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FRP 工作台 切割机 设计

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FRP工作台T形槽切割机的设计,FRP,工作台,切割机,设计

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编号无锡太湖学院毕业设计（论文）相关资料题目： FRP工作台T形槽切割机的设计 信机 系 机械工程及自动化专业学 号： 0923207学生姓名： 施 娟 指导教师： 唐正宁（职称：副教授 ） （职称： ）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: FRP工作台T形槽切割机的设计 信机 系 机械工程及自动化 专业学 号： 0923207 学生姓名： 施 娟 指导教师： 唐正宁 （职称：副教授 ） （职称： ） 2012年 11月 29日 课题来源FRP材料制作的Table top是Siemens公司生产的医疗设备中的一个部件，根据使用要求，必须在Table top上加工T形槽。为此，需要设计相应的专用加工设备。本课题来源于企业实际生产，将所学知识与实践紧密结合。科学依据（包括课题的科学意义；国内外研究概况、水平和发展趋势；应用前景等）一，课题的科学意义随着科技和经济的快速发展，社会对自动化设备的需求越来越大，FRP板的切割也不例外，它要求其实现的功能多，速度快，切割有力度，多方位切割，为此，FRP板材切割机也就被研发出来了。FRP-(Fiberglass-Reinforced Plastics )纤维增强复合塑料由增强纤维和基体组成。纤维（或晶须）的直径很小，一般在10m以下，缺陷较少又较小，断裂应变约为千分之三十以内，是脆性材料，易损伤、断裂和受到腐蚀。基体相对于纤维来说，强度、模量都要低很多，但可以经受住大的应变，往往具有粘弹性和弹塑性，是韧性材料。故FRP有如下特性：轻质高强、耐腐蚀性能好、电性能好、热性能良好、可设计性好、工艺性优良。本课题以设计为主线，设计一台切割机，用于在FRP工作台上加工T形槽。通过本设计不仅可以解决实际生产加工问题，而且可以了解产品设计过程，锻炼工程设计和解决实际问题的能力，为产品设计打下良好的基础。二，研究概况，水平和发展趋势在机械加工过程中，板材切割常用方式有手工切割、半自动切割机切割及数控切割机切割。手工切割灵活方便，但手工切割质量差、尺寸误差大、材料浪费大、后续加工工作量大，同时劳动条件恶劣，生产效率低。半自动切割机中仿形切割机，切割工件的质量较好，由于其使用切割模具，不适合于单件、小批量和大工件切割。其它类型半自动切割机虽然降低了工人劳动强度，但其功能简单，只适合一些较规则形状的零件切割。数控切割相对手动和半自动切割方式来说，可有效地提高板材切割的效率、切割质量，减轻操作者的劳动强度。目前在我国的一些中小企业甚至在一些大型企业中使用手工切割和半自动切割方式还较为普遍。三，应用前景FRP工作台T形槽切割机是一种专用设备，可以提高生产效率，切割精度比较高，使用轻巧简便，成本低廉。在能满足精度要求的前提下，不失为一种好的选择，应用前景良好。研究内容设计一种由电动机传动，以及气压系统固定夹紧的切割机，实现FRP工作台T行槽的切割，原理方案的功能实现，总体方案的设计，结构形式，结构参数，工作参数的设计。研究内容如下：1.阅读外文资料，并翻译与所学专业或课题相关的外文文献5000字左右，语句通顺、流畅、准确；2.根据加工产品上T形槽的具体结构和加工要求，拟定分析设备设计方案；3.运用三维软件进行三维造型设计、三维装配； 4.进行设备气动控制设计；5.绘制设备总装图和关键零部件二维工程图；6. 编写设计说明书，符合本科论文的格式要求，语言简洁、流畅、层次分明； 毕业设计作为重要的学习实践环节，本人在设计工作中做到积极主动，严谨细致、灵活，使设计方案、工程图纸符合国家标准。拟采取的研究方法、技术路线、实验方案及可行性分析各部分系统：电动机带动皮带轮，控制前后进程。气动系统进行固定。手动调节角度。确定动力问题后，采取先难后易的方法，先解决T型槽的角度如何改变，采用类似半圆仪的带有角度标志的半圆盘控制刀具的角度，使刀具可以小范围的转动，解决T型槽的角度问题。传动系统，电动机带动丝杠运转，丝杠上设计可以安装刀具装置，通过丝杠旋转运动，控制刀具的横向运动以电动机转速控制刀具运动速度。当FRP板材按正确位置放好后，采用杠杆原理与气动系统，使固定系统下移，将板材压紧在切割机底座上，从而实现板材固定。切割完成后，气压系统松动，释放压力即可。可行性分析：通过各主要部分系统的确定，主要问题可以解决。细节问题也可以方便的处理，总的来说，此方案的可行性没有问题。研究计划及预期成果1. 2012.11.1211.24 搜集、阅读、整理相关文献资料，英文翻译，明确切割机的功能，预期达到的性能指标等。2. 2012.11.2511.30 撰写开题报告。3. 2012.12.012013.3.1 完成实训。4. 2013.3.044.19 总体结构设计。电动机的选择，参数确定，强度计算，刀具的材料和类型的选择，刀具和重要部件的有限元分析，优化设计，通过三维造型设计，进行三维装配。完成设计规定量的设计图纸。 5. 2013.4.225.17 编写设计说明书。6. 2013.5.205.25 修改设计图纸，完成设计任务。 预期成果：完成设备的机械部分、气动部分、电气部分的设计及相关计算，绘制好机械部分设计的CAD三维图和主要二维图纸。 特色或创新之处 以电动机为动力源，结构紧凑，制造方便。割据角可调节，以适应要求形状的T形槽的加工。刀夹自行润滑，刀轮更换方便，加上传动装置，自动运行停止功能。该切割机具有手动和半自动相结合的特点，操作简便，灵活，运转平稳，从而使降低劳动强度，提高生产效率。已具备的条件和尚需解决的问题已具备的条件：1. 已了解切割机的工作原理和工作路线，对切割机的结构已经有初步的认识，也具备机械设计，包括电机的选择和刀具的选择等的能力。2. 能够运用软件进行设计、模拟。3. 相关资料齐全，有了初步可行的设计方案，技术条件具备。尚需解决的问题：设计方案有待完善，软件应用的熟练程度需要加强，刀具旋转精度与丝杠间隙的精度尚需进一步分析提高。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Design efficiency optimization of one-dimensional multi-stage axial-flow compressorAbstract 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; Optimization1. IntroductionThe 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 122.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 1720. 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 enthalpyspecific 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 enthalpyspecific 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 compressorFig. 2. Enthalpyentropy diagram of a n-stage compressorFig. 3. Velocity triangle of an intermediate stageFig. 4. Enthalpyentropy 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 compressorThe 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 procedureOnce 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 withTable 1Relative areas for the stationsStation ()1234567Relative area 10.9360.8860.8090.7290.7010.647Table 2Original and optimal design plans参数上限下限原始数据最佳数据=0.732=0.732=0.732=0.6=0.59=0.59=0.49=0.59549080.589172.685874.911666.5570359049.5045.0045.0045.00549084.133876.343177.5568.2003359049.5045.0045.0045.00549066.41159.708069.058255.7046359049.541845.0045.0046.6157549089.9990.0090.9989.6147031.0891.04591.09131.093031.1481.14741.15491.0798031.4241.39701.39001.2624031.4241.41171,。41981.2624031.5651.53721.60911.3345031.6181.63381.66711.44500.90200.90500.90740.89555. Numerical exampleIn the calculations, , , , n = 3, R = 286.96 J/(kgK), , 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. ConclusionIn 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 and artificial neural network algorithms for practical compressor optimization.中文译文：一维多级轴流压缩机性能的解析优化摘要 对多级压缩机的优化设计模型，本文假设固定的流道形状以入口和出口的动叶绝对角度，静叶的绝对角度和静叶及每一级的入口和出口的相对气体密度作为设计变量，得到压缩机基元级的基本方程和多级压缩机的解析关系。用数值实例来说明多级压缩机的各种参数对最优性能的影响。关键词 轴流压缩机 效率 分析关系 优化 1 引言轴流式压缩机的设计是工艺技术的一部分，如果缺乏准确的预测将影响设计过程。至今还没有公认的方法可使新的设计参数达到一个足够精确的值，通过应用一些已经取得新进展的数值优化技术，以完成单级和多级轴流式压缩机的设计。计算流体动力学（CFD）和许多更准确的方法特别是发展计算的CFD技术，已经应用到许多轴流式压缩机的平面和三维优化设计。它仍然是使用一维流体力学理论用数值实例来计算压缩机的最佳设计。Boiko通过以下假设提出了详细的数学模型用以优化设计单级和多级轴流涡轮：（1）固定的轴向均匀速度分布（2）固定流动路径的形状分布，并获得了理想的优化结果。陈林根等人也采用了类似的想法，通过假设一个固定的轴向速度分布的优化设计提出了设计单级轴流式压缩机一种数学模型。在本文中为优化设计多级轴流压缩机的模型，提出了假设一个固定的流道形状，以入口和出口的动叶绝对角度，静叶的绝对角度和静叶及每一级的入口和出口的相对气体密度作为设计变量，分析压缩机的每个阶段之间的关系，用数值实例来说明多级压缩机的各种参数对最优性能的影响。2 基元级的基本方程考虑图1所示由n级组成的轴流压缩机, 其某一压缩过程焓熵图和中间级的速度三角形见图2和图3，相应的中间级的具体焓熵图如图4，按一维理论作级的性能计算。按一般情况列出轴流压缩机中气体流动的能量方程和连续方程,工作流体和叶轮的速度。在不同级的轴向流速不为常数，即考虑, () 时的能量和流量方程。在下列假定下分析轴流压缩机的工作: 相对于稳定回转的动叶、静叶和导向叶片机构, 气体流动是稳定的； 流体是可压缩、无黏性和不导热的； 通过级的流体质量流量为定值；在实际工质的情况下, 压缩过程是均匀的；本级出口绝对气流角为下一级进口角绝对气流角；忽略进出口管道的影响。 在每一级的具体焓如下： （1） （2）第阶段的动叶和静叶的焓值损失总额计算如下： （3） （4）其中是第阶段动叶叶片轮廓总损失系数，是第阶段静叶叶片轮廓总损失的系数。 图1 n级轴流式压缩机的流量路径。叶片轮廓损失系数和是工作流体和叶片的几何功能参数。它们可以使用各种方法及视作常量来计算。当和看做工作流体和叶片的几何功能参数时，可以使用Ref迭代的方法来计算损失系数。使用迭代方法解决计算损失系数：（1）选择和初始值，然后计算各级的参数。（2）计算的，值，重复第一步，直到计算值和原值之间的差异足够小。第阶段理论所需计算得： （5）第阶段实际所需计算得： 图2 n级压缩机的焓熵图 图3 中间级的速度三角形 图4 中间级的焓熵图 （6）基元级反应度定义为。因此有： （7） 在这里，视作速度系数,它们的计算为:和 （8） （9）3 级组的数