部分的油膜厚度分布状态和油膜压力分布状态。 由于考虑到弹性变形， 由 EHL 技术分析得到的油膜厚度相比于由刚体分析得到的几乎差不多大，并且在压力区域几乎一 致。比起刚体分析，由 EHL 分析给出的压力分布区域在圆周方向更宽，并且显示出较低的 油膜压力最大值。显然，从图 4(b)显示的轴中心轨迹和 EHL 技术分析表明，轴中心是明显 偏离曲轴间隙的。 在图 4(a)中显示了，在曲轴角度约 250 度时，由 EHL 技术分析决定的油膜压力最大值 不同于由传统 EHL 技术分析所决定的数值。当轴向油膜断裂面的一边运动时，曲柄角度调 整符合从上部金属到下部金属负载的转变。 图 4（e）显示的是油膜压力分布状态和曲柄角度在 250 度时的油填充比例。从这个图 表上明显看出，考虑油膜历史曲线的分析显示出，轴承正压力区域的发展由于油量的缺乏 而受限制，并且给出相比于传统 EHL 分析所得到的更高的油膜压力。 考虑油膜历史的 EHL 分析 常规 EHL 分析 压力 分布 油比 例 (e)压力分布在考虑油膜历史的 EHL 分析和不考虑油膜历史分析的情况下的差异(在曲轴角度 250 度) 3.1.2 曲柄销外形在轴承特性上的影响我们评估曲柄销外形在大型连杆头轴承特性上的影响。图 5 中显示的是我们研究的三 种曲柄销外形，即(1)直线型，(2)桶形，(3)曲线形。图 6 显示的是油膜压力最大值和油 膜厚度最小值的分析结果。这表明，在直线外形的曲柄销(1)显示出更大的油膜厚度，并 且适合于曲轴的操作条件。 3.2 主 要 影 响 这一方法不仅可以适用于分析大型连杆头轴承，而且也适用于大部分轴承和小型连杆 头轴承。下面是个一台 S6r 引擎的 4 号主轴承的分析实例。表格 2 显示了轴承的规格，图 7 显示了包括轴承负载和主要轴承的发动机框架结构的 FEM 模型。 直线形 桶形 曲线形 图5 曲柄销外形的形式 曲线形 桶形 直线形 图6 由于销的几何误差引起的曲轴特性差异 （a）曲轴负载 （b）发动机框架结构的FEM模型 图7 主要影响的计算情况表格 2 S6r 引擎四号主轴承的尺寸参数 轴承直径 轴承宽度 径向间隙 引擎速度 润滑油粘度 油槽 140mm 53mm 0.07mm 1800rpm 10cp 水平向和侧面，45，100 毫米范围 图 8(a)所示的油膜压力最大值的时间变化历史记录曲线，显示了刚体分析评估的压力 值高于 EHL 分析的结果。图 8(b) 明显显示了曲柄角度在 245 度时的压力分布状态，考虑 了油膜历史记录曲线的 EHL 分析认为，轴承上由于润滑油的不足而导致油膜压力发展受限 制,并且得出比常规 EHL 分析结果更高的油膜压力数值。因此,有人认为，主要轴承的分析 显示了和大型连杆头轴承分析相类似的趋势。 考虑油膜历史的 EHL 分析 常规 EHL 分析 压力 分布 油比 例 (b)压力分布在考虑油膜历史的 EHL 分析和不考虑油膜历史记录分析的情况下的差异(曲轴角度 270 度) 图8 轴承主要特性分析的计算结果 4结论 考虑油膜历史的 EHL 分析 常规 EHL 分析 刚体分析 曲柄角度(deg) (a)最大变化，油膜压力的时间变化曲线EHL分析和研究油膜历史记录曲线的EHL分析是做为一种能够改善发动机曲轴系统可靠 性的先进技术来提出的。 为了给轻型高功率引擎的开发设计轴承，我们必须用上述评估技术来保证高度的可靠 性，并且提高三维CAD设计系统化技术连接的便利性。这项开发的部分是与Truck&BusResearch、开发中心和三菱汽车公司合作的。Mitsubishi Heavy Industries, Ltd.Technical Review Vol.39 No.1 (Feb. 2002)16Crank Bearing Design Based on 3-D Elasto-hydrodynamicLubrication TheoryTakero Makino*1Toshimitsu Koga*2The demand for high efficient diesel engines makes ample design of crank shaft bearings difficult. Deformationdue to bearing load is in millimeters, while bearing oil film thickness is less than several microns. This paper detailsthe application of 3-D elasto-hydrodynamic lubrication theory to crank shaft bearings for 4-stroke diesel engines. Thetheory includes bearing deformation and oil film history in a bearing gap.XYBearingShaftFig. 1 System of coordinates*1 Nagasaki Research & Development Center, Technical Headquarters*2 General Machinery & Special Vehicle Headquarters1. IntroductionRecently, the output of internal combustion engineshas been increased and their weight has been reduced.As the result of this, bearings are used under severeroperating conditions, and deformation of housings ofconnecting rod big-end bearings and main bearingshave significant influence upon the bearingcharacteristics. To solve this problem, MitsubishiHeavy Industries, Ltd. (hereinafter referred to asMHI) has developed a bearing characteristicprediction method(1)applying the elasto-hydrodynamic lubrication (EHL) theory for thesedynamically loaded bearings, and used the method todesign and evaluate MHIs large-bore diesel engines.The EHL analysis coupling oil film pressure withelastic deformation of bearing surface enables toconsider the influence of local bearing surfacedeformation and evaluate the quantitative bearingperformance with remarkably improved accuracy ofbearing characteristic prediction as compared to theconventional analysis assuming bearings to be rigidbodies.In addition, in these years, MHI has introducedthe EHL analysis considering oil film history thatcouples oil film pressure with bearing deformationwhile tracing the oil filling ratio in bearing clearanceaccording to the time history to improve theevaluation accuracy.This report introduces examples of application ofthis technique to connecting rod big-end bearings andmain bearings of MHIs 4-stroke diesel engines.2. Theory2.1 Basic equationsFig. 11 shows the system of coordinates used in thisreport.Reynolds equation that controls the oil filmpressure distribution p is expressed as eq. (1).(1)When the equation (1) and the following equationsof force equilibrium are combined into a simultaneoussystem of equations with respect to time t, informationon shaft center locus and oil film thickness can beobtained.(2)(3)The oil film thickness h is expressed by eq. (4) inconsideration of the initial geometrical clearance,misalignment of shaft, and elastic deformation.(4)where,: Viscosity pressure coefficientx: Misalignment around X axisy: Misalignment around Y axisc r: Bearing radial clearancee x: Eccentricity in X directione y: Eccentricity in Y directionh : Oil film thicknessL : ComplianceN : Engine speed: Areap: Oil film pressureMitsubishi Heavy Industries, Ltd.Technical Review Vol.39 No.1 (Feb. 2002)17: Oil filling ratio: Bearing circumferential coordinatet : TimeU : Sliding velocityW x: X direction loadW y: Y direction loadX: Horizontal coordinateY: Vertical coordinateZ: Axial coordinate2.2 Analyzing technique2.2.1 EHL analysis considering oil film historyWe developed an EHL analysis considering themovement of oil in the bearing clearance based on theconcept of oil film history proposed by Jones(2). MHIsconventional EHL analysis calculated the pressuredistribution on the assumption that the wholebearing surface is covered with lube oil, replacedthe negative pressure region obtained as the resultof the calculation with the ambient pressure andregarded the boundary of the region as the oil filmrupture boundary. Under this boundary condition,flow continuity on the oil film rupture boundarycannot be filled. On the other hand, the EHLanalysis considering oil film history follows the oilfilling ratio with time so that the flow continuity isfilled.For bearings that receive fluctuating load, such asengine bearings, the bearing area where positive oilfilm pressure can develop is restricted, and, therefore,the EHL analysis may show higher pressure than thatobtained by the conventional EHL analysis. This wasverified by actual measurement on the main bearingsof MHIs large-bore marine diesel engines. The EHLanalysis considering oil film history is used for designand evaluation as a useful analysis tool.2.2.2 Calculating methodThe oil film pressure p and shaft eccentricities exand ey can be obtained by the simultaneous equations(1) to (4). Since the equation (1) is nonlinear withrespect to the oil film pressure, we used Newton-Raphson method to determine them. We used thefinite element method (FEM) (Galerkin method) fornumerical calculation, four-point isoparametricelements as elements and Skyline method fornumerical solution of the linear equation system. Todetermine the oil film rupture boundary, we improvedand used the technique of Kumar, et al.(3)that appliesthe oil film history algorithm to the finite elementmethod. Fig. 22 shows the flow of the calculation.3. Case study3.1 Connecting rod big-end bearingUsing the connecting rod big-end bearing of an S3Lengine, we compared the rigid body analysis and theEHL analysis, and compared the EHL analysisconsidering oil film history and the conventional EHLanalysis.Moreover, we evaluated the influence of the shapeof the crank pin on the bearing characteristics of theconnecting rod big-end bearing.Table 11 shows the specifications for the bearing,and Fig. 3 shows the load vector that affects thebearing and the 3D FEM model used to calculate thecompliance.StartReading of input dataSetting of initial valuesSetting of oil film rupture boundaryAssumption of journal eccentricity and translation speedCalculation of oil film thickness distributionFind non-ruptured oil film regionCoefficient matrix of Newton Raphson method Calculation of correction values of pressuredistribution and eccentricityJudgment of convergence of load and eccentricityJudgment of convergence atcurrent crank angleJudgment of convergence of 1 cycleOutputEndCalculation of changein oil filling ratioRenewal of time stepTable 1 Dimensions of S3L engine connecting rod bigg-end bearing Bearing diameter 48 mm Bearing width 21 mm Radial clearance 0.025 mm Rod length 145 mm Stroke 78.5 mmEngine speed 3 600 rpm Lube oil viscosity 10 cP Fig. 2 Flow of calculationMitsubishi Heavy Industries, Ltd.Technical Review Vol.39 No.1 (Feb. 2002)18-5 000-25 00025 000-50 0005 00000XYZ360o300o120o240o60o0o(a) Bearing load (b) FEM model of big-end180o720540360180801001201401601802006040200(a) Change in max. oil film pressure with timeMax. oil film pressure (MPa)Crank angle (deg): EHL analysis considering oil film history: Conventional EHL analysis: Rigid body analysis-0.0500.050-0.0250.0250.000-0.05 -0.0 25 0.0250 0.05(b) Shaft center locus: EHL analysis considering oil film history: Conventional EHL analysis: Rigid body analysis3.1.1 Comparison of analyzing techniquesFig. 4(a) shows the change in the maximum oil filmpressure with time. At almost all timing points, theoil film pressure determined by the rigid body analysisis higher than that obtained by the EHL analysis. Thisis apparently indicated, for example, at the explosiontiming at a crank angle of about 10 where the loadis relatively large. The maximum oil film pressuredetermined by the EHL analysis is 133 (MPa), whilethat determined by the rigid body analysis is 180(MPa). Figs .4 (c) and (d) show the oil film thicknessdistribution and the oil film pressure distribution inthe bearing center part at a crank angle of 10. Inconsideration of the elastic deformation, the oil filmthickness obtained by the EHL analysis is larger ascompared to that obtained by the rigid body analysisand almost uniform in the pressure region. Comparedwith the rigid body analysis, the EHL analysis givespressure distribution wider in the circumferentialdirection and shows lower maximum oil film pressure.As is evident from the shaft center locus shown inFig. 3 Conditions of calculation for bearingChange in bearing load and 3D solid model for calcula-tion of deformation are shown.Results of analysis of characteristics ofbig-end bearingWhen the elastic deformation is taken into con-sideration, the calculated oil film pressure de-creases, and the calculated oil film thickness in-creases. When the oil film history is taken intoconsideration, the bearing area where positive oilfilm pressure can develop is restricted, and thepressure peak becomes higher.608010012014016020018040200 90 180 270 360(c) Oil film pressure in bearing center part (at crank angle of 10o )Oil film pressure (MPa)Circumferential coordinate of bearing (deg): Conventional EHL analysis: Rigid body analysis1020304050600 90 180 270 360(d) Oil film thickness in bearing center part (at crank angle of 10o )Circumferential coordinate of bearing (deg): Conventional EHL analysis: Rigid body analysis1010(e) Difference in pressure distribution between EHL analysis considering oil film history and EHL analysis without considering oil film history (at crank angle of 250o )EHL analysis considering oil film historyConventional EHL analysisPressure distribu-tionOil filling ratioFig. 4Mitsubishi Heavy Industries, Ltd.Technical Review Vol.39 No.1 (Feb. 2002)19Fig. 4(b), the EHL analysis indicates that the shaftcenter is remarkably eccentric over the bearingclearance.In Fig. 4(a), the maximum oil film pressure at acrank angle of about 250 determined by the EHLanalysis considering oil film history is different fromthat determined by the conventional EHL analysis.This crank angle timing corresponds to the point ofshift of load from the upper metal to the lower metal,where the shaft moves toward the side having an oilfilm rupture.Fig. 4(e) shows the oil film pressure distributionand oil filling ratio at a crank angle of 250. As isevident from this figure, the analysis considering oilfilm history shows that the bearing area wherepositive oil film pressure can develop is restrictedowing to the shortage of oil, and gives higher oil filmpressure as compared to that given by theconventional EHL analysis.3.1.2 Influence of crank pin shape on bearingcharacteristicsWe evaluated the influence of the crank pin shapeon the characteristics of connecting rod big-endbearings. We examined three crank pin shapesshown in Fig. 55, i.e. (1) straight, (2) barrel shape,and (3) hourglass shape. Fig. 66 shows the results ofanalysis of the maximum oil film pressure andminimum oil film thickness. It is revealed that thestraight crank pin (1) gives higher oil film thicknessand is suitable for the operating conditions of thebearings.3.2 Main bearingThis analyzing technique is applicable not only tobig-end bearings, but also to main bearings and small-end bearings. An example of analysis of the No.4 mainbearing of an S6R engine is shown below. Table 2shows the specifications for the bearing, and Fig. 7shows an FEM model of the engine frame includingthe bearing load and main bearing.The time history of maximum oil film pressurez=1z=-1z=1z=-1z=1z=-12120015010050064 46012345220Max. oil film pressure (MPa)Hourglass shapeStraightBarrel shapeTable 2 Dimensions of S6R engine No. 4 main bearing Bearing diameter 140 mm Bearing width 53 mm Radial clearance 0.07 mm Engine speed 1 800 rpm Lube oil viscosity 10 cP Oil groove Horizontal and lateral, 45, 100 mm wide -50 000 50 0000100 000-100 000-100 000100 000-50 00050 000XYZ300o120o240o180o60o0o(a) Bearing load (b) FEM model of engine frameshown in Fig. 8 (a) indicates that the rigid bodyanalysis estimates the pressure higher as comparedto the EHL analysis. As is evident from the pressuredistribution at a crank angle of 245 shown in Fig. 8(b), the EHL analysis considering oil film historyrestricts the bearing area where positive oil filmpressure can develop owing to the shortage of lubeoil and gives higher oil film pressure than thatobtained by the conventional EHL analysis.Accordingly, it is considered that the analysis of mainbearings shows a sim