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編號 無錫太湖學院 畢業(yè)設計（論文） 相關(guān)資料 題目： 磁力式擰瓶機的設計及工程分析 信機 系 機械工程及自動化專業(yè) 學 號： 　　　 0923190　　　　 學生姓名： 仲曉斌 指導教師： 　何雪明（職稱：副教授） （職稱： ） 2013年5月25日 目 錄 一、畢業(yè)設計（論文）開題報告 二、畢業(yè)設計（論文）外文資料翻譯及原文 三、學生“畢業(yè)論文（論文）計劃、進度、檢查及落實表” 四、實習鑒定表  無錫太湖學院 畢業(yè)設計（論文） 開題報告 題目： 磁力式擰瓶機的設計及工程分析 信機 系 機械工程及自動化 專業(yè) 學 號： 0923190 學生姓名： 仲曉斌 指導教師： 何雪明（職稱：副教授 ） （職稱： ） 2012年11月12日 課題來源 來源于工廠 科學依據(jù)（包括課題的科學意義；國內(nèi)外研究概況、水平和發(fā)展趨勢；應用前景等） （1）課題科學意義 擰瓶機是封口機的一種，它廣泛用于玻璃瓶或PET 瓶的螺紋蓋封口。 這種封蓋事先加工出內(nèi)螺紋，螺紋有單頭或多頭之分，如藥瓶多用單頭螺紋，罐頭瓶多用多頭螺紋，是靠旋轉(zhuǎn)封蓋，將蓋旋緊于容器口部.由于螺紋蓋具有封口快捷、開啟方便及開啟后瓶又可重新旋上等優(yōu)點，所以一些不含氣的液料, 諸如飲料、酒類、調(diào)味料、化妝品及藥品、嬰兒食品等瓶包裝的封口中大量采用螺紋蓋封口。在大型的自動化灌裝線上, 擰瓶機一般與灌裝機聯(lián)動, 并且作一體機型設計,從而減小灌裝至封蓋的行程, 使生產(chǎn)線結(jié)構(gòu)更為緊湊。目前已有全自動洗瓶機、全自動灌裝機、全自動擰瓶機三合一的機型。 為了減少包裝破損和運輸重量，并滿足消費者的安全需要，許多大型零售商都要求飲料和食品生產(chǎn)商采用塑料包裝。由于螺紋蓋有封口快捷、開啟方便及開啟后瓶又可重新封好等優(yōu)點, 使其在許多產(chǎn)品的包裝中應用越來越廣泛, 諸如飲料、酒類、調(diào)味料、化妝品及藥品等瓶包裝的封口就大量采用螺紋蓋封口。為了提高自動化生產(chǎn)線上瓶裝產(chǎn)品密封包裝的旋蓋問題，提高生產(chǎn)效率，保證產(chǎn)品質(zhì)量，特進行本課題自動擰瓶機機構(gòu)的設計研究。 （2）擰瓶機的研究狀況及其發(fā)展前景 國內(nèi)外已經(jīng)有相當成熟的封口機技術(shù)，形成了相當成熟的生產(chǎn)線，各種有特定功能的封口機、擰瓶機也在生產(chǎn)生活中隨處可見，技術(shù)不斷創(chuàng)新和改良，形式多樣化發(fā)展。 目前國內(nèi)自主研發(fā)的擰瓶機存在可靠性低、穩(wěn)定性差、旋蓋質(zhì)量低、返工率高等問題，國內(nèi)灌裝生產(chǎn)線中廣泛使用的擰瓶機大多為直線式擰瓶機，采用瓶頸掛蓋。經(jīng)定位、預封后使蓋平穩(wěn)坐落在瓶口上，最后由皮帶對蓋頂部搓壓摩擦而將蓋旋緊。旋蓋頭主要結(jié)構(gòu)型式經(jīng)歷了彈簧摩擦片式和磁力耦合式2種。彈簧摩擦片式在滿足恒扭矩要求方面效果較差，如經(jīng)長時間使用后彈簧力會減小，摩擦片使用一段時間后也需進行更換和調(diào)整。目前，國內(nèi)普遍使用的旋蓋頭為磁力耦合式。 擰瓶機是飲料灌裝過程中旋緊瓶蓋的專用設備,工作時必須保證適宜的旋緊力矩。力矩過小, 瓶蓋旋不緊; 力矩過大, 易損壞瓶嘴和瓶蓋。為此, 我們在吸收國外同類先進設備的基礎(chǔ)上研制了一種利用磁力傳遞扭力矩實現(xiàn)瓶蓋旋緊的旋蓋頭, 能根據(jù)需要方便地設定、調(diào)整旋緊力矩的大小, 并能適用于不同高度的瓶子。 研究內(nèi)容 （1）擰瓶機總體結(jié)構(gòu)設計 進行擰瓶機結(jié)構(gòu)總體方案設計，分析擰瓶機功能組成部分，進行最優(yōu)化選擇設計，讓其實現(xiàn)。 （2）擰瓶機的組成以及各部件設計 包括圓柱凸輪、理蓋裝置、轉(zhuǎn)盤、輸送軌道和旋蓋頭的設計。 （3）擰瓶機傳動部分的設計 包括電動機的選擇、減速器的選擇、帶傳動的設計、軸的校核、鍵的選擇、滾動軸承的選擇和錐齒輪的計算等。 （4）擰瓶機控制系統(tǒng) 分析選用哪種控制系統(tǒng)比較好 擬采取的研究方法、技術(shù)路線、實驗方案及可行性分析 （1）明確設計依據(jù)、原則和技術(shù)要求。 （2）查閱資料，分析現(xiàn)有的擰瓶機的優(yōu)缺點，參考其方案設計確定本設計的整體方案， 并對該方案進行優(yōu)化設計。 （3）對于擰瓶機進行設計并進行總體結(jié)構(gòu)的設計。 （4）利用UG進行三維模型設計，檢查各個零部件之間是否存在干涉，導出重要零部件的二維圖 （5）寫出具體的說明書。 研究計劃及預期成果 研究計劃： 2012年11月12日-2013年1月20日：按照任務書要求查閱論文相關(guān)參考資料，填寫畢業(yè)設計開題報告書。 2013年1月21日-2013年3月15日：填寫畢業(yè)實習報告。 2013年3月16日-2013年3月22日：按照要求修改畢業(yè)設計開題報告。 2013年3月23日-2013年4月20日：學習并翻譯一篇與畢業(yè)設計相關(guān)的英文材料。 2013年4月22日-2013年5月3日：擰瓶機的總體設計，利用UG繪制擰瓶機簡單的3D模型。 2013年5月4日-2013年5月10日：擰瓶機的部件設計。利用UG繪制擰瓶機器各部件的3D模型。 2013年5月11日-2013年5月20日：畢業(yè)論文撰寫和修改，并用UG出圖。 預期成果： 旋蓋頭利用磁能產(chǎn)生的力來進行旋蓋，能夠適應不同高度的瓶子。生產(chǎn)效率達到了4000至5000瓶/時。 特色或創(chuàng)新之處 ① 適用于不同高度的瓶子。 ② 旋力可調(diào)、定位更加可靠。 ③ 利用通電產(chǎn)生磁力來進行旋蓋。 已具備的條件和尚需解決的問題 ① 實驗方案思路已經(jīng)非常明確，已經(jīng)具備使用利用UG進行三維制圖。 ② 使用UG繪圖的能力尚需加強。 ③ 不會仿真。 ④ 設計的擰瓶機還存在很多的不足，如某些地方的設計考慮的還不夠多，還需完善和改進，自動化程度還不夠高，成本較高等。 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(lǐng)導簽名： 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. Consequ