磁力式拧瓶机的设计及工程分析[三维UG]

喜欢就充值下载吧。。资源目录里展示的文件全都有，，请放心下载，，有疑问咨询QQ:414951605或者1304139763 ======================== 喜欢就充值下载吧。。资源目录里展示的文件全都有，，请放心下载，，有疑问咨询QQ:414951605或者1304139763 ======================== 编号 无锡太湖学院 毕业设计（论文） 相关资料 题目： 磁力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： 　　　 0923190　　　　 学生姓名： 仲晓斌 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： 磁力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： 0923190 学生姓名： 仲晓斌 指导教师： 何雪明（职称：副教授 ） （职称： ） 2012年11月12日 课题来源 来源于工厂 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） （1）课题科学意义 拧瓶机是封口机的一种，它广泛用于玻璃瓶或PET 瓶的螺纹盖封口。 这种封盖事先加工出内螺纹，螺纹有单头或多头之分，如药瓶多用单头螺纹，罐头瓶多用多头螺纹，是靠旋转封盖，将盖旋紧于容器口部.由于螺纹盖具有封口快捷、开启方便及开启后瓶又可重新旋上等优点，所以一些不含气的液料, 诸如饮料、酒类、调味料、化妆品及药品、婴儿食品等瓶包装的封口中大量采用螺纹盖封口。在大型的自动化灌装线上, 拧瓶机一般与灌装机联动, 并且作一体机型设计,从而减小灌装至封盖的行程, 使生产线结构更为紧凑。目前已有全自动洗瓶机、全自动灌装机、全自动拧瓶机三合一的机型。 为了减少包装破损和运输重量，并满足消费者的安全需要，许多大型零售商都要求饮料和食品生产商采用塑料包装。由于螺纹盖有封口快捷、开启方便及开启后瓶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。为了提高自动化生产线上瓶装产品密封包装的旋盖问题，提高生产效率，保证产品质量，特进行本课题自动拧瓶机机构的设计研究。 （2）拧瓶机的研究状况及其发展前景 国内外已经有相当成熟的封口机技术，形成了相当成熟的生产线，各种有特定功能的封口机、拧瓶机也在生产生活中随处可见，技术不断创新和改良，形式多样化发展。 目前国内自主研发的拧瓶机存在可靠性低、稳定性差、旋盖质量低、返工率高等问题，国内灌装生产线中广泛使用的拧瓶机大多为直线式拧瓶机，采用瓶颈挂盖。经定位、预封后使盖平稳坐落在瓶口上，最后由皮带对盖顶部搓压摩擦而将盖旋紧。旋盖头主要结构型式经历了弹簧摩擦片式和磁力耦合式2种。弹簧摩擦片式在满足恒扭矩要求方面效果较差，如经长时间使用后弹簧力会减小，摩擦片使用一段时间后也需进行更换和调整。目前，国内普遍使用的旋盖头为磁力耦合式。 拧瓶机是饮料灌装过程中旋紧瓶盖的专用设备,工作时必须保证适宜的旋紧力矩。力矩过小, 瓶盖旋不紧; 力矩过大, 易损坏瓶嘴和瓶盖。为此, 我们在吸收国外同类先进设备的基础上研制了一种利用磁力传递扭力矩实现瓶盖旋紧的旋盖头, 能根据需要方便地设定、调整旋紧力矩的大小, 并能适用于不同高度的瓶子。 研究内容 （1）拧瓶机总体结构设计 进行拧瓶机结构总体方案设计，分析拧瓶机功能组成部分，进行最优化选择设计，让其实现。 （2）拧瓶机的组成以及各部件设计 包括圆柱凸轮、理盖装置、转盘、输送轨道和旋盖头的设计。 （3）拧瓶机传动部分的设计 包括电动机的选择、减速器的选择、带传动的设计、轴的校核、键的选择、滚动轴承的选择和锥齿轮的计算等。 （4）拧瓶机控制系统 分析选用哪种控制系统比较好 拟采取的研究方法、技术路线、实验方案及可行性分析 （1）明确设计依据、原则和技术要求。 （2）查阅资料，分析现有的拧瓶机的优缺点，参考其方案设计确定本设计的整体方案， 并对该方案进行优化设计。 （3）对于拧瓶机进行设计并进行总体结构的设计。 （4）利用UG进行三维模型设计，检查各个零部件之间是否存在干涉，导出重要零部件的二维图 （5）写出具体的说明书。 研究计划及预期成果 研究计划： 2012年11月12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21日-2013年3月15日：填写毕业实习报告。 2013年3月16日-2013年3月22日：按照要求修改毕业设计开题报告。 2013年3月23日-2013年4月20日：学习并翻译一篇与毕业设计相关的英文材料。 2013年4月22日-2013年5月3日：拧瓶机的总体设计，利用UG绘制拧瓶机简单的3D模型。 2013年5月4日-2013年5月10日：拧瓶机的部件设计。利用UG绘制拧瓶机器各部件的3D模型。 2013年5月11日-2013年5月20日：毕业论文撰写和修改，并用UG出图。 预期成果： 旋盖头利用磁能产生的力来进行旋盖，能够适应不同高度的瓶子。生产效率达到了4000至5000瓶/时。 特色或创新之处 ① 适用于不同高度的瓶子。 ② 旋力可调、定位更加可靠。 ③ 利用通电产生磁力来进行旋盖。 已具备的条件和尚需解决的问题 ① 实验方案思路已经非常明确，已经具备使用利用UG进行三维制图。 ② 使用UG绘图的能力尚需加强。 ③ 不会仿真。 ④ 设计的拧瓶机还存在很多的不足，如某些地方的设计考虑的还不够多，还需完善和改进，自动化程度还不够高，成本较高等。 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 1 Introduction The screw compressor is one of the most common types of machine used to compress gases. Its construction is simple in that it essentially comprises only a pair of meshing rotors, with helical grooves machined in them, contained in a casing, which fits closely round them. The rotors and casing are separated by very small clearances. The rotors are driven by an external motor and mesh like gears in such a manner that, as they rotate, the space formed between them and the casing is reduced progressively. Thus, any gas trapped in this case is compressed. The geometry of such machines is complex and the flow of the gas being compressed within them occurs in three stages. Firstly, gas enters between the lobes, through an inlet port at one end of the casing during the start of rotation. As rotation continues, the space between the rotors no longer lines up with the inlet port and the gas is trapped and thus compressed. Finally, after further rotation, the opposite ends of the rotors pass a second port at the other end of the casing, through which the gas is discharged. The whole process is repeated between successive pairs of lobes to create a continuous but pulsating flow of gas from low to high pressure. These machines are mainly used for the supply of compressed air in the building industry, the food, process and pharmaceutical industries and, where required, in the metallurgical industry and for pneumatic transport. They are also used extensively for compression of refrigerants in refrigeration and air conditioning systems and of hydrocarbon gases in the chemical industry. Their relatively rapid acceptance over the past thirty years is due to their relatively high rotational speeds compared to other types of positive displacement machine, which makes them compact, their ability to maintain high efficiencies over a wide range of operating pressures and flow rates and their long service life and high reliability. Consequently, they constitute a substantial percentage of all positive displacement compressors now sold and currently in operation. The main reasons for this success are the development of novel rotor profiles, which have drastically reduced internal leakage, and advanced machine tools, which can manufacture the most complex shapes to tolerances of the order of 3 micrometers at an acceptable cost. Rotor profile enhancement is still the most promising means of further improving screw compressors and rational procedures are now being developed both to replace earlier empirically derived shapes and also to vary the proportions of the selected profile to obtain the best result for the application for which the compressor is required. Despite their wide usage, due to the complexity of their internal geometry and the non-steady nature of the processes within them, up till recently, only approximate analytical methods have been available to predict their performance. Thus, although it is known that their elements are distorted both by the heavy loads imposed by pressure induced forces and through temperature changes within them, no methods were available to predict the magnitude of these distortions accurately, nor how they affect the overall performance of the machine. In addition, improved modelling of flow patterns within the machine can lead to better porting design. Also, more accurate determination of bearing loads and how they fluctuate enable better choices of bearings to be made. Finally, if rotor and casing distortion, as a result of temperature and pressure changes within the compressor, can be estimated reliably, machining procedures can be devised to minimise their adverse effects. Screw machines operate on a variety of working fluids, which may be gases, dry vapour or multi-phase mixtures with phase changes taking place within the machine. They may involve oil flooding, or other fluids injected during the compression or expansion process, or be without any form of internal lubrication. Their geometry may vary depending on the number of lobes in each rotor, the basic rotor profile and the relative proportions of each rotor lobe segment. It follows that there is no universal configuration which would be the best for all applications. Hence, detailed thermodynamic analysis of the compression process and evaluation of the influence of the various design parameters on performance is more important to obtain the best results from these machines than from other types which could be used for the same application. A set of well defined criteria governed by an optimisation procedure is therefore a prerequisite for achieving the best design for each application. Such guidelines are also essential for the further improvement of existing screw machine designs and broadening their range of uses. Fleming et al., 1998 gives a good contemporary review of screw compressor modelling, design and application. A mathematical model of the thermodynamic and fluid flow processes within positive displacement machines, which is valid for both the screw compressor and expander modes of operation, is presented in this Monograph. It includes the use of the equations of conservation of mass, momentum and energy applied to an instantaneous control volume of trapped fluid within the machine with allowance for fluid leakage, oil or other fluid injection, heat transfer and the assumption of real fluid properties. By simultaneous solution of these equations, pressure-volume diagrams may be derived of the entire admission, discharge and compression or expansion process within the machine. A screw machine is defined by the rotor profile which is here generated by use of a general gearing algorithm and the port shape and size. This algorithm demonstrates the meshing condition which, when solved explicitly, enables a variety of rotor primary arcs to be defined either analytically or by discrete point curves. Its use greatly simplifies the design since only primary arcs need to be specified and these can be located on either the main or gate rotor or even on any other rotor including a rack, which is a rotor of infinite radius. The most efficient profiles have been obtained from a combined rotor-rack generation procedure. The rotor profile generation processor, thermofluid solver and optimizer,together with pre-processing facilities for the input data and graphical post processing and CAD interface, have been incorporated into a design tool in the form of a general computer code which provides a suitable tool for analysis and optimization of the lobe profiles and other geometrical and physical parameters. The Monograph outlines the adopted rationale and method of modelling, compares the shapes of the new and conventional profiles and illustrates potential improvements achieved with the new design when applied to dry and oil-flooded air compressors as well as to refrigeration screw compressors. The first part of the Monograph gives a review of recent developments in screw compressors. The second part presents the method of mathematical definition of the general case of screw machine rotors and describes the details of lobe shape specification. It focuses on a new lobe profile of a slender shape with thinner lobes in the main rotor, which yields a larger cross-sectional area and shorter sealing lines resulting in higher delivery rates for the same tip speed. The third part describes a model of the thermodynamics of the compression-expansion processes, discusses some modelling issues and compares the shapes of new and conventional profiles. It illustrates the potential improvements achievable with the new design applied to dry and oil-flooded air compressors as well as to refrigeration screw compressors. The selection of the best gate rotor tip radius is given as an example of how mathematical modelling may be used to optimise the design and the machine’s operating conditions. The fourth part describes the design of a high efficiency screw compressor with new rotor profiles. A well proven mathematical model of the compression process within positive displacement machines was used to determine the optimum rotor size and speed, the volume ratio and the oil injection position and jet diameter. In addition, modern design concepts such as an open suction port and early exposure of the discharge port were included, together with improved bearing and seal specification, to maximise the compressor efficiency. The prototypes were tested and compared with the best compressors currently on the market. The measured specific power input appeared to be lower than any published values for other equivalent compressors currently manufactured. Both the predicted advantages of the new rotor profile and the superiority of the design procedure were thereby confirmed. 1.1 Basic Concepts Thermodynamic machines for the compression and expansion of gases and vapours are the key components of the vast majority of power generation and refrigeration systems and essential for the production of compressed air and gases needed by industry. Such machines can be broadly classified by their mode of operation as either turbomachines or those of the positive displacement type. Turbomachines effect pressure changes mainly by dynamic effects, related to the change of momentum imparted to the fluids passing through them. These are associated with the steady flow of fluids at high velocities and hence these machines are compact and best suited for relatively large mass flow rates. Thus compressors and turbines of this type are mainly used in the power generation industry, where, as a result of huge investment in research and development programmes, they are designed and built to attain thermodynamic efficiencies of more than 90% in large scale power production plant. However, the production rate of machines of this type is relatively small and worldwide, is only of the order of some tens of thousands of units per annum. Positive displacement machines effect pressure changes by admitting a fixed mass of fluid into a working chamber where it is confined and then compressed or expanded and, from which it is finally discharged. Such machines must operate more or less intermittently. Such intermittent operation is relatively slow and hence these machines are comparatively large. They are therefore better suited for smaller mass flow rates and power inputs and outputs. A number of types of machine operate on this principle such as reciprocating, vane, scroll and rotary piston machines. In general, positive displacement machines have a wide range of application, particularly in the fields of refrigeration and compressed air production and their total world production rate is in excess of 200 million units per annum. Paradoxically, but possibly because these machines are produced by comparatively small companies with limited resources, relatively little is spent on research and development programmes on them and there are very few academic institutions in the world which are actively promoting their improvement. One of the most successful positive displacement machines currently in use is the screw or twin screw compressor. Its principle of operation, as indicated in Fig. 1.1, is based on volumetric changes in three dimensions rather than two. As shown, it consists, essentially, of a pair of meshing helical lobed rotors, contained in a casing. The spaces formed between the lobes on each rotor form a series of working chambers in which gas or vapour is contained. Beginning at the top and in front of the rotors, shown in the light shaded portion of Fig. 1.1a, there is a starting point for each chamber where the trapped volume is initially zero. As rotation proceeds in the direction of the arrows, the volume of that chamber then increases as the line of contact between the rotor with convex lobes, known as the main rotor, and the adjacent lobe of the gate rotor Fig. 1.1. Screw Compressor Rotors advances along the axis of the rotors towards the rear. On completion of one revolution i.e. 360◦ by the main rotor, the volume of the chamber is then a maximum and extends in helical form along virtually the entire length of the rotor. Further rotation then leads to reengagement of the main lobe with the succeeding gate lobe by a line of contact starting at the bottom and front of the rotors and advancing to the rear, as shown in the dark shaded portions in Fig. 1.1b. Thus, the trapped volume starts to decrease. On completion of a further 360◦ of rotation by the main rotor, the trapped volume returns to zero. The dark shaded portions in Fig. 1.1 show the enclosed region where therotors are surrounded by the casing, which fits closely round them, while the light shaded areas show the regions of the rotors, which are exposed to external pressure. Thus the large light shaded area in Fig. 1.1a corresponds to the low pressure port while the small light shaded region between shaft ends B and D in Fig. 1.1b corresponds to the high pressure port. Exposure of the space between the rotor lobes to the suction port, as their front ends pass across it, allows the gas to fill the passages formed between them and the casing until the trapped volume is a maximum. Further rotation then leads to cut off of the chamber from the port and progressive reduction in the trapped volume. This leads to axial and bending forces on the rotors and also to contact forces between the rotor lobes. The compression process continues until the required pressure is reached when the rear ends of the passages are exposed to the discharge port through which the gas flows out at approximately constant pressure. It can be appreciated from examination of Fig. 1.1, is that if the direction of rotation of the rotors is reversed, then gas will flow into the machine through the high pressure port and out through the low pressure port and it will act as an expander. The machine will also work as an expander when rotating in the same direction as a compressor provided that the suction and discharge ports are positioned on the opposite sides of the casing to those shown since this is effectively the same as reversing the direction of rotation relative to the ports. When operating as a compressor, mechanical power must be supplied to shaft A to rotate the machine. When acting as an expander, it will rotate automatically and power generated within it will be supplied externally through shaft A. The meshing action of the lobes, as they rotate, is the same as that of helical gears but, in addition, their shape must be such that at any contact position, a sealing line is formed between the rotors and between the rotors and the casing in order to prevent internal leakage between successive trapped passages. A further requirement is that the passages between the lobes should be as large as possible, in order to maximise the fluid displacement per revolution. Also, the contact forces between the rotors should be low in order to minimise internal friction losses. A typical screw rotor profile is shown in Fig. 1.2, where a configuration of 5–6 lobes on the main and gate rotors is presented. The meshing rotors are shown with their sealing lines, for the axial plane on the left and for the cross-sectional plane in the centre. Also, the clearance distribution between the two rotor racks in the transverse plane, scaled 50 times (6) is given above. Fig. 1.2. Screw rotor profile: (1) main, (2) gate, (3) rotor external and (4) pitch circles, (5) sealing line, (6) clearance distribution and (7) rotor flow area between the rotors and housing Oil injected Oil Free Fig. 1.3. Oil Injected and Oil Free Compressors Screw machines have a number of advantages over other positive displacement types. Firstly, unlike reciprocating machines, the moving parts all rotate and hence can run at much higher speeds. Secondly, unlike vane machines, the contact forces within them are low, which makes them very reliable. Thirdly, and far less well appreciated, unlike the reciprocating, scroll and vane machines, all the sealing lines of contact which define the boundaries of each cell chamber, decrease in length as the size of the working chamber decreases and the pressure within it rises. This minimises the escape of gas from the chamber due to leakage during the compression or expansion process. 1.2 Types of Screw Compressors Screw compressors may be broadly classified into two types. These are shown in Fig. 1.3 where machines with the same size rotors are compared: 1.2.1 The Oil Injected Machine This relies on relatively large masses of oil injected with the compressed gas in order to lubricate the rotor motion, seal the gaps and reduce the temperature rise during compression. It requires no internal seals, is simple in mechanical design, cheap to manufacture and highly efficient. Conseque