【作者】 刘娜；

【导师】 李国祥；

【作者基本信息】 山东大学 ， 热能工程， 2012， 博士

【摘要】 缸套—活塞环摩擦副与主轴承是柴油机的关键摩擦副,其摩擦润滑性能对整机性能有很大影响。本文通过理论分析、数值计算与试验技术相结合的方法,研究了柴油机缸套—活塞环摩擦副与主轴承两大关键摩擦副的摩擦润滑性能,主要研究工作有：根据气体流量方程与质量守恒定律建立了活塞环的漏气计算模型,就所分析的某直列六缸柴油机进行了漏气分析,为润滑分析提供载荷条件与边界条件；然后基于平均雷诺方程,建立了缸套—活塞环摩擦副润滑磨损性能分析的计算模型,所建立的模型综合考虑了粗糙度对润滑性能的影响,考虑了润滑油的变密度和变粘度效应；推导了平均雷诺方程的差分离散格式,给出了求解平均雷诺方程的方法；并就所分析的柴油机进行了润滑磨损分析,分析结果表明：该柴油机的缸套—活塞环摩擦副在压缩上止点附近的一段曲轴转角时间内,膜厚比小于4,处于混合润滑状态,其余曲轴转角时间内膜厚比大于4,处于流体动力润滑状态；最大油膜压力出现在压缩上止点附近,为燃烧室内最高压力的两倍多；最大摩擦力同样出现在压缩上止点附近,主要是由于油膜厚度较小,摩擦副两表面产生固体接触而导致,其余时刻内摩擦力数值较小。最大摩擦功耗出现在压缩上止点附近。缸套的粘着磨损深度最大值位于第一道环上止点处。根据所建立的缸套—活塞环润滑计算模型,考查了缸套—活塞环摩擦副润滑性能的影响因素——润滑油粘度、润滑油温度、摩擦副表面粗糙度、活塞环结构参数等；并且考虑润滑油的非牛顿效应,以幂律流体为例,建立了适用于缸套—活塞环润滑计算的润滑模型,给出了平均雷诺方程的差分格式,进行了求解。缸套—活塞环润滑计算结果表明：减小缸套—活塞环摩擦副的表面粗糙度,最小膜厚比呈增加趋势,处于混合润滑区域的几率减小,润滑状况得以改善；最大摩擦力和最大摩擦功耗都有所降低。润滑油粘度增大,可以改善上下止点处的润滑状况,但其余曲轴转角时刻的摩擦力却增加；随润滑油温度的降低,摩擦副的润滑形式处于混合润滑与边界润滑的几率减小,干摩擦几率减小最大摩擦力有所降低,但全膜润滑时摩擦力增大；从整个循环工况的总体而言,除去膜厚比较小的时刻,其余时刻随润滑油温度的降低,摩擦功耗增大；所以从润滑性能角度而言,润滑油要保持一定温度,不能过高或过低。随活塞环桶面高度降低,活塞环侧面轮廓曲线平缓,有利于挤压效应的形成,最小油膜厚度在上下止点位置处有所增加,其余曲轴转角时刻由于动压效应的减弱,最小油膜厚度有所降低；在上下止点位置附近,摩擦力降低,其余曲轴转角时刻内,摩擦力增大；表明降低活塞环桶面高度可以改善上下止点位置处的润滑性能,但却减弱其他曲轴转角时刻的动压效应,因而要合理选取活塞环桶面高度值,使之发挥最好效应。活塞环环高增加,增大了摩擦副的接触面积,使单位面积承载力降低,增加了膜厚,改善了润滑性能,但同时因为接触面积加大,流体阻力有可能增大而导致摩擦力与摩擦功耗增大,所以应综合考虑,选取合适环高；忽略粘度随压力的变化,最小油膜厚度的计算结果比不忽略粘度变化时小最大摩擦力与最大摩擦功耗比不忽略粘度变化时高。当考虑润滑油的非牛顿效应时,幂律指数不同,油膜压力分布形状保持一致,最大油膜压力出现的位置一致,变化趋势一致。但油膜压力的数值随幂律指数的增大而增大,即,幂律指数越大,油膜承载能力越强。对于幂律指数为0.9的润滑油而言,改变活塞环环高与摩擦副表面粗糙度,对膜厚比与油膜压力的影响与润滑油为牛顿流体时影响基本一致。试验分析是缸套—活塞环摩擦磨损特性研究的重要方面,对于模拟试验来说,重点和难点是如何准确模拟真实情况。传统的磨损试验—般采用摩擦副零件材料的部分切片作为试样进行研究,并且模拟条件也与柴油机真实工作环境有—定差距,本文从摩擦学系统分析的角度出发,模拟试验中采用更为精准的模拟准则,选用真实的完整缸套—活塞环进行配对,模拟条件中加入温度条件,以及真实工况中存在的酸性环境等,在专用缸套—活塞环往复摩擦磨损试验机上进行摩擦磨损过程的模拟试验,考察压缩上止点处缸套与第—道活塞环的磨损情况,并且为计算柴油机实际工作循环中缸套的磨损量提供当量粘着磨损系数。本文通过测量缸套、活塞环的磨损量数据以及摩擦磨损过程中的实时摩擦力数据等,进行了4对不同材料的缸套—活塞环摩擦特性的对比分析。结果表明：该柴油机缸套—活塞环的最佳配副为等温淬火气缸套-喷钼活塞环配副。在分析主轴承的润滑特性时,弹性变形因素不能忽略,需要在润滑方程组中加入弹性变形方程共同求解。本文采用多体动力学与主轴承弹流润滑相耦合的方法,对柴油机进行主轴承的润滑性能研究。计算获取的主要参数有主轴承载荷、主轴承轴心轨迹、最小油膜厚度、最大油膜压力以及摩擦功耗等。研究结果表明：主轴承载荷中,沿气缸垂向方向(Z向)的载荷大于活塞主副推力面(Y向)的载荷；第2、3、5、6主轴承的Z向载荷峰值较大；第1、4、7主轴承的Z向载荷峰值较小；第1、7主轴承的轴心轨迹类似,轴心的主要运动方式为圆周运动,很少有高速向心运动,穴蚀现象发生几率较低；第2主轴承的轴心向心运动几率也不高,因而润滑性能也较好；第4主轴承的轴心轨迹大部分落在60°CA-240°CA区域,造成此区域油膜厚度很小,容易造成磨损；第3、5、6主轴承的轴心轨迹存在较为明显的向心运动,容易产生穴蚀。第1主轴承的最小油膜厚度最小,需要引起特别注意；最大油膜压力出现位置与Z向轴承力峰值位置基本一致。而最小油膜厚度位置处与轴承力峰值载荷位置不一定一致；摩擦功耗的变化趋势与最大油膜压力的变化趋势基本相同,最大摩擦功耗出现的位置与轴承力峰值位置基本一致。本文还对单缸熄火时的主轴承润滑性能进行了研究。机体的结构复杂,受载形式较为复杂,承受缸盖、轴承盖螺栓预紧力、轴瓦过盈力等静力作用与气缸爆压力、曲柄连杆机构、及活塞组惯性力与活塞的敲击力等动载作用。随内燃机性能不断提高,曲轴的工作条件日益苛刻,强度要求更加严格。机体与曲轴的强度对内燃机工程有重要意义。本文对润滑在机体、曲轴动态强度分析方而进行了探索性研究,在强度分析中引入润滑因素的影响。在机体—曲轴系统动力学润滑耦合分析的基础上,获取机体与曲轴在一个完整工作循环中的动态应力分布,进行了机体与曲轴的动态疲劳强度分析,获取各部位的疲劳安全系数,考查了机体与曲轴的动态强度。通过获取整个工作循环中机体各主轴承对应各曲轴转角下的轴承总压,进行主轴承壁的动态强度分析。分析结果表明：从机体横剖面上看,该柴油机机体的动态疲劳安全系数较小处位于机体顶面上机体与缸套的安装孔边沿与机体底部缸间横隔板上主轴承座孔肋的上端面；从机体纵剖面上看,机体上安全系数较小的位置一般位于气缸间横隔板处,出现位置与静态疲劳强度分析结果安全系数较小位置基本一致。该柴油机第6主轴承壁安全系数最小的位置位于机座下方切口的过渡圆角处；安全系数次小的位置位于止推轴承安装孔下方的凹槽处。曲轴上较小的安全系数均出现在曲柄销与主轴颈的圆角上,疲劳系数较小的区域与静态疲劳强度分析结果基本一致。曲柄销与主轴颈的圆角为强度较薄弱部位,在建立曲轴网格模型时,需要对这些圆角处进行细化。更多还原

【Abstract】 Cylinder liner-piston rings friction pair and main bearings are key friction pairs on a diesel engine, the friction and lubrication properties of which have a great influence on the performance of the engine as a whole. This paper studies the friction and lubrication properties of the cylinder liner-piston rings friction pair and main bearings of diesel engines by combining theoretical analysis, numerical calculation and experimental techniques. The study mainly covers:First, according to the gas flow equation and law of conservation of mass, this paper establishes a leakage calculation model for the piston ring to analyze the leakage of an inline six-cylinder diesel engine; then based on the average Reynolds equation builds a computational model for analysis of the lubrication properties of a cylinder liner-piston ring friction pair, and in the model the effect of roughness on lubrication properties and the effects of variable density and variable viscosity lubricants are considered; this paper derives the difference scheme of discrete and gives the solution to solve the average Reynolds equation; this paper also summarizes the problems that occur in the solving process based on the established lubrication analysis model and gives the matters that need attention. This paper analyzes the lubrication of the given diesel engine. The results show that the film thickness ratio of the cylinder liner-piston ring friction pair at a certain period of crank angle near the top dead center of compression is smaller than4, in the mixed lubrication state. For the rest of the crank angle time, the film thickness ratio is greater than4, in the hydrodynamic lubrication state; the maximum film pressure occurs near the top dead center of compression, more than twice the maximum pressure in the combustion chamber; the maximum frictional force also occurs near the top dead center of compression, mainly caused by the solid contact between two surfaces of the friction pair due to thin film. At other time, the internal frictional force is small. The maximum friction power loss occurs near the top dead center of compression. The maximum adhesive wear depth of the cylinder liner lies at the TDC of the first ring. Second, according to the established cylinder liner-piston ring lubrication computational model, this paper examines the factors that affect the lubrication properties of cylinder liner-piston ring-lubricant viscosity, lubricant temperature, friction pair surface roughness, and piston ring structure parameters; this paper also considers the non-Newtonian of lubricant, derives the power law fluid lubrication model applicable to the calculation of cylinder liner-piston ring lubrication, gives the difference scheme of average Reynolds equation and the solution. The cylinder liner-piston ring lubrication calculation results show that by reducing the surface roughness of cylinder liner-piston ring friction pair, the minimum film thickness tends to increase, the probability of occurrence of the mixed lubrication zone reduces, and the lubrication conditions are improved; the maximum friction and maximum friction power consumption are reduced. The increased lubricant viscosity can improve the lubrication conditions at the top and bottom dead centers, but the friction at other crank angle time slightly increases; with the lubricant temperature decreasing, the probability of the friction pair in the form of mixed lubrication and boundary lubrication decreases, and the probability of dry friction decreases; the maximum friction decreases, but the full-film lubrication friction slightly increases; from the perspective of the entire cycle, except the time when the film thickness is relatively small, the friction power loss increases at other time as the lubricant temperature decreases; therefore, for the purpose of lubrication properties, the lubricant needs to maintain a certain temperature, neither too high nor too low. With the height of piston ring barrel surface reduced, the piston ring’s side profile curve becomes gentle, which is conducive to the formation of the squeeze effect. The minimum film thickness increases at the top and bottom dead centers, and decreases at other crank angle time as the hydrodynamic effect weakens; in the vicinity of the top and bottom dead centers, the friction slightly decreases, and at the rest of crank angle time, the friction slightly increases; therefore, reducing the height of piston ring barrel surface may improve the lubrication properties at the top and bottom dead centers, but diminishes hydrodynamic effect at other crank angle time, and thus the piston ring barrel surface height value should be reasonably selected to achieve the best results. The increase of piston ring height increases the contact area of the friction pair, reduces the carrying capacity of unit area, increases the film thickness, improves the lubrication properties; but at the same time, because of the increased contact area, the fluid resistance may increase and lead to higher friction and friction power loss. So, comprehensive considerations should be made to select the right ring height; when the change of viscosity with pressure is ignored, the calculation results of the minimum film thickness are smaller than the situation when the change in viscosity is not ignored, but the maximum friction and maximum friction power loss are much higher. When considering the non-Newtonian effect of the lubricant, with different power law coefficients, the distribution shape of the oil film pressure is uniform, the locations with the maximum oil film pressure are consistent, and the change trend is consistent. But the value of oil film pressure increases with the increase of the power law coefficient, that is, the greater the power law coefficient is, the higher the carrying capacity of the oil film is. For lubricant with a power law coefficient of0.9, the influence of changing the piston ring height and the friction pair’s surface roughness on the film thickness ratio and oil film pressure is basically the same as the situation when the lubricant is a Newtonian fluid.Experimental analysis has been an important aspect for the study on the friction wear characteristics of cylinder liner-piston ring. For a simulation experiment, the emphasis and difficulty is how to accurately simulate the actual situation. In traditional wear experiments, slices of the material of friction pair part are commonly used as a sample for study. Also, the simulation conditions are somewhat different from the real working environment of the diesel engine. This chapter, from the perspective of traditional analysis of tribology, designs more accurate simulation criteria for the simulation experiment, in which real and complete cylinder liner-piston rings are used for pair, temperature conditions, as well as acidic environment that occurs in reality are included. The simulation experiment of friction wear process is conducted on a dedicated cylinder-piston ring reciprocating friction wear experimenting machine. Great efforts are made to make the wear form and wear conditions closer to the actual working conditions to improve the experiment credibility. By measuring the wear data and real-time friction data in the friction wear process of cylinder liner and piston ring, the comparative analysis of friction characteristics of cylinder liner-piston rings made of four different materials is performed. The comparative analysis is performed through the friction and wear experiment. The analysis results show that combining the experiment results, the best match for the diesel engine cylinder liner-piston ring is austempering cylinder liner-sprayed molybdenum piston ring.In analyzing the lubrication characteristics of the main bearings, the elastic deformation cannot be ignored. The elastic deformation equation should be included in the lubrication equation set to seek solution. In this paper, the multi-body dynamics and main bearings EHL are coupled to study the lubrication properties of the diesel engine. The main parameters obtained for the calculation include the main bearing loads, main bearing axis center orbit, minimum film thickness, maximum film pressure, and friction power loss. The study results show that:in the main bearing loads, the load in the direction perpendicular to the cylinder (Z-direction) is greater than the load in the main and auxiliary thrust surface direction (Y-direction); the peak loads in the Z-direction of2nd,3rd,5th,6th main bearings are huge; the peak loads in the Z-direction of1st,4th,7th main bearings are small; the axis orbits of1st,7th main bearings are similar, mainly in circular motion, rarely in high-speed centripetal motion, and the occurrence probability of cavitation is low; the occurrence probability of axis centripetal motion of the2nd main bearing is not high, thus the lubrication properties are good; most of the axis orbit of the4th main bearing falls in the60℃A-240℃A zone, resulting in a small oil film thickness and high vulnerability to wear; the axis orbits of the3rd,5th and6th main bearings show a significant centripetal motion, prone to cause cavitation. The minimum film thickness of the1st main bearing is the smallest, which needs special attention; the position where the maximum film pressure occurs is basically consistent with the position where the Z-direction bearing peak load occurs. While the position of minimum film thickness is not necessarily consistent with the bearing peak load position; the friction power loss trend is basically the same as that of the maximum film pressure, and the position of peak power friction power loss is basically consistent with the bearing·peak load position. The main bearing lubrication properties when the single cylinder misfires are studied also.Engine block, as the skeleton of an engine, has a very complex structure. The loads it is subject to are also very complex, including static loads such as pre-tightening loads applied by cylinder head and bearing cap bolts, and interference force of bearing pad, and dynamic loads such as cylinder detonation pressure, inertia force of crank linkage mechanism and piston group, and striking force of piston. With the continuous improvement of the performance of internal combustion engine, the operating conditions of crankshaft are increasingly severe and the requirements on strength are more stringent. To consider the lubrication effect on the block&crankshft strength analysis, based on the coupled analysis of the engine block-crankshaft system’s dynamics and tribology, this paper obtains the dynamic stress distribution of engine block and crankshaft in a complete cycle; analyzes the dynamic fatigue strength of engine block and crankshaft, acquires the fatigue safety factors of various parts, and examines the dynamic strength of engine block and crankshaft. The dynamic strength analysis of the main bearing wall is also carried out. The analysis results show that when viewed from the transverse section of the engine block, the smaller dynamic fatigue safety factors for the diesel engine block are located at the edge of the mounting holes of the engine block and cylinder liner on the top surface of the engine block, and the upper end surface of the bearing saddle bore rib on the lateral bulkhead at the engine block bottom; when viewed from longitudinal section of the engine block, the smaller safety factors are generally located at the lateral bulkhead between cylinders, basically consistent with the occurrence positions of smaller safety factors according to analysis results of static fatigue strength; all the smaller safety factors on the crankshaft occur on the fillets of the crank pin and main journal. The zone of smaller fatigue safety factors is basically consistent with the analysis results of static fatigue strength. The minimum safety factor of the main bearing wall lies at the transition fillet of the incision transition fillet below the main bearing cover. The fillets of crank pin and main journal have a relatively weak strength. In establishing the grid model for crankshaft, these fillets need to be refined.更多还原