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编号无锡太湖学院毕业设计（论文）相关资料题目： 机械式拧瓶机的设计及工程分析 信机 系 机械工程及自动化专业学 号： 0923116学生姓名： 吴 建 军 指导教师： 何雪明 （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： 机械式拧瓶机的设计及工程分析 信机 系 机械工程及自动化 专业学 号： 0923116 学生姓名： 吴 建 军 指导教师： 何雪明 （职称：副教授 ） （职称： ）2012年11月20日 课题来源 工厂科学依据（1）课题科学意义 拧瓶机是自动灌装生产线的主要设备之一，用于玻璃瓶或PET瓶的螺纹盖封口。随着社会的发展和人民生活水平的提高，人们对产品的包装质量的要求也越来越高。由于螺纹盖具有封口快捷，开启方便及开启瓶后又可重新封好等优点，使其在许多产品的包装中应用越来越广泛，诸如饮料，酒类，调味料，化妆品及药品等瓶包装的封口就大量采用螺纹盖封口。目前现有的国产同类机型的封盖机的产量，速度和自动化程度都相对落后。为了适应现代包装机高速，高效和高可靠性生产的需要，研制了一种回转式拧瓶机，该机采用多工位回转式结构，机电气一体化，具有效率高，速度快，可靠性好和自动化程度高等优点。（2）拧瓶机的研究状况及其发展前景 提高自动化程度是包装机械发展重要的趋势。产品和产量居世界之首的美国十分重视白装机械与计算机紧密结合，实现机电一体化控制，将自动化操作程序、数据收集系统、自动检验系统更多用于包装机械之中。日本则长于微电子技术，用以开那个值包装机械，有效地促进了无人操作和自动化程度的提高。在计量、制造和技术性能等方面居于世界领先地位的德国也高度重视提高自动化程度。几年前，德国包装机械系统设计时，自动化技术在整个系统操作及运行中还占30%，现在已占到50%以上。研究内容 了解数拧瓶机的工作原理，国内外的研究发展现状； 完成拧瓶机总体方案设计； 完成零部件的选型计算、结构强度校核； 熟练掌握有关计算机绘图软件，并绘制装配图和零件图纸，折合A0不少于2.5张； 完成设计说明书的撰写，并翻译外文资料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）实验方案实验通过通过UG软件进行动态模拟，分析材料是否在预定寿命内失效，利用有限元分析方法，分析其数学求解原理，根据实际情况进行离散化，进而求解分析。仿真技术的运用可以消除加工程序中的错误，有效地检查出工件加工过程中可能存在的干涉，从而保证机床与人员的安全，提高加工效率，改善加工质量，显著降低生产成本。（2）研究方法通过参阅借来的参考资料和上网查阅相关信息，并对拧瓶机进行实体观察，认真研究上体机结构，了解拧瓶机工作原理，与指导老师交流来完成对拧瓶机上体结构的毕业设计研究计划及预期成果研究计划： 2012年11月10日-2013年1月25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；完成一篇英文文献翻译2013年1月26日-2月3日：进行专业实训并对UG软件学习使用；2013年2月3日-2月17日：填写毕业实习实训报告；2013年2月17日-3月9日：对拧瓶机进行整体结构设计；2013年3月10日-4月5日：开始拧瓶机主传动系统设计；2013年4月5日-5月5日：进给传动系统设计；2013年5月6日-6月1日：基础支撑件和辅助装置设计；并完善毕业论文以及相关资料，为答辩做好充分的准备。预期成果：（1） 达到预期的实验结论：按照计划安排进行本设计，完成的拧瓶机上体机基本可以实现预期功能,。熟练掌握有关计算机绘图软件，并绘制装配图和零件图纸，折合A0不少于2.5张；完成设计说明书的撰写，并翻译外文资料1篇。特色或创新之处（1）主题明确，有针对性,稳定, 易操作, 通用性强。 （2）使用简易，功能完善。 （3）使生产率得到较大的提高，一改以前的单线生产已具备的条件和尚需解决的问题 （1）技术条件：拧瓶机的总目标已经了解，同时明确了实现总目标应该采用的策略，对UG的能够较熟练的使用。 （2）尚未解决的问题：恰当地选择机架材料，刚度、强度、稳定性满足要求的前提下获得好的经济性。对拧瓶机的结构和整体设计还未确定 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Applications4.1 IntroductionThis chapter demonstrates the scope of the method developed for the three-dimensional analysis of a screw compressor. The CFD package used in this case was COMET developed by ICCM GmbH Hamburg, today a part of CD-Adapco. The analysis of the flow and performance characteristics of a number of types of screw machines is performed to demonstrate a variety of parameters used for grid generation and calculation.The first example is concerned with a dry air screw compressor. A common compressor casing is used with two alternative pairs of rotors. The rotors have identical overall geometric properties but different lobe profiles. The application of the adaptation technique enables convenient grid generation for geometrically different rotors. The results obtained by three dimensional modelling are compared with those derived from a one-dimensional model, previously verified by comparison with experimental data.The relative advantages of each rotor profile are demonstrated.The second example shows the application of three dimensional flow analysis to the simulation of an oil injected air compressor. The results, thus obtained, are compared with test results obtained by the authors from a compressor and test rig, designed and built at City University. They are presented in the form of both integral parameters and a p-indicator diagram. Calculations based on the assumptions of the laminar flow are compared to those of turbulent flow. The effect of grid size on the results is also considered and shown here.The third example gives the analysis of an oil injected compressor in an ammonia refrigeration plant.This utilises the real fluid property subroutines in the process calculations and demonstrates the blow hole area and the leakage flow through the compressor clearances.The fourth example presents two cases, one of a dry screw compressor to show the influence of thermal expansion of the rotor on screw compressor performance and one of a high pressure oil-flooded screw compressor to show the influence of high pressure loads upon the compressor performance. 4.2 Flow in a Dry Screw CompressorDry screw compressors are commonly used to produce pressurised air, free of any oil. A typical example of such a machine, similar in configuration to the compressor modelled, is shown in Figure 4-1. This is a single stage machine with 4 male and 6 female rotor lobes. The male and female rotor outer diameters are 142.380 mm and 135.820 mm respectively, while their centre lines are 108.4 mm apart. The rotor length to main diameter ratio l/d=1.77. Thus, the rotor length is 252.0 mm. The male rotor with wrap angle =248.40 is driven at a speed of 6000 rpm by an electric motor through a gearbox. The male and female rotors are synchronised through timing gears with the same ratio as that of the compressor rotor lobes i.e. 1.5. The female rotor speed is therefore 4000 rpm. The male rotor tip speed is then 44.7m/s, which is a relatively low value for a dry air compressor. The working chamber is sealed from its bearings by a combination of lip and labyrinth seals.Each rotor is supported by one radial and one axial bearing, on the discharge end, and one radial bearing on the suction end of the compressor. The bearings are loaded by a high frequency force, which varies due to the pressure change within the working chamber. Both radial and axial forces, as well as the torque change with a frequency of 4 times the rotational speed. This corresponds to 400Hz and coincides with the number of working cycles that occur within the compressor per unit time.Figure 4-1 Cross section of a dry screw compressorThe compressor takes in air from the atmosphere and discharges it to a receiver at a constant output pressure of 3 bar. Although the pressure rise is moderate, leakage through radial gaps of 150 m is substantial. In many studies and modelling ,procedures, volumetric losses are assumed to be a linear function of the cross sectional area and the square root of pressure difference, assuming that the interlobe clearance is kept more or less constant by the synchronising gears. The leakage through the clearances is then proportional to the clearance gap and the length of the leakage line. However, a large clearance gap is needed to prevent contact with the housing caused by rotor deformation due to the pressure and temperature changes within the working chamber. Hence, the only way to reduce leakage is to minimise the length of the sealing line. This can be achieved by careful design of the screw rotor profile. Although minimising,leakage is an important means of improving a screw compressor efficiency, it is not the only one. Another is to increase the flow area between the lobes and thereby increase the compressor flow capacity, thereby reducing the relative effect of leakage. Modern profile generation methods take these various effects into account by means of optimisation procedures which lead to enlargement of the male rotor interlobes and reduction in the female rotor lobes. The female rotor lobes are thereby strengthened and their deformation thus reduced. To demonstrate the improvements possible from rotor profile optimisation, a three dimensional flow analysis has been carried out for two different rotor profiles within the same compressor casing, as shown in Figure 4-2. Both rotors are of the “N” type and rack generated.Figure 4-2N Rotors, Case-1 upper, Case-2 lowerCase 1 is an older design, similar in shape to SRM “D” rotors. Its features imply that there is a large torque on the female rotor, the sealing line is relatively long and the female lobes are relatively weak.Case 2, shown on the bottom of Figure 4-2, has rotors optimised for operating on dry air. The female rotor is stronger and the male rotor is weaker. This results in higher delivery, a relatively shorter sealing line and less torque on the female rotor. All these features help to improve screw compressor performance.The results of these two analyses are presented in the form of velocity distributions in the planes defined by cross-sections A-A and B-B, shown in Figure 4-1.In the case of this study, the effect of rotor profile changes on compressor integral performance parameters can be predicted fairly accurately with one-dimensional models, even if some of the detailed assumptions made in such analytical models are inaccurate. Hence the integral results obtained from the three-dimensional analysis are compared with those from a one-dimensional model.4.2.1 Grid Generation for a Dry Screw CompressorIn Case-1, the rotors are mapped with 52 numerical cells along the interlobe on the male rotor and 36 cells along each interlobe on the female rotor in the circumferential direction. This gives 208 and 216 numerical cells respectively in the circumferential direction for the male and female rotors. A total of 6 cells in the radial direction and 97 cells in the axial direction is specified for both rotors. This arrangement results in a numerical mesh with 327090 cells for the entire machine. The cross section for the Case-1 rotors is shown in Figure 4-3. The female rotor is relatively thin and has a large radius on the lobe tip. Therefore, it is more easily mapped than in Case-2 where the tip radius is smaller, as shown in Figure 4-4.Figure 4-3 Cross section through the numerical mesh for Case-1 rotorsThe rotors in Case 2 are mapped with 60 cells along the male rotor lobe and 40 cells along the female lobe, which gives 240 cells along both rotors in the circumferential direction. In the radial direction, the rotors are mapped with 6 cells while 111 cells are selected for mapping along the rotor axis. Thus, the entire working chamber for this compressor has 406570 cells. In this case, different mesh sizes are applied and different criteria are chosen for the boundary adaptation of these rotors. The main adaptation criterion selected for the rotors is the local radius curvature with a grid point ratio of 0.3 to obtain the desired quality of distribution along the rotor boundaries. By this means, the more curved rotors are mapped with only a slight increase in the grid size to obtain a reasonable value of the grid aspect ratio. To obtain a similar grid aspect ratio without adaptation, 85 cells would have been required instead of 60 along one interlobe on the female rotor. This would give 510 cells in the circumferential direction on each of rotors. If the number of cells in the radial direction is also increased to be 8 instead of 6 but the number of cells along axis is kept constant, the entire grid would contain more then a million cells which would, in turn, result in a significantly longer calculation time and an increased requirement for computer memory. Figure 4-4 Cross section through the numerical mesh for Case-2 rotors4.2.2 Mathematical Model for a Dry Screw CompressorThe mathematical model used is based on the momentum, energy and mass conservation equations as given in Chapter 2. The equation for space law conservation is calculated in the model in order to obtain cell face velocities caused by the mesh movement. The system of equations is closed by Stokes, Fouriers and Ficks laws and the equation of state for an ideal gas. This defines all the properties needed for the solution of the governing equations.4.2.3 Comparison of the Two Different Rotor Profiles The results obtained for both Case 1 and Case 2 compressors are presented here. To establish the full range of working conditions and to obtain an increase of pressure from 1 to 3 bars between the compressor suction and discharge, 15 time steps were required. A further 25 time steps were then needed to complete the full compressor cycle. Each time step needed about 30 minutes running time on an 800 MHz AMD Athlon processor. The computer memory required was about 400 MB.In Figure 4-5 the velocity vectors in the cross and axial sections are compared. The top diagram is given for Case-1 rotors and the bottom one for Case-2. As may be seen, the Case 2 rotors realised a smoother velocity distribution than the Case 1 rotors. This may have some advantage and could have increased the compressor adiabatic efficiency by reduction in flow drag losses. In both cases, recirculation within the entrapped working chamber occurs as consequence of the drag forces in the air as shown in the figure. On the other hand, different fluid flow patterns can be observed in the suction port. The velocities within the working chambers and the suction and discharge ports are kept relatively low while the flow through the clearance gaps changes rapidly and easily reaches sonic velocity. Figure 4-5 Velocity field in the compressor cross section for Case1 and Case2 rotorsFigure 4-6 Velocity field in the compressor axial section for Case1 and Case2 rotorsThese differences are confirmed in the view of the vertical compressor section through the female rotor axis, shown in Figure 4-6. In Case 2, lower velocities are achieved not only in the working chamber but also in the suction and discharge ports. In the suction port, this is significant because of the fluid recirculation which appears at the end of the port. This recirculation causes losses which cannot be recovered later in the compression process. Therefore, many compressors are designed with only an axial port instead of both, radial and axial ports. Such a situation reduces suction dynamic losses caused by recirculation but, on the other hand, increases the velocity in the suction chamber which in turn decreases efficiency. Some of these problems can be avoided only by the design of screw compressor rotors with larger lobes and a bigger swept volume and a shape which allows the suction process to be completed more easily. However, rotor profile design based on existing one-dimensional procedures neglects flow variations in the ports and hence is inferior for this purpose. In such cases, only a full three dimensional approach such as this, will be effective. 中文译文应用4.1简介本章介绍了对螺杆压缩机的三维分析开发的方法的范围。在这种情况下，采用由ICCM GmbH Hamburg开发的CFD软件，现在是CD-Adapco的一部分。对一定数量的螺杆机器的类型的流程和性能特性的分析是用来展示用于栅格一代和演算的各种各样的参量。 第一个例子是关于一个干螺杆空气压缩机。一个常见的压缩机外壳是使用两个可选双转子。转子具有相同的整体几何性质但是有不同的叶剖面。适应技术的应用可以方便使网格生成几何不同的转子。三维模型得到的结果与从一个一维模型获得的那些比较，以前被核实与实验数据相比。演示了每个转子配置文件的相对优势。第二个例子显示了三维流动分析模拟注入油空气压缩机的应用。如此得到的结果与从压缩机的作者和试验台，设计和建造城市大学通过以下方式获得的测试结果进行了比较。他们提出了两个积分的形式参数和一个p-示意图。计算基于的假设是层流与湍流流动的那些进行比较。网格尺寸对计算结果的影响也被认为是在这里。第三个例子给出了油中注入的制冷压缩机的分析。这利用了现实的流体属性的过程中计算的子程序，并演示吹孔区域和通过压缩机的间隙泄漏流。第四个例子呈现两种情况，一是显示的干式螺杆压缩机的转子的螺杆式压缩机的性能，热膨胀的影响和高压油没螺杆式压缩机中的一个，以显示的影响高压负荷时压缩机的性能。4.2干燥螺丝压缩机的流程 干燥螺丝压缩机是常用的生产被加压的空气，不需要任何油。这样机器的一个典型的例子，在配置与被塑造的压缩机相似，在表4-1显示。这是一个有4个阳性和6个阴性转子叶单级机。阳性和阴性的转子外直径分别为142.380毫米和135.820毫米，而他们的中心线108.4毫米。转子长度的主直径比L / D = 1.77。因此，转子长度252毫米。阳转子与包角= 248.40在每分钟6000转的速度驱动，通过齿轮箱由一个电动马达。阳性和阴性的转子通过定时齿轮同步与压缩机转子裂片即1.5的相同比率。因此，阴性的转子转速为每分钟4000转。阳转子叶尖速度然后44.7米/ s，这是相对低的值，为干燥的空气压缩机。工作腔密封从它的轴承，由唇，迷宫式密封的组合。每个转子是由一个径向和轴向轴承和一个径向轴承在放电结束后吸入端的压缩机。轴承是由一个高频力加载，它会因在工作腔的压力变化而变化。径向和轴向的力，以及频率的旋转速度的4倍的转矩变化。这对应于400Hz和发生在压缩机内的每单位时间的工作周期数一致。 压缩机以空气从大气排到一个接收器3个恒定的输出压力。虽然压力上升是温和的，经过150径向间隙泄漏是巨大的。在许多研究和建模过程中，容积损失被认为是一个线性函数的横截面积和压差的平方根假设叶片间间隙保持或多或少不变的同步齿轮。然后，通过该间隙的泄漏间隙和泄漏管路的长度成比例。然而，一个大的间隙是必要的，以防止转子变形，由于工作腔内的压力和温度的变化所造成的与壳体接触。因此，减少泄漏的唯一方法是将密封线长度。这可以通过仔细的螺杆转子型线设计实现。尽管最小化泄漏是一个重要的手 图4-1 干式螺杆压缩机的截面段,提高了螺杆压缩机效率,却不是唯一的一个。另一个是提高叶流之间的区域,从而提高压气机叶流量,从而减少了相对效应的泄漏。现代配置生成方法把这些不同的影响考虑通过优化程序，导致扩大阳转子叶片和减少阴性转子叶。阴性的转子叶是加强及其变形从而降低。为了证明可能从转子齿形优化，改善已进行了三维流场计算在两个不同的转子型线在同一个压缩机壳体，如图4-2所示。生成两个转子的“N ”型和机架。例1是一个比较老的设计，形状类似SRM “D”的转子。它的特点意味着阴转子上，有一个大的转矩，密封线是比较长的相对较弱阴性叶。显示在图4-2的底部，例2的转子的优化操作在干燥的空气。阴性
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