氣門搖臂軸支座零件工藝及工裝設計[銑Φ26面]【含CAD高清圖紙、文檔所見所得】

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高 柴 油 機 單 位 功 率 的 重 量 也 顯 著 降 低 為 了 節(jié) 能 各 國 都 在 注 重 改 善 燃 燒 過 程 研 究 燃 用 低 質 燃 油 和 非 石 油 制 品 燃 料 此 外 降 低 摩 擦 損 失 廣 泛 采 用 廢 氣 渦 輪 增 壓 并 提 高 增 壓 度 進 一 步 輕 量 化 高 速 化 低 油 耗 低 噪 聲 和 低 污 染 都 是 柴 油 機 的 重 要 發(fā) 展 方 向 研究內容 了解氣門搖臂零件的工作原理 國內外的研究發(fā)展現狀 完成氣門搖臂零件的總體方案設計 完成有關零部件的選型計算 結構強度校核 熟練掌握計算機 CAD 繪圖軟件 并繪制裝配圖和零件圖紙 折合 A0 不少于 2 5 張 完成說明書的撰寫 并且翻譯外文資料 1 篇 擬采取的研究方法 技術路線 實驗方案及可行性分析 1 技術路線 首先根據氣門搖臂零件的特殊性對其造型等方面的設計需求進行分析 從整體 上把握其設計原則 然后對不同的功能區(qū)域進行單獨的研究分析 總結出符合工 程學要求的設計理論 最后將整體的設計分析和每一部分的設計相結合 尋找有 效的結合點并進行統一協調 最終設計出高質量 高檔次的產品 2 研究方法 測試出氣門搖臂各零件的尺寸 剛度 獲得大量的實驗數據 對實驗數據進行分析處理 為建立氣門搖臂力學模型與分析作了必要的準備 3 實驗方案 確定具體設計方案 零件的工藝分析及生產類型的確定 零件的工藝分析 研究計劃及預期成果 1 研究計劃 2012 年 10 月 28 日 2012 年 11 月 16 日 學習并翻譯一篇與畢業(yè)設計相關的英文材料 2012 年 11 月 20 日 2013 年 1 月 20 日 按照任務書要求查閱論文相關參考資料 填寫 畢業(yè)設計開題報告書 2013 年 1 月 25 日 2013 年 2 月 10 日 填寫畢業(yè)實習報告 2013 年 2 月 20 日 2013 年 3 月 10 日 按照要求修改畢業(yè)設計開題報告 2013 年 3 月 19 日 2013 年 3 月 30 日 氣門搖臂軸支座銑 18 孔端面的夾具結構設計 2013 年 4 月 1 日 2013 年 4 月 25 日 CAD 繪圖 2013 年 4 月 26 日 2013 年 5 月 21 日 畢業(yè)論文撰寫和修改工作 2 預期成果 我國市場前景廣闊 產品質量性能逐漸滿足要求 因此產品的發(fā)展必須由單純的 追求技術上的完善 轉向產品外觀質量的提高 放到與技術改進放到同等重要的位置 通過本課題的研究 產品必定以合理的色彩以及人性化的結構方式提高自己的附加值 吸引到更多地客戶 加大自己產品的市場占有率 提高在行業(yè)中的競爭力 特色或創(chuàng)新之處 1 通用性好 氣門搖臂軸支座銑 18 孔端面在設計過程中 考略到通用性 因此留 有余地 因此除搬運外 還可以焊接噴漆等 2 工作效率 提高了勞動生產效率 同時也降低了成本 已具備的條件和尚需解決的問題 1 夾具的構造應與其用途和生產規(guī)模相適應 正確處理好質量 效率 方便性與 經濟性四者的關系 2 保證使用方便 要便于裝卸 便于夾緊 便于測量 便于觀察 便于排屑排 液 便于安裝運輸 保證安全第一 3 注意結構工藝 對加工 裝配 維修通盤考慮 降低成本 指導教師意見 指導教師簽名 年 月 日 教研室 學科組 研究所 意見 教研室主任簽名 年 月 日 系意見 主管領導簽名 年 月 日 英文原文 Experimental investigation of laser surface textured parallel thrust bearings Performance enhancements by laser surface texturing LST of parallel thrust bearings is experimentally investigated Test results 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 is presented showing the bene ts of LST in terms of increased clearance and reduced friction KEY WORDS uid lm bearings slider bearings surface texturing 1 Introduction The 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 generally because of some mechanism that relaxes one or more of the assumptions of the classical theory A stable uid lm with su cient load carrying capacity in parallel sliding surfaces can be obtained for example with macro or micro surface structure of di erent types These include waviness 1 and protruding microasperities 2 4 A good literature review on the subject can be found in Ref 5 More recently laser surface texturing LST 6 8 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 e ective clearance reduction in the sliding direction and in this respect it is identical to the par tial LST concept described in ref 10 for generating hydrostatic e ect 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 e ect of surface texturing on load carrying capacity was analyzed The optimum parameters of the dimples were found in order to obtain maximum load carrying capacity A micro dimple collective e ect 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 textured bearings 2 Background A 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 depth and area density Sp As a result of the hydrodynamic pressure generated by the dimples the sliding surfaces will be separated by a clearance depending on the sliding velocity U the uid viscosity l and the external load It was found in Ref 12 that an optimum ratio exists for the parameter that provides maximum dimensionless load carrying capacity where L is the 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 Experimental The 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 nish by 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 An electrical 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 e ects 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 to the 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 coe cient 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 coe cient at a 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 coe cient 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 1500 and 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 discussion As 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 di erences 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 roughness remained 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 6 and 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 between 33 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 coe cient 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 coe cient of the two partial LST bearings is very similar with slightly lower values in the case of the more e cient unidirectional bearing The friction coe cient of the untextured bearing is much larger compared to that of the LST bearings At 1500 rpm gure 9 and the highest load of 460 N the friction coe cient 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 coe cients is somewhat higher due 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 wrong However while a postmortem yields good information it is better to avoid the process altogether by specifying the bearing correctly in The first place To 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 fail About 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 environment Fortunately a bearing failure caused by environment or handling contamination is preventable and a simple visual examination can easily identify the cause Conducting a postmortem il1ustrates what to look for on a failed or failing bearing Then 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 loading such 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 races Raceway dents also produce noise vibration and increased torque A similar defect is a pattern of elliptical dents caused by balls vibrating between raceways while the bearing is not turning This problem is called false brinelling It occurs on equipment in transit or that vibrates when not in operation In addition debris created by false brinelling acts like an abrasive further contaminating the bearing Unlike brinelling false binelling is often indicated by a reddish color from fretting corrosion in the lubricant False brinelling is prevented by eliminating vibration sources and keeping the bearing well lubricated Isolation pads on the equipment or a separate foundation may be required to reduce environmental vibration Also a light preload on the bearing helps keep the balls and raceway in tight contact Preloading also helps prevent false brinelling during transit Seizures can be caused by a lack of internal clearance improper lubrication or excessive loading Before seizing excessive friction and heat softens the bearing steel Overheated bearings often change color usually to blue black or straw colored Friction also causes stress in the retainer which can break and hasten bearing failure Premature material fatigue is caused by a high load or excessive preload When these conditions are unavoidable bearing life should be carefully calculated so that a maintenance scheme can be worked out Another solution for fighting premature fatigue is changing material When standard bearing materials such as 440C or SAE 52100 do not guarantee sufficient life specialty materials can be recommended In addition when the problem is traced back to excessive loading a higher capacity bearing or different configuration may be used Creep is less common than premature fatigue In bearings it is caused by excessive clearance between bore and shaft that allows the bore to rotate on the shaft Creep can be expensive because it causes damage to other components in addition to the bearing 0ther more likely creep indicators are scratches scuff marks or discoloration to shaft and bore To prevent creep damage the bearing housing and shaft fittings should be visually checked Misalignment is related to creep in that it is mounting related If races are misaligned or cocked The balls track in a noncircumferencial path The problem is incorrect mounting or tolerancing or insufficient squareness of the bearing mounting site Misalignment of more than 1 4 can cause an early failure Contaminated lubricant is often more difficult to detect than misalignment or creep Contamination shows as premature wear Solid contaminants become an abrasive in the lubricant In addition insufficient lubrication between ball and retainer wears and weakens the retainer In this situation lubrication is critical if the retainer is a fully machined type Ribbon or crown retainers in contrast allow lubricants to more easily reach all surfaces Rust is a form of moisture contamination and often indicates the wrong mat