資源目錄里展示的全都有，所見即所得。下載后全都有,請放心下載。原稿可自行編輯修改=【QQ：401339828 或11970985 有疑問可加】 本科畢業(yè)設計(論文)選題審批表 題目名稱 2X-70旋片式真空泵設計 指導教師 職稱 講師 指導教師 職稱 題目來源 （1）教師擬訂；（2）學生建議；（3）企業(yè)和社會征集；（4）教師科研 課程類別 1.設計 2、論文 選題依據(jù) 旋片真空泵是真空行業(yè)量大面廣的產品，廣泛應用于冶金、醫(yī)藥、化工、電子等行業(yè)。本論文以2X-70為例，從泵的原理、整體結構、進油機構及防返油機構以及防噴油結構、冷卻方式結構等諸多方面進行了設計計算。采取了壓差供油的方式，并設置電磁閥以控制返油；在排氣閥處減少積油，并設置油氣分離裝置防止噴油；水冷卻的結構設計等措施。提供了清潔的真空作業(yè)環(huán)境，改善了旋片泵的使用性能，降低功率消耗，提高泵的使用壽命。 通過該課題的設計，可以讓學生將所學的專業(yè)知識融會貫通，建立起工程理念，掌握機械設計的和制造的一般過程。為從事機械行業(yè)的相關工作打下良好的基礎。 已有研究基礎 教研室 審核意見 教研室主任簽字： 2013年 9 月 日 學院 審批意見 負責人簽字： 2013年 9 月 日 畢業(yè)設計（論文）任務書 所在學院 專業(yè) 機械設計制造及其自動化 班級 學生姓名 學號 指導教師 題 目 2X-70旋片式真空泵設計 一、畢業(yè)設計（論文）工作內容與基本要求：（目標、任務、途徑、方法，應掌握的原始資料（數(shù)據(jù)）、參考資料（文獻）以及設計技術要求、注意事項等）（紙張不夠可加頁） 主要任務與目標： 2X-70兩級旋片泵技術指標： 1） 極限真空（無氣鎮(zhèn)）：達到6.7×101。 2） 名義抽速：70L/S。 3） 功率：小于5.5KW。 4） 進氣口內徑：10mm。 5） 溫升：80℃~85℃。 6） 噴油：泵工作穩(wěn)定以后，一分鐘內沒有噴油現(xiàn)象。 7） 噪聲：聲功率級﹤70db(A)。 8） 壽命：連續(xù)運轉500小時性能不變。 已知要求的立式注塑機的合模力為100噸。要求外形尺寸不宜過大，以經濟可靠為前提設計該注塑機合模部分機械結構。 1、 翻譯2篇與本課題相關的近幾年的英文文獻，文獻翻譯每篇要求在2000字以上； 2、 查閱和整理文獻并提交一篇反映課題內容的文獻綜述，文獻綜述在3000字以上； 3、 根據(jù)以上的技術指標設計旋片真空泵，要求獨立完成裝配圖，在此基礎上完成部分零件圖的設計，提交一份開題報告； 4、 按照開題報告的進度計劃，獨立進行合模結構設計所需的數(shù)據(jù)計算，結合相關 課程中涉及的經驗公式與經驗數(shù)據(jù)，撰寫論文，論文正文不少于10000字。 研究途徑與方法： 1、 結合所學專業(yè)課程，通過查閱相關資料，溫習相關CAD軟件，完成畢業(yè)設計； 2、 查閱注塑機相關信息，完成旋片真空泵的結構設計，結合專業(yè)課程制定畢業(yè)設計計劃，搭建論文正文主體框架，繪制二維裝配圖、拆畫主要零件圖、計算總體方案設計過程中所涉及的重要數(shù)據(jù)，校核主軸的剛度與強度，完善設計骨架，匯整豐富說明書內容，最后對格式進行標準化處理，檢索并翻譯外文資料，按論文指導手冊的要求完成畢業(yè)設計全部內容。 推薦資料、參考文獻： [1] 楊乃康編，真空獲得設備，冶金工業(yè)出版社，2005 [2] 成大先主編，《機械設計手冊》，化學工業(yè)出版社，2004 [3] 機械工程手冊編委會編，機械工程手冊（第二版），機械工業(yè)出版社，1995 [4] 鄭經緯 吳天星. 機械原理（第七版）. 北京：高等教育出版社，1996 [5] 曹龍華 蔣希成. 平面連桿機構綜合. 北京：高等教育出版社，1990 [6] 王三民. 機械原理與設計課程設計. 北京：機械工業(yè)出版社，2004 [7] 陸鳳儀. 機械原理課程設計. 北京：機械工業(yè)出版社，2001 [8]丁東升，計算機輔助注塑機設計關鍵技術研究[J]，東南大學碩士學位論文 2006（7） 設計技術要求： 1、 注明該泵的技術指標； 2、 注明該泵的外形尺寸； 3、 注明零件圖中各零件材料、表面處理要求及其他特殊要求等。 注意事項： 1、 零件圖需要有圖框、零件尺寸標注、技術要求需規(guī)范并符合制圖標準； 2、 要求2D圖總量折合為2張A0圖以上的量； 3、最終稿2D圖需轉成PDF形式保薦并提交電子文檔； 4、英文翻譯需注明原文出處，并附上PDF格式原文。 二、畢業(yè)論文進度計劃 序號 各階段工作內容 起訖日期 備注 1 畢業(yè)選題、下發(fā)任務 2013.09-2013.10 2 提交開題報告、外文翻譯、文獻綜述； 2013.10-2013.11 3 初定總體設計方案、初畫裝配圖 2013.11-2013.12 4 確定機械機構，繪制裝配圖 2013.12-2014.01 5 拆畫零件圖、撰寫技術說明書 2014.01-2014.02 6 完成圖紙的修改，完成畢業(yè)設計論文撰寫 2014.03-2014.04 7 打印、膠裝、答辯資格審核 2014.04-2014.05 8 準備答辯 2014.04-2014.06 三、專業(yè)（教研室）審批意見： 審批人（簽字）： 工作任務與工作量要求：原則上查閱文獻資料不少于12篇，其中外文資料不少于2篇；文獻綜述不少于3000字；文獻翻譯不少于2000字；畢業(yè)設計說明書或論文1篇不少于10000字。 提交相關圖紙、實驗報告、調研報告、譯文等其它形式的成果。畢業(yè)設計（論文）撰寫規(guī)范及有關要求，請查閱《畢業(yè)設計（論文）撰寫規(guī)范》。 備注：學生一人一題，指導教師對每一名學生下達一份《畢業(yè)設計（論文）任務書》。 寧波大紅鷹學院 畢業(yè)設計（論文）外文翻譯 所在學院： 宋體四號加粗 班 級： 姓 名： 學 號： 指導教師： 合作導師： 2013 年 11 月 15 日 原文： 題目 Research of an unattended intelligentized control system of air compressor for supplying constant-pressure air Lingen Chen , Jun Luo , Fengrui Sun , Chih Wu Postgraduate School, Naval University of Engineering, Wuhan, 430033, PR China Mechanical Engineering Department, US Naval Academy, Annapolis MN21402, USA Available online 28 November 2007 Abstract A model for the optimal design of a multi-stage compressor, assuming a fixed configuration of the flow-path, is presented.The absolute inlet and exit angles of the rotor, the absolute exit angle of the stator, and the relative gas densities at the inlet and exit stations of the stator, of every stage, are taken as the design variables. Analytical relations of the compressor elemental stage and the multi-stage compressor are obtained. Numerical examples are provided to illustrate the effects of various parameters on the optimal performance of the multi-stage compressor. 2007 Elsevier Ltd. All rights reserved. Keywords: Multi-stage axial-flow compressor; Efficiency; Analytical relation; Optimization 1. Introduction The design of the axial-flow compressor is partially an art. The lack of accurate prediction influences the design process. Until today, there are no methods currently available that permit the prediction of the values of these quantities to a sufficient accuracy for a new design. Some progresses has been achieved via the application of numerical optimization techniques to single- and multi-stage axial-flow compressor design [1–22].Especially with the development of computational fluid-dynamics (CFD), many more accurate methods of calculating have been presented in many references in which the techniques of CFD have been applied to two- and three-dimensional optimal designs of axial-flow compressors [17–20]. However, it is still of worthwhile significance to calculate, using one-dimensional flow-theory, the optimal design of compressors. Boiko [23] presented a detailed mathematical model for the optimal design of single- and multi-stage axial-flow turbines by assuming (i) a fixed distribution of axial velocities or (ii) a fixed flow-path shape, and obtained the corresponding optimized results. Using a similar idea, Chen et al. [22] presented a mathematical model for the optimal design of a single-stage axial-flow compressor by assuming a fixed distribution of axial velocities.In this paper, a model for the optimal design of a multi-stage axial-flow compressor, by assuming a fixed flow path shape, is presented. The absolute inlet and exit angles of the rotor, the absolute exit angle of the stator, and the relative gas densities at the inlet and exit stations of the stator, of each stage, are taken as the design variables. Analytical relations of the compressor stage are obtained. Numerical examples are provided to illustrate the effects of various parameters on the optimal performance of the multi-stage compressor 2. Fundamental equations for elemental-stage compressor Consider a n-stage axial-flow compressor – see Fig. 1. Fig. 2 shows the specific enthalpy–specific entropy diagram of this compressor. For a n-stage axial-flow compressor, there are (2n + 1) section stations. The stage velocity triangle of an intermediate stage (i.e. jth stage) is shown in Fig. 3. The corresponding specific enthalpy–specific entropy diagram is shown in Fig. 4. The performance calculation of multi-stage compressor is performed using one-dimensional flow theory. The analysis begins with the energy and continuity equations, and the axial-flow velocities of the working fluid and wheel velocities at the different stations in the compressor are not considered as constant, that is, , (), where i denotes the ith station and j denotes the jth stage. The major assumptions made in the method are as follows ? The working fluid flows stably relative to the vanes, stators and rotors, which rotate at a fixed speed. ? The working fluid is compressible, non-viscous and adiabatic. ? The mass-flow rate of the working fluid is constant. ? The compression process is homogeneous in the working fluid. ? The absolute outlet angle of the working fluid, in jth stage, is equal to the absolute inlet angle of the working fluid in (j+1)th stage. ? The effects of intake and outlet piping are neglected. The specific enthalpies at every station are as follows （1） （2） The total profile losses of the jth stage rotor and the stator are calculated as follows: （3） （4） Whereis the total profile loss coefficient of jth stage rotor-blade and is that of jth stage-stator blade. Fig. 1. Flow-path of a n-stage axial-flow compressor Fig. 2. Enthalpy–entropy diagram of a n-stage compressor Fig. 3. Velocity triangle of an intermediate stage Fig. 4. Enthalpy–entropy diagram of an intermediate stage. The blade profile loss-coefficients and are functions of parameters of the working fluid and blade geometry. They can be calculated using various methods and are considered to be constants. When and are functions of the parameters of the working fluid and blade geometry, the loss coefficients can be calculated using the method of Ref. [24], which was employed and described in Ref. [21]. The optimization problem can be solved using the iterative method: (1) First, select the original values of and and then calculate the parameters of the stage. (2) Secondly, calculate the values of and , and repeat the first step until the differences between the calculated values and the original ones are small enough. The work required by the jth stage is （5） The work required by the jth rotor is: （6） The degree of reaction of the jth stage compressor is defined as . Hence, one has （7） Where， are the velocity coefficients, and they are defined as: andThe constraint conditions can be obtained from the energy-balance equation for the one-dimensional flow （8） （9） 3. Mathematical model for the behaviour of the multi-stage compressor The compression work required by each stage is. The total compression work required by the multi-stage compressor is . The stagnation isentropic enthalpy rise of every stage is . The sum of the stagnation isentropic enthalpy rise of each stage is, while the stagnation isentropic enthalpy rise of the multi-stage compressor is . One has,The stagnation isentropic efficiency of the multi-stage axial-flow compressor is （10） The total energy-balance of a n-stage compressor gives: （11） Eq. (11) can be rewritten as …. (12) For convenience, in order to make the constraints dimensionless, some parameters are defined: （13） （14） （15） （16） Where are the aerodynamic functions, and , where is the stagnation sound velocity and ，is the relative area, is the relative density, where l is the height of the blade, and is flow coefficient. Introducing the isentropic coefficient used by Boiko [23], one has （17） Where （18） Therefore, the constraint conditions can be rewritten as： （19） （20） （21） and the stagnation isentropic efficiency of the multi-stage axial-flow compressor can be rewritten as （22） Where is isentropic work coefficient of the multi-stage. The isentropic work coefficient of each stage is defined as .Now the optimization problem is to search the optimal values of and for finding the maximum value of the objective function under the constraints of Eqs. (19)~(21). 4. Solution procedure Once the system variables, the objective function, and the constraints are defined, a suitable method has to be adopted to determine the values of the design variables that maximize the objective function while satisfying the given constraints. The present optimization model is a non-linear programming procedure with Table 1Relative areas for the stations Station () 1 2 3 4 5 6 7 Relative area 1 0.936 0.886 0.809 0.729 0.701 0.647 Table 2Original and optimal design plans 參數(shù) 上限 下限 原始數(shù)據(jù) 最佳數(shù)據(jù) =0.732 =0.732 =0.732 =0.6 =0.59 =0.59 =0.49 =0.59 54 90 80.5891 72.6858 74.9116 66.5570 35 90 49.50 45.00 45.00 45.00 54 90 84.1338 76.3431 77.55 68.2003 35 90 49.50 45.00 45.00 45.00 54 90 66.411 59.7080 69.0582 55.7046 35 90 49.5418 45.00 45.00 46.6157 54 90 89.99 90.00 90.99 89.6147 0 3 1.089 1.0459 1.0913 1.093 0 3 1.148 1.1474 1.1549 1.0798 0 3 1.424 1.3970 1.3900 1.2624 0 3 1.424 1.4117 1,。4198 1.2624 0 3 1.565 1.5372 1.6091 1.3345 0 3 1.618 1.6338 1.6671 1.4450 0.9020 0.9050 0.9074 0.8955 5. Numerical example In the calculations, ,, , , n = 3, R = 286.96 J/(kg·K), , and are set. The relative areas at every station are listed in Table 1. It should be pointed out that there will be some influence on the relation of the optimization objective with these dimensionless parameters if are functions of the working fluid parameters and geometry parameters of the flow-path configuration. However, the relation obtained will not change qualitatively. For a 3-stage compressor, there are 13 design variables and 7 constraint conditions. Besides, the lower and upper limit value constraints of the 13 design variables should also be considered in the calculations. The lower and upper limits of the optimization variables, the original design plan, and the optimization results for different flow coefficients and work coefficients are listed in Table 2. It can be seen that the optimization procedure is effective and practical. The calculations show that the optimal stagnation isentropic efficiency is an increasing function of the work coefficient and a decreasing function of the flow coefficient. The effect of the work coefficient on the optimal stagnation isentropic-efficiency is larger than that of the flow coefficient. Also for various values你of the flow coefficients and work coefficients, the optimal absolute exit-angle of the last stage always approaches . 6. Conclusion In this paper, the efficiency optimization of a multi-stage axial-flow compressor for a fixed flow shape has been studied using one-dimensional flow-theory. The universal characteristic relation of the compressor be haviour is obtained. Numerical examples are presented. The results can provide some guidance as to the performance analysis and optimization of the multi-stage compressor. This is a preliminary study. It will be necessary to use multi-objective numerical optimization techniques [11–13,20,21,25–29] and artificial neural network algorithms [10,19,30,31] for practical compressor optimization. References [1] Wall RA. Axial-flow compressor performance prediction. AGARD-LS-83 1976(June):4.1–4.34. 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Chen et al. / Applied Energy 85 (2008) 625–633 633 譯文： 題目 一個用來提供恒定空氣壓力的無人值守的智能化控制系統(tǒng)的空氣壓縮機的研究 Lingen Chen Jun Luo Fengrui Sun Chih Wu 摘要 對多級壓縮機的優(yōu)化設計模型，本文假設固定的流道形狀以入口和出口的動葉絕對角度，靜葉的絕對角度和靜葉及每一級的入口和出口的相對氣體密度作為設計變量，得到壓縮機基元級的基本方程和多級壓縮機的解析關系。用數(shù)值實例來說明多級壓縮機的各種參數(shù)對最優(yōu)性能的影響。 關鍵詞 軸流壓縮機 效率 分析關系 優(yōu)化 1 引言 軸流式壓縮機的設計是工藝技術的一部分，如果缺乏準確的預測將影響設計過程。至今還沒有公認的方法可使新的設計參數(shù)達到一個足夠精確的值，通過應用一些已經取得新進展的數(shù)值優(yōu)化技術，以完成單級和多級軸流式壓縮機的設計。計算流體動力學（CFD）和許多更準確的方法特別是發(fā)展計算的CFD技術，已經應用到許多軸流式壓縮機的平面和三維優(yōu)化設計。它仍然是使用一維流體力學理論用數(shù)值實例來計算壓縮機的最佳設計。Boiko通過以下假設提出了詳細的數(shù)學模型用以優(yōu)化設計單級和多級軸流渦輪：（1）固定的軸向均勻速度分布（2）固定流動路徑的形狀分布，并獲得了理想的優(yōu)化結果。陳林根等人也采用了類似的想法，通過假設一個固定的軸向速度分布的優(yōu)化設計提出了設計單級軸流式壓縮機一種數(shù)學模型。在本文中為優(yōu)化設計多級軸流壓縮機的模型，提出了假設一個固定的流道形狀，以入口和出口的動葉絕對角度，靜葉的絕對角度和靜葉及每一級的入口和出口的相對氣體密度作為設計變量，分析壓縮機的每個階段之間的關系，用數(shù)值實例來說明多級壓縮機的各種參數(shù)對最優(yōu)性能的影響。 2 基元級的基本方程 考慮圖1所示由n級組成的軸流壓縮機, 其某一壓縮過程焓熵圖和中間級的速度三角形見圖2和圖3，相應的中間級的具體焓熵圖如圖4，按一維理論作級的性能計算。按一般情況列出軸流壓縮機中氣體流動的能量方程和連續(xù)方程,工作流體和葉輪的速度。在不同級的軸向流速不為常數(shù)，即考慮, () 時的能量和流量方程。在下列假定下分析軸流壓縮機的工作: ·相對于穩(wěn)定回轉的動葉、靜葉和導向葉片機構, 氣體流動是穩(wěn)定的； ·流體是可壓縮、無黏性和不導熱的； ·通過級的流體質量流量為定值； ·在實際工質的情況下, 壓縮過程是均勻的； ·本級出口絕對氣流角為下一級進口角絕對氣流角； ·忽略進出口管道的影響。 在每一級的具體焓如下： （1） （2） 第階段的動葉和靜葉的焓值損失總額計算如下： （3） （4） 其中是第階段動葉葉片輪廓總損失系數(shù)，是第階段靜葉葉片輪廓總損 失的系數(shù)。 圖1 n級軸流式壓縮機的流量路徑。 葉片輪廓損失系數(shù)和是工作流體和葉片的幾何功能參數(shù)。它們可以使用各種方法及視作常量來計算。當和看做工作流體和葉片的幾何功能參數(shù)時，可以使用Ref迭代的方法來計算損失系數(shù)。使用迭代方法解決計算損失系數(shù)： （1）選擇和初始值，然后計算各級的參數(shù)。 （2）計算的，值，重復第一步，直到計算值和原值之間的差異足夠小。 第階段理論所需計算得： （5） 第階段實際所需計算得： 圖2 n級壓縮機的焓熵圖 圖3 中間級的速度三角形 圖4 中間級的焓熵圖 （6） 基元級反應度定義為。因此有： （7） 在這里，視作速度系數(shù),它們的計算為: 和 （8） （9） 3 級組的數(shù)學模型 壓縮機各級的比壓縮功為則總的比耗功為， 各級的滯止等熵能量頭為，則級組各級滯止等熵比壓縮功總和為，級組等熵比壓縮功為, 則為壓縮機的重熱系數(shù)。根據(jù)定義,多級壓縮機通流部分滯止等熵效率為： 求解確定各級能量頭的分配： （11） 方程式（11）同樣可以寫作： …. (12) 出于方便，一些參數(shù)簡化約束計算做了如下定義： （13） （14） （15） （16） 這里 是氣動力函數(shù)，在這里的是滯止聲速相對應的，且 是相對面積，是相對密度，是葉片高 是流量系數(shù)。 通過Boiko的論文引入等熵線系數(shù)，一個是： （17） 這里 （18） 因此約束條件也可寫作 （19） （20） （21） 在這里多級軸流式壓縮機滯止等熵線的效率計算如下： （22） 這里是多級壓縮機的等熵工作系數(shù)，每一級的等熵工作系數(shù)是。 現(xiàn)在的優(yōu)化問題是尋找和的最佳值，來找出在方程（19~21)約束下的目標函數(shù)的最大值。 4 結論 一旦這些系統(tǒng)和定義的常數(shù)按目標實現(xiàn)自己系統(tǒng)功能，在他最理想的環(huán)境下達到預計函數(shù)最大的程度。其呈現(xiàn)的并非是一個線性的而是一階梯函數(shù)。本優(yōu)化模型是（2n +1）約束功能和一個n級軸流壓縮機（4n + 1）變量的非線性規(guī)劃程序。例如改善外部法或SUMT法，對于這樣的問題Powell采用在無約束極小化技術與一維最小的拋物線插值方法。人們已經發(fā)現(xiàn)是非常有作用的。 表1 各級相對面積 級 () 1 2 3 4 5 6 7 相對面積 1 0.936 0.886 0.809 0.729 0.701 0.647 表2 原始數(shù)據(jù)和設計計劃 參數(shù) 上限 下限 原始數(shù)據(jù) 最佳數(shù)據(jù) =0.732 =0.732 =0.732 =0.6 =0.59 =0.59 =0.49 =0.59 54 90 80.5891 72.6858 74.9116 66.5570 35 90 49.50 45.00 45.00 45.00 54 90 84.1338 76.3431 77.55 68.2003 35 90 49.50 45.00 45.00 45.00 54 90 66.411 59.7080 69.0582 55.7046 35 90 49.5418 45.00 45.00 46.6157 54 90 89.99 90.00 90.99 89.6147 0 3 1.089 1.0459 1.0913 1.093 0 3 1.148 1.1474 1.1549 1.0798 0 3 1.424 1.3970 1.3900 1.2624 0 3 1.424 1.4117 1,。4198 1.2624 0 3 1.565 1.5372 1.6091 1.3345 0 3 1.618 1.6338 1.6671 1.4450 0.9020 0.9050 0.9074 0.8955 5 數(shù)值計算例子 在計算中，做，，，，，,則為0.04, 為0.025和為0.02的設置。表1列出了在每個級的相對面積。應當指出會有一些優(yōu)化目標的關系與這些量綱的影響是工作流體參數(shù)的功能和流動路徑的幾何參數(shù)設置。然而，得到的關系不會改變流體性質。對于3級壓縮機中，有13個設計變量和7個約束條件。此外，較低上限約束的13個設計變量的值也應考慮在計算中。優(yōu)化變量的上限和下限，原來的設計方案中優(yōu)化不同流量系數(shù)和工作系數(shù)的結果列于表2。由此可以看出，優(yōu)化程序