气门摇臂轴支座 加工工艺及铣Φ26面夹具设计课程设计
收藏
资源目录
压缩包内文档预览：(预览前20页/共26页)
编号：94593685
类型：共享资源
大小：2.34MB
格式：ZIP
上传时间：2020-09-14
上传人：好资料QQ****51605
认证信息
个人认证
孙**（实名认证）
江苏
IP属地：江苏
20
积分
- 关 键 词：
-
气门摇臂轴支座
加工工艺及铣Φ26面夹具设计课程设计
气门
摇臂
支座
加工
工艺
26
夹具
设计
课程设计
- 资源描述：
-
喜欢就充值下载吧。。。资源目录里展示的全都有，，下载后全都有,,请放心下载，原稿所见即得所得哦=====================喜欢就充值下载吧。。。资源目录里展示的全都有，，下载后全都有,,请放心下载，原稿所见即得所得哦=====================喜欢就充值下载吧。。。资源目录里展示的全都有，，下载后全都有,,请放心下载，原稿所见即得所得哦=====================喜欢就充值下载吧。。。资源目录里展示的全都有，，下载后全都有,,请放心下载，原稿所见即得所得哦=====================
- 内容简介：
-
编号无锡太湖学院毕业设计（论文）相关资料题目： 摇臂零件工艺及工装设计 信机 系 机械工程及自动化专业学 号： 0923140学生姓名： 司舒晖 指导教师： 许文（职称：副教授） 2013年5月25日目 录一、毕业设计（论文）开题报告二、毕业设计（论文）外文资料翻译及原文三、学生“毕业论文（论文）计划、进度、检查及落实表”四、实习鉴定表无锡太湖学院毕业设计（论文）开题报告题目： 摇臂零件工艺及工装设计 信机 系 机械工程及自动化 专业学 号： 0923140 学生姓名： 司舒晖 指导教师： 许文 （职称：副教授 ） 2012年11月14日 课题来源自拟题目科学依据（1）课题科学意义随着现代社会进程的加快，柴油机发挥的社会作用不可估量，特别是在社会工业化之后，柴油机作为动力内燃机的一种，在社会的各个领域无处不在，为社会创造着巨大的效益。在这领域中，柴油机所发挥的作用也是不尽相同，所以根据作用的需要，柴油机也被设计出了很多种型号，各种型号功率不同，发挥的作用大小也就不一样，创造出的价值也不一样。但是柴油机的污染排放也是一个不小的社会问题，随着社会的发展，人类对生活质量要求的提高，而高污染排放的柴油机必定不能满足人类的这一生活需求，但是柴油机已经是社会发展不可缺少的一个重要零部分，彻底取代柴油机在目前的技术条件下似乎还不太可能。（2）研究状况及其发展前景：随着社会的需要，柴油机生产数量将不断的增长，而气门摇臂轴支座是柴油机上不可或缺的零件，也就是意味着气门摇臂轴支座的生产数量将是与日俱增，为了创造出更大的效益，设计出轻便，经久耐用，便于生产的气门摇臂轴支座这一零件是很有必要的。柴油机具有热效率高的显著优点，其应用范围越来越广。随着强化程度的提高，柴油机单位功率的重量也显著降低。为了节能，各国都在注重改善燃烧过程，研究燃用低质燃油和非石油制品燃料。此外，降低摩擦损失、广泛采用废气涡轮增压并提高增压度、进一步轻量化、高速化、低油耗、低噪声和低污染，都是柴油机的重要发展方向。研究内容了解气门摇臂零件的工作原理，国内外的研究发展现状； 完成气门摇臂零件的总体方案设计； 完成有关零部件的选型计算、结构强度校核； 熟练掌握计算机CAD绘图软件，并绘制装配图和零件图纸，折合A0不少于2.5张； 完成说明书的撰写，并且翻译外文资料1篇。拟采取的研究方法、技术路线、实验方案及可行性分析1）技术路线首先根据气门摇臂零件的特殊性对其造型等方面的设计需求进行分析，从整体上把握其设计原则；然后对不同的功能区域进行单独的研究分析，总结出符合工程学要求的设计理论；最后将整体的设计分析和每一部分的设计相结合，寻找有效的结合点并进行统一协调，最终设计出高质量、高档次的产品。（2）研究方法 测试出气门摇臂各零件的尺寸、刚度，获得大量的实验数据。 对实验数据进行分析处理，为建立气门摇臂力学模型与分析作了必要的准备。（3）实验方案 确定具体设计方案：零件的工艺分析及生产类型的确定，零件的工艺分析研究计划及预期成果（1）研究计划：2012年10月28日-2012年11月16日：学习并翻译一篇与毕业设计相关的英文材料2012年11月20日-2013年1月20日：按照任务书要求查阅论文相关参考资料，填写毕业设计开题报告书。2013年1月25日-2013年2月10日：填写毕业实习报告。2013年2月20日-2013年3月10日：按照要求修改毕业设计开题报告。2013年3月19日-2013年3月30日：气门摇臂轴支座铣18孔端面的夹具结构设计。2013年4月1日-2013年4月25日：CAD绘图。2013年4月26日-2013年5月21日：毕业论文撰写和修改工作。（2）预期成果：我国市场前景广阔，产品质量性能逐渐满足要求，因此产品的发展必须由单纯的追求技术上的完善，转向产品外观质量的提高，放到与技术改进放到同等重要的位置，通过本课题的研究，产品必定以合理的色彩以及人性化的结构方式提高自己的附加值，吸引到更多地客户，加大自己产品的市场占有率，提高在行业中的竞争力。特色或创新之处1通用性好，气门摇臂轴支座铣18孔端面在设计过程中，考略到通用性，因此留有余地，因此除搬运外，还可以焊接喷漆等。2工作效率，提高了劳动生产效率，同时也降低了成本。已具备的条件和尚需解决的问题（1）.夹具的构造应与其用途和生产规模相适应，正确处理好质量、效率、方便性与经济性四者的关系。 （2）.保证使用方便，要便于装卸、便于夹紧、便于测量、便于观察、便于排屑排液、便于安装运输，保证安全第一。 （3）.注意结构工艺，对加工、装配、维修通盘考虑，降低成本。指导教师意见 指导教师签名：年 月 日教研室（学科组、研究所）意见 教研室主任签名： 年 月 日系意见 主管领导签名： 年 月 日英文原文Experimental investigation of laser surface textured parallel thrust bearingsPerformance enhancements by laser surface texturing (LST) of parallel-thrust bearings is experimentally investigated. Testresults are compared with a theoretical model and good correlation is found over the relevant operating conditions. A compari-son of the performance of unidirectional and bi-directional partial-LST bearings with that of a baseline, untextured bearing ispresented showing the benets of LST in terms of increased clearance and reduced friction.KEY WORDS: uid lm bearings, slider bearings, surface texturing1. IntroductionThe classical theory of hydrodynamic lubrication yields linear (Couette) velocity distribution with zero pressure gradients between smooth parallel surfaces under steady-state sliding. This results in an unstable hydrodynamic lm that would collapse under any external force acting normal to the surfaces. However, experience shows that stable lubricating lms can develop between parallel sliding surfaces, generallybecause of some mechanism that relaxes one or more of the assumptions of the classical theory.A stable uid lm with sucient load-carrying capacity in parallel sliding surfaces can be obtained, for example, with macro or micro surface structure of dierent types. These include waviness 1 and protruding microasperities 24. A good literature review on the subject can be found in Ref. 5. More recently, laser surface texturing (LST) 68, as well as inlet roughening by longitudinal or transverse grooves 9 were suggested to provide load capacity in parallel sliding. The inlet roughness concept of Tonder 9 is based on eective clearance reduction in the slidingdirection and in this respect it is identical to the par- tial-LST concept described in ref. 10 for generating hydrostatic eect in high-pressure mechanical seals.Very recently Wang et al. 11 demonstrated experimentally a doubling of the load-carrying capacity for the surface- texture design by reactive ion etching of SiC parallel-thrust bearings sliding in water. These simple parallel thrust bearings are usually found in seal-less pumps where the pumped uid is used as the lubricant for the bearings. Due to the parallel sliding their performance is poorer than more sophisticated tapered or stepped bearings. Brizmer et al. 12 demon-strated the potential of laser surface texturing in the form of regular micro-dimples for providing load-carrying capacity with parallel-thrust bearings. A model of a textured parallel slider was developed and the eect of surface texturing on load-carrying capacitywas analyzed. The optimum parameters of the dimples were found in order to obtain maximum load-carrying capacity. A micro-dimple collective eect was identi-ed that is capable of generating substantial load-carrying capacity, approaching that of optimumconventional thrust bearings. The purpose of the present paper is to investigate experimentally the validity of the model described in Ref. 12 by testing practical thrust bearings and comparing the performance of LST bearings with that of the theoretical predictions and with the performance of standard non-texturedbearings2. BackgroundA cross section of the basic model that was analyzed in Ref. 12 is shown in figure 1. A slider having a width B is partially textured over a portion Bp =B of its width. The textured surface consists of multiple dimples with a diameter，depthand area density Sp. As a result of the hydrodynamic pressure generated by the dimples the sliding surfaces will be separated by a clearancedepending on the sliding velocity U, the uid viscosity l and the external loadIt was found in Ref. 12 that an optimum ratio exists for the parameter that provides maximum dimensionless load-carrying capacity where L isthe bearing length, and this optimum value is hp=1.25. It was further found in Ref. 12 that an optimum value exists for the textured portion a depending onthe bearing aspect ratio L/B. This behavior is shown in gure 2 for a bearing with L/B = 0.75 at various values of the area density Sp. As can be seen in the range of Sp values from 0.18 to 0.72 the optimum a value varies from 0.7 to 0.55, respectively. It can also be seen from gure 2 that for a 0.85 no optimum value exists for Sp and the maximum load W increases with increasing Sp. Hence, the largest area density that can be practically obtained with the laser texturing is desired. It is also interesting to note from gure 2 the advantage of partial-LST (a 1) over the full LST (a = 1) for bearing applications. At Sp= 0.5, for example, the load W at a = 0.6 is about three times higher than its value at a = 1. A full account of this behavior is given in Ref. 12.3. ExperimentalThe tested bearings consist of sintered SiC disks 10 mm thick, having 85 mm outer diameter and 40 mm inner diameter. Each bearing (see gure 3) comprises a at rotor (a) and a six-pad stator (b). The bearings were provided with an original surface nishby lapping to a roughness average Ra= 0.03 lm. Each pad has an aspect ratio of 0.75 when its width is measured along the mean diameter of the stator. The photographs of two partial-LST stators are shown in gure 4 where the textured areas appear as brighter matt surfaces. The rst stator indicated (a) is a unidirectional bearing with the partial-LST adjacent to the leading edge of each pad, similar to the model shown in gure 1. The second stator (b) is a bi-directional version of a partial-LST bearing having two equal textured portions, a/2, on each of the pad ends. The laser texturing parameters were the following; dimple depth, dimple diameter and dimple area density Sp= 0.60.03. These dimple dimensions were obtained with 4 pulses of 30 ns duration and 4 mJ each using a 5 kHz pulsating Nd:YAG laser. The textured portion of the unidirectional bearing was a= 0.73 and that of the bi-directional bearing was a= 0.63. As can be seen from gure 2 both these a values should produce load-carrying capacity vary close to the maximum theoretical value.The test rig is shown schematically in gure 5. Anelectrical motor turns a spindle to which an upper holder of the rotor is attached. A second lower holder of the stator is xed to a housing, which rests on a journal bearing and an axial loading mechanism that can freely move in the axial direction. An arm that presses against a load cell and thereby permits friction torque measurements prevents the free rotation of this housing. Axial loading is provided by means of dead weights on a lever and is measured with a second load cell. A proximity probe that is attached to the lower holder of the stator allows on-line measurements of the clearance change between rotor and stator as the hydrodynamic eects cause axial movement of the housing to which the stator holder is xed. Tap water is supplied by gravity from a large tank to the center of the bearing and the leakage from the bearing is collected and re-circulated. A thermocouple adjacent tothe outer diameter of the bearing allows monitoring of the water temperature as the water exit the bearing. A PC is used to collect and process data on-line. Hence,the instantaneous clearance, friction coecient, bearing speed and exit water temperature can be monitored constantly. The test protocol includes identifying a reference “zero” point for the clearance measurements by rst loading and then unloading a stationary bearing over the full load range. Then the lowest axial load is applied, the water supply valve is opened and the motor turned on. Axial loading is increased by steps of 40 N and each load step is maintained for 5 min following the stabilization of the friction coecient ata steady-state value. The bearing speed and water temperature are monitored throughout the test for any irregularities. The test ends when a maximum axial load of 460 N is reached or if the friction coecient exceeds a value of 0.35. At the end of the last load step the motor and water supply are turned o and the reference for the clearance measurements is rechecked. Tests are performed at two speeds of 1500and 3000 rpm corresponding to average sliding velocities of 4.9 and 9.8 m/s, respectively and each test is repeated at least three times.4. Results and discussionAs a rst step the validity of the theoretical model in Ref. 12 was examined by comparing the theoretical and experimental results of bearing clearance versus bearing load for a unidirectional partial-LST bearing. The results are shown in gure 6 for the two speeds of 1500 and 3000 rpm where the solid and dashed lines correspond to the model and experiment, respectively. As can be seen, the agreement between the model and the experiment is good, with dierences of less than 10%, as long as the load is above 150 N. At lower loads the measured experimental clearances are much larger than the model predictions, particularly at the higher speed of 3000 rpm where at 120 N the measured clearance is 20 lm, which is about 60% higher than the predicted value. It turns out that the combination of such large clearances and relatively low viscosity of the water may result in turbulent uid lm. Hence, the assumption of laminar ow on which the solution of the Reynolds equation in Ref. 12 is based may be violated making the model invalid especially at the higher speed and lowest load. In order to be consistent with the model of Ref. 12 it was decided to limit further comparisons to loads above 150 N.It should be noted here that the rst attempts to test the baseline untextured bearing with the original surface nish of Ra= 0.03 lm on both the stator and rotor failed due to extremely high friction even at the lower loads. On the other hand the partial-LST bearing ran smoothly throughout the load range. It was found that the post-LST lapping to completely remove about 2 lm height bulges, which are formed during texturing around the rims of the dimples, resulted in a slightly rougher surface with Ra= 0.04 lm. Hence, the baseline untextured stator was also lapped to the same rough-ness of the partial-LST stator and all subsequent tests were performed with the same Ra value of 0.04 lm for all the tested stators. The rotor surface roughnessremained, the original one namely, 0.03 lm. Figure 7 presents the experimental results for the clearance as a function of the load for a partial-LST unidirectional bearing (see stator in gure 4(a) and a baseline untextured bearing. The comparison is made at the two speeds of 1500 and 3000 rpm. The area density of the dimples in the partial-LST bearing is Sp= 0.6 and the textured portion is a 0:734. The load range extends from 160 to 460 N. The upper load was determined by the test-rig limitation that did not permit higher loading. It is clear from gure 7 that the partial-LST bearing operates at substantially larger clearances than the untextured bearing. At the maximum load of 460 N and speed of 1500 rpm the partial-LST bearing has a clearance of 6 lm while the untextured bearing clearance is only 1.7 lm. At 3000 rpm the clearances are 6.6and 2.2 lm for the LST and untextured bearings, respectively. As can be seen from gure 7 this ratio of about 3 in favor of the partial-LST bearing is maintained over the entire load range.Figure 8 presents the results for the bi-directionalbearing (see stator in gure 4(b). In this case the LST parameters are Sp 0:614 and a 0:633. The clearances of the bi-directional partial-LST bearing are lower compared to these of the unidirectional bearing at the same load. At 460 N load the clearance for the 1500 rpm is 4.1 lm and for the 3000 rpm it is 6 lm. These values represent a reduction of clearance between33 and 10% compared to the unidirectional case. However, as can be seen from gure 8 the performance of the partial-LST bi-directional bearing is still substantially better than that of the untextured bearing. The friction coecient of partial-LST unidirectional and bi-directional bearings was compared with that of the untextured bearing in gures 9 and 10 for the two speeds of 1500 and 3000 rpm, respectively. As can be seen the friction coecient of the two partial-LST bearings is very similar with slightly lower values in the case of the more ecient unidirectional bearing. The friction coecient of the untextured bearing ismuch larger compared to that of the LST bearings. At 1500 rpm (gure 9) and the highest load of 460 N the friction coecient of the untextured bearing is about 0.025 compared to about 0.01 for the LST bearings.At the lowest load of 160 N the values are about 0.06 for the untextured bearing and around 0.02 for the LST bearings. Hence, the friction values of the untextured bearing are between 2.5 and 3 times higher than the corresponding values for the partial-LST bearings over the entire load range. Similar results were obtained at the velocity of 3000 rpm (gure 10) but the level of the friction coecients is somewhat higherdue to the higher speed. The much higher friction of the untextured bearing is due to the much smaller clearances of this bearing (see gures 7 and 8) that result in higher viscous shear.Bearings fail for a number of reasons，but the most common are misapplication，contamination，improper lubricant，shipping or handling damage，and misalignment. The problem is often not difficult to diagnose because a failed bearing usually leaves telltale signs about what went wrongHowever，while a postmortem yields good information，it is better to avoid the process altogether by specifying the bearing correctly in The first placeTo do this，it is useful to review the manufacturers sizing guidelines and operating characteristics for the selected bearing.Equally critical is a study of requirements for noise, torque, and runout, as well as possible exposure to contaminants, hostile liquids, and temperature extremes. This can provide further clues as to whether a bearing is right for a job.1 Why bearings failAbout 40% of ball bearing failures are caused by contamination from dust, dirt, shavings, and corrosion. Contamination also causes torque and noise problems, and is often the result of improper handling or the application environmentFortunately, a bearing failure caused by environment or handling contamination is preventable，and a simple visual examination can easily identify the causeConducting a postmortem il1ustrates what to look for on a failed or failing bearingThen，understanding the mechanism behind the failure, such as brinelling or fatigue, helps eliminate the source of the problem.Brinelling is one type of bearing failure easily avoided by proper handing and assembly. It is characterized by indentations in the bearing raceway caused by shock loadingsuch as when a bearing is dropped-or incorrect assembly. Brinelling usually occurs when loads exceed the material yield point(350,000 psi in SAE 52100 chrome steel)It may also be caused by improper assembly, Which places a load across the racesRaceway dents also produce noise，vibration，and increased torque.A similar defect is a pattern of elliptical dents caused by balls vibrating between raceways while th
- 温馨提示:
1: 本站所有资源如无特殊说明，都需要本地电脑安装OFFICE2007和PDF阅读器。图纸软件为CAD,CAXA,PROE,UG,SolidWorks等.压缩文件请下载最新的WinRAR软件解压。
2: 本站的文档不包含任何第三方提供的附件图纸等，如果需要附件，请联系上传者。文件的所有权益归上传用户所有。
3.本站RAR压缩包中若带图纸，网页内容里面会有图纸预览，若没有图纸预览就没有图纸。
4. 未经权益所有人同意不得将文件中的内容挪作商业或盈利用途。
5. 人人文库网仅提供信息存储空间，仅对用户上传内容的表现方式做保护处理，对用户上传分享的文档内容本身不做任何修改或编辑，并不能对任何下载内容负责。
6. 下载文件中如有侵权或不适当内容，请与我们联系，我们立即纠正。
7. 本站不保证下载资源的准确性、安全性和完整性, 同时也不承担用户因使用这些下载资源对自己和他人造成任何形式的伤害或损失。
人人文库网所有资源均是用户自行上传分享，仅供网友学习交流，未经上传用户书面授权，请勿作他用。
川公网安备: 51019002004831号