新式灌裝機(jī)的設(shè)計(jì)與工程分析【含CAD圖紙】

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無(wú)錫太湖學(xué)院 畢業(yè)設(shè)計(jì)（論文） 開(kāi)題報(bào)告 題目： 新式灌裝機(jī)的設(shè)計(jì)及工程分析 信機(jī) 系 機(jī)械工程及自動(dòng)化 專(zhuān)業(yè) 學(xué) 號(hào)： 0923001 學(xué)生姓名： 朱 艷 指導(dǎo)教師： 何雪明（職稱(chēng)：副教授 ） （職稱(chēng)： ） 2012年11月25日 課題來(lái)源 工廠 科學(xué)依據(jù)（包括課題的科學(xué)意義；國(guó)內(nèi)研究概況、水平和發(fā)展趨勢(shì)；應(yīng)用前景等） （1）課題科學(xué)意義 隨著科學(xué)技術(shù)的日新月異,人民的生活水平顯著提高,消費(fèi)習(xí)慣也發(fā)生了相應(yīng)的變化,人們對(duì)消費(fèi)品的包裝提出了更高的要求,而在包裝行業(yè)中占比例最大的是液態(tài)產(chǎn)品的包裝。這是由于液體包裝涉及多行業(yè)多品種,如飲料方面的果汁、牛奶、礦泉水、啤酒等;調(diào)味品方面的醬油、醋等;藥品方面的針劑、糖漿、氣霧劑等;農(nóng)藥乳劑、化工產(chǎn)品的各種瓶裝、化妝品等,要滿(mǎn)足日益增長(zhǎng)的液體產(chǎn)品的需要,就應(yīng)大力發(fā)展液體產(chǎn)品的包裝機(jī)械。 灌裝機(jī)是酒水、飲料類(lèi)等食品加工行業(yè)的關(guān)鍵設(shè)備之一。傳統(tǒng)罐裝機(jī)性能比較單一，自動(dòng)化程度低、通用性差，灌裝速度調(diào)整不方便，而且難以適用瓶形、液體物料及灌裝規(guī)格的變化。大量事實(shí)表明，實(shí)現(xiàn)灌裝的機(jī)械化和自動(dòng)化，尤其是實(shí)現(xiàn)具有高度靈活性的自動(dòng)包裝線，不僅體現(xiàn)了現(xiàn)代生產(chǎn)的發(fā)展方向，同時(shí)也可以獲得巨大的經(jīng)濟(jì)效益。它能改善產(chǎn)品質(zhì)量，加強(qiáng)市場(chǎng)競(jìng)爭(zhēng)能力；能改善勞動(dòng)條件，避免污染危害環(huán)境；能節(jié)約原材料，減少浪費(fèi)，降低成本；能提高生產(chǎn)效率，加速產(chǎn)品的不斷更新。實(shí)現(xiàn)灌裝機(jī)械化和建立現(xiàn)代包裝工業(yè)，將會(huì)更好地適應(yīng)市場(chǎng)的實(shí)際需要，更加合理地利用勞動(dòng)力，為社會(huì)多創(chuàng)造財(cái)富。 （2） 灌裝機(jī)的研究狀況及其發(fā)展前景 灌裝機(jī)械制造水平的發(fā)展體現(xiàn)在灌裝閥核心技術(shù)的發(fā)展上：從重力灌裝到先進(jìn)的抽真空等壓灌裝；從平面閥密封結(jié)構(gòu)到堆形閥密封結(jié)構(gòu)。這些都體現(xiàn)了技術(shù)的進(jìn)步與發(fā)展。我國(guó)的灌裝機(jī)械制造業(yè)，經(jīng)歷了仿制、引進(jìn)技術(shù)、消化吸收、創(chuàng)新、 自主開(kāi)發(fā)的過(guò)程，技術(shù)進(jìn)步及創(chuàng)新的速度更快，而且在不斷縮小與國(guó)外先進(jìn)技術(shù)之間的距離?，F(xiàn)代灌裝技術(shù)的目標(biāo)是精確、高效、自動(dòng)化。精確的灌裝量，灌裝過(guò)程的高速、可靠，盡量小的液損，整條生產(chǎn)線的最優(yōu)化控制，都由于電子技術(shù)的實(shí)際應(yīng)用而成為可能。隨著灌裝機(jī)械競(jìng)爭(zhēng)加劇，現(xiàn)代的科技技術(shù)發(fā)展正朝著以下方向發(fā)展：（1）、機(jī)械功能多元化（2）、結(jié)構(gòu)設(shè)計(jì)標(biāo)準(zhǔn)化、模組化（3）、控制智能化（4）、結(jié)構(gòu)運(yùn)動(dòng)高精度化未來(lái)的灌裝工業(yè)將配合產(chǎn)業(yè)自動(dòng)化趨勢(shì)，朝著研發(fā)技術(shù)、人才及高速包裝機(jī)等方向進(jìn)行，從而推動(dòng)灌裝行業(yè)的不斷前進(jìn)。 研究?jī)?nèi)容 ① 了解灌裝機(jī)的工作原理以及各個(gè)主要部分的工作性能，國(guó)內(nèi)外的研究發(fā)展現(xiàn)狀； ② 完成液體旋轉(zhuǎn)型灌裝機(jī)的總體方案設(shè)計(jì)； ③ 對(duì)比和分析完成對(duì)各個(gè)部件的選擇與設(shè)計(jì)的基本思想和開(kāi)發(fā)思路，對(duì)系統(tǒng)的各個(gè) 組成機(jī)構(gòu)進(jìn)行了整體一致性地分析； ④ 完成各零部件的選型計(jì)算、結(jié)構(gòu)強(qiáng)度校核； ⑤ 熟練使用UG繪圖軟件繪制三維立體圖，并繪制出處各部件的裝配工程圖及零件 工程圖； ⑥ 完成設(shè)計(jì)論文的撰寫(xiě)，并翻譯外文資料一篇。 擬采取的研究方法、技術(shù)路線、實(shí)驗(yàn)方案及可行性分析 （1） 研究方法 ① 搜尋資料，先確定整體灌裝機(jī)采用哪種輸送形式以及灌裝對(duì)象以及一些必要 的設(shè)計(jì)參數(shù)，確定各部分的尺寸及動(dòng)力分析。 ② 用UG繪制出三維實(shí)體動(dòng)力學(xué)模型，進(jìn)行仿真與分析，并及時(shí)修改。 （2） 技術(shù)路線 首先，通過(guò)對(duì)比和分析完成了對(duì)各個(gè)子系統(tǒng)環(huán)節(jié)的選擇與設(shè)計(jì)的基本思想和開(kāi)發(fā)思路；其次，對(duì)設(shè)備進(jìn)行了準(zhǔn)確度等級(jí)的檢測(cè)實(shí)驗(yàn)和商品實(shí)例分析;最后，對(duì)系統(tǒng)的各個(gè)組成機(jī)構(gòu)進(jìn)行了整體一致性地分析和控制。 （3） 實(shí)驗(yàn)方案 對(duì)灌裝機(jī)根據(jù)灌裝無(wú)氣液體的要求，選用在常壓狀態(tài)下的旋轉(zhuǎn)型灌裝機(jī)，從 而使它能連續(xù)工作，保證了工作效率。其主要部件灌裝閥采用控制液位定量式常 壓法灌裝閥，供瓶、托瓶機(jī)構(gòu)采用凸輪機(jī)構(gòu)、齒輪機(jī)構(gòu)等幾種常用機(jī)構(gòu)加以設(shè)計(jì) 分析 。 研究計(jì)劃及預(yù)期成果 研究計(jì)劃： 2012年11月10號(hào)-2013年11月30號(hào)：查閱有關(guān)于論文的參考資料，并初步填寫(xiě)畢業(yè) 設(shè)計(jì)開(kāi)題報(bào)告書(shū)； 2012年12月01號(hào)-2013年03月04號(hào)：學(xué)習(xí)UG軟件的使用，并以一減速機(jī)為例，進(jìn) 行繪制三維圖及工程圖； 2013年03月04號(hào)-2013年03月11號(hào)：學(xué)習(xí)并翻譯一篇與畢業(yè)設(shè)計(jì)相關(guān)的外文資料； 2013年03月12號(hào)-2013年04月28號(hào)：液體旋轉(zhuǎn)灌裝機(jī)的結(jié)構(gòu)設(shè)計(jì)以及UG圖； 2013年04月28號(hào)-2013年05月10號(hào)：液體旋轉(zhuǎn)灌裝機(jī)三維UG圖； 2013年05月11號(hào)-2013年05月19號(hào)：開(kāi)題報(bào)告的修改、畢業(yè)論文撰寫(xiě)及修改工作。 預(yù)期成果： 我國(guó)市場(chǎng)前景廣闊，產(chǎn)品質(zhì)量的提高逐漸滿(mǎn)足要求，因此產(chǎn)品不僅要有技術(shù)上的 完善，也要有外觀上質(zhì)量的提高。通過(guò)本課題的研究，產(chǎn)品必定以合理的色彩以及更加人性化的結(jié)構(gòu)方式來(lái)提高自己的附價(jià)值，從而更好的吸引顧客人群，提高自己的市場(chǎng)競(jìng)爭(zhēng)能力，從而創(chuàng)造更大的收益。 特色或創(chuàng)新之處 ① 使用UG軟件仿真，效果明顯，方便改變參量，能夠直觀判斷實(shí)驗(yàn)結(jié)果。 ② 采用固定某些參量、改變某些參量來(lái)研究問(wèn)題的方法，思路清晰，簡(jiǎn)潔明了，行之有效。 已具備的條件和尚需解決的問(wèn)題 ① 實(shí)驗(yàn)方案思路已經(jīng)非常明確，已經(jīng)具備使用UG繪制三維圖及操作方面的知識(shí)。 ② 使用UG模擬仿真的能力尚需加強(qiáng)。 指導(dǎo)教師意見(jiàn) 指導(dǎo)教師簽名： 年 月 日 教研室（學(xué)科組、研究所）意見(jiàn) 教研室主任簽名： 年 月 日 系意見(jiàn) 主管領(lǐng)導(dǎo)簽名： 年 月 日 英文原文 Screw Machine Design Holmes, 1999 reported that even higher accuracy was achieved on the new Holroyd vitrifying thread-grinding machine, thus keeping the manufacturing tolerances within 3 μm even in large batch production. This means that, as far as rotor production alone is concerned, clearances between the rotors can be as small as 12 μm. Screw machines are used today for different applications both as compressors and expanders. Fig. 1.4. Screw compressor mechanical parts Fig. 1.5. Cross section of a screw compressor with gear box For optimum performance from them a specific design and operating mode is needed for each application. Hence, it is not possible to produce efficient machines by the specification of a universal rotor configuration or set of working parameters, even for a restricted class of machines. Industrial compressors are required to compress air, refrigerants and process gases. For each application their design must differ to obtain the most desirable result. Typically, refrigeration and process gas compressors, which operate for long periods, must have a high efficiency. In the case of air compressors, especially for mobile applications, efficiency may be less important than size and cost. Oil free compressed air is delivered almost exclusively by screw compressors. The situation is becoming similar for the case of process gas compression. In the field of refrigeration, reciprocating and vane compressors are continuously being replaced by screw and a dramatic increase in the needs for refrigeration compressors is expected in the next few years. The range of screw compressors sizes currently manufactured is covered by male rotor diameters of 75 to 620mm and this permits the delivery of compressed gas flow rates of 0.6m3/min to 600m3/min. A pressure ratio of 3.5 is attainable in them from a single stage for dry compressors and up to 15 for oil flooded ones. Normal pressure differences are up to 15 bars, but maximum pressure differences sometimes exceed 40 bars. Typically, for oil flooded air compression applications, the volumetric efficiency of these machines now exceeds 90% and the specific power input has been reduced to values which were regarded as unattainable only a few years ago. 1.1 Screw Compressor Practice The Swedish company SRM was a pioneer and they are still leaders in the field of screw compressor practice. Other companies, like Compair U.K., Atlas- Copco in Belgium, Ingersol-Rand and Gardner Denver in the USA and GHH in Germany follow them closely. York, Trane and Carrier lead in screw compressor applications for refrigeration and air conditioning. Japanese screw compressor manufacturers, like Hitachi, Mycom and Kobe-Steel are also well known. Many relatively new screw compressor companies have been founded in the Middle and Far East. New markets in China and India and in other developing countries open new screw compressor factories. Although not directly involved in compressor production the British company, Holroyd, are the largest screw rotor manufacturer, they are world leaders in tool design and tool machine production for screw compressor rotors. Despite the increasing popularity of screw compressors, public knowledge and understanding of them is still limited. Three screw compressor textbooks were published in Russian in the early nineteen sixties. Sakun, 1960 gives a full description of circular, elliptic and cycloidal profile generation and a reproducible presentation of a Russian asymmetric profile named SKBK. The profile generation in his book was based on an envelope approach. Andreev, 1961 repeats the theory of screw profiles and makes a contribution to rotor tool profile generation theory. Golovintsov’s textbook, 1964, is more general but its section on screw compressors is both interesting and informative. Asomov, 1977, also in Russian, gave a reproducible presentation of the SRM asymmetric profile, five years after it was patented, together with the classic Lysholm Profile. Two textbooks have been published in German. Rinder, 1979, presented a profile generation method based on gear theory to reconstruct the SRM asymmetric profile, seven years after it was patented. Konka, 1988, published some engineering aspects of screw compressors. Only recently a number of textbooks have been published in English, which deal with screw compressors. O’Neill, 1993, on industrial compressors and Arbon, 1994, on rotary twin shaft compressors. There are a few compressor manufacturers’ handbooks on screw compressors and a number of brochures giving useful information on them, but these are either classified or not in the public domain. Some of them, like the SRM Data Book, although available only to SRM licensees, are cited in literature on screw compressors. There is an extraordinarily large number of patents on screw compressors. Literally thousands have appeared in the past thirty years, of which SRM, alone, holds 750. The patents deal with various aspects of these machines, but especially with their rotor profiles. The SRM patents of Nilson, 1952, for the symmetric profile, Shibbie, 1979, for the asymmetric profile and Astberg 1982, for the “D” profile are the most widely quoted in reference literature on this topic. Ohman, 1999, introduced the “G” profile for SRM. Other examples of successful profiling patents may also be mentioned, namely: Atlas-Copco,Compair with Hough, 1984, Gardner Denver with Edstroem, 1974, Hitachi with Kasuya, 1983, and Ingersoll-Rand with Bowman, 1983. More recently, several highly successful patents were granted to relatively small companies such as Fu Sheng, Lee, 1988, and Hanbel, Chia-Hsing, 1995. A new approach to profile generation, using a rack as the basis for the primary curves, was proposed by Rinder, 1987, and Stosic, 1996. All patented profiles were generated by a procedure but information on the methods used is hardly disclosed either in the patents or in accompanying publications. Thus it took many years before these procedures became known. Examples of this are: Margolis, who published his derivation of the symmetric circular profile in 1977, 32 years after it had been patented and Rinder, who used gear meshing criteria to reproduce the SRM asymmetric profile in 1979, 9 years after patent publication. It may also be mentioned that Tang, 1995, derived the SRM “D” profile analytically as part of a PhD thesis 13 years after the patent publication. Many other aspects of screw compressors were also patented. These include nearly all their most well known characteristics, such as, oil flooding, the suction and discharge ports following the rotor tip helices, axial force compensation, unloading, the slide valve and the economiser port, most of which were filed by SRM. However, other companies were also keen to file patents. The general impression gained is that patent experts are as important for screw compressor development as engineers There is a surprising lack of screw compressor publications in the technical literature. Lysholm’s papers in 1942 and 1966 were a mid twentieth century exception, but he did not include any details of the profiling details which he introduced to reduce the blow-hole area. Thus, journal papers like those of Stosic et al., 1997, 1998, may be regarded as an exception. In recent years, publication of screw compressor materials in journals has become more common through the International Institution of Refrigeration Stosic, 1992 and Fujiwara, 1995, the IMechE, with papers by Smith, 1996, Fleming, 1994, 1998 and Stosic 1998, and the ASME, by Hanjalic, 1997. Together these made more information available than the total published in all previous years. Stosic’s, 1998 paper is a typical example of the modern practice of timely publishing. There are three compressor conferences which deal exclusively or partly with screw compressors. These are the biennial compressor technology conference, held at Purdue University in the USA, the IMechE international conference on compressors and their systems, in England and the “VDI Schraubenkompressoren Tagung” in Dortmund, Germany. Despite the number of papers on screw compressors published at these events, only a few of them contain useful information on rotor profiling and compressor design. Typical Purdue papers cited as publications from which a reader can gain information on this are: Edstroem, 1992, Stosic, 1994 and Singh, 1984, 1990. Zhang, 1992,indicates that they used envelope theory to calculate some geometric features of their rotors. The Dortmund proceedings give some interesting papers such as that by Rinder, 1984, who presented the rack generation of a screw rotor profile, including a fully reproducible pattern based on gear theory. Hanjalic, 1994 and Holmes, 1994, give more details on profiling, manufacturing and control. Kauder, 1994, 1998 and Stosic, 1998 are typical examples of successful university reseach applied to real engineering. Sauls, 1994, 1998, may be regarded as an example of fine engineering work. The London compressor conference included some interesting papers like those of Edstroem, 1989 and Stosic et al., 1999. Many reference textbooks on gears give useful background for screw rotor profiling. However all of them are limited to the classical gear conjugate action condition. Litvin, 1968 and 1956–1994 may be regarded as an exception to this practice, in giving gearing theory which can be applied directly to screw compressor profiling. 1.2 Recent Developments The efficient operation of screw compressors is mainly dependent on proper rotor design. An additional and important requirement for the successful design of all types of compressor is an ability to predict accurately the effects on performance of the change in any design parameter such as clearance, rotor profile shape, oil or fluid injection position and rate, rotor diameter and proportions and speed. Now, when clearances are tight and internal leakage rates have become small, further improvements are only possible by the introduction of more refined design principles. The main requirement is to improve the rotor profiles so that the internal flow area through the compressor is maximised while the leakage path is minimised and internal friction, due to relative motion between the contacting rotor surfaces, is made as small as possible. Although it may seem that rotor profiling is now in a fully developed state, this is far from true. In fact there is room for substantial improvement. The most promising seems to be through rack profile generation which gives stronger but lighter rotors with higher throughput and lower contact stress. The latter enables lower viscosity lubricant to be used. Rotor housings with better shaped ports can be designed using a multivariable optimization technique. This reduces flow losses thus permitting higher rotor speeds and more compact machines. Improvements in compressor bearing design achieved in recent years now enable process fluid lubrication in some cases. Also seals are more efficient today. All these give scope for more effective and more efficient screw compressors. 1.2.1 Rotor Profiles The practice predominantly used for the generation of screw compressor rotor profiles is to create primary profile curves on one of the real screw rotors and to generate a corresponding secondary profile curve on the other rotor using some appropriate conjugate action criterion. Any curve can be used as a primary one, but traditionally the circle is the most commonly used. All circles with centres on the pitch circles generate a similar circle on another rotor. It is the same if the circle centres are at the rotor axes. Circles with centres offset from the pitch circles and other curves, like ellipses, parabolae and hyperbolae have elaborate counterparts. They produce generated curves, so called trochoids, on the other rotor. Similarly, points located on one rotor will cut epi- or hypocycloids on the other rotor. For decades the skill needed to produce rotors was limited to the choice of a primary arc which would enable the derivation of a suitable secondary profile. The symmetric circular profile consists of circles only, Lysholm’s asymmetric profile, apart from pitch circle centered circles, introduced a set of cycloids on the high pressure side, forming the first asymmetric screw rotor profile. The SRM asymmetric profile employs an offset circle on the low pressure side of the gate rotor, followed later by the SKBK profile introducing the same on the main rotor. In the both cases the evolved curves were given analytically as epi- or hypocycloids. The SRM “D” profile consists exclusively of circles, almost all of them eccentrically positioned on the main or gate rotor. All patents following, give primary curves on one rotor and secondary, generated curves on the other rotor, all probably based on derivations of classical gearing or some other similar condition. More recently, the circles have been gradually replaced by other curves, such as elipses in the FuSheng profiles, parabolae in the Compair and Hitachi profiles and hyperbolae in the “hyper” profile. The hyperbola in the latest profiles seems to be the most appropriate replacement giving the best ratio of rotor displacement to sealing line length. Another practice to generate screw rotor profile curves is to use imaginary, or “non-physical” rotors. Since all gearing equations are independent of the coordinate system in which they are expressed, it is possible to define primary arcs as given curves using a coordinate system which is independent of both rotors. By this means, in many cases the defining equations may be simplified. Also, the use of one coordinate system to define all the curves further simplifies the design process. Typically, the template is specified in a rotor independent coordinate system. The same is valid for a rotor of infinite radius which is a rack. From this, a secondary arc on some of the rotors is obtained by a procedure, which is called “rack generation”. The first ever published patent on rack generation by Menssen, 1977, lacks practicality but conveniently uses the theory. Rinder, 1987 and recently Stosic, 1996 give a better basis for profile Generation. An efficient screw compressor needs a rotor profile which has a large flow cross section area, a short sealing line and a small blow-hole area. The larger the cross section area the higher the flow rate for the same rotor sizes and rotor speeds. Shorter sealing lines and a smaller blow-hole reduce leakages. Higher flow and smaller leakage rates both increase the compressor volumetric efficiency, which is the rate of flow delivered as a fraction of the sum of the flow plus leakages. This in turn increases the adiabatic efficiency because less power is wasted in the compression of gas which is recirculated internally. The optimum choice between blow hole and flow areas depends on the compressor duty since for low pressure differences the leakage rate will be relatively small and hence the gains achieved by a large cross section area may outweigh the losses associated with a larger blow-hole. Similar considerations determine the best choice for the number of lobes since fewer lobes imply greater flow area but increased pressure difference between them. As precise manufacture permits rotor clearances to be reduced, despite oil flooding, the likelihood of direct rotor contact is increased. Hard rotor contact leads to deformation of the gate rotor, increased contact forces and ultimately rotor seizure. Hence the profile should be designed so that the risk of seizure is minimised. The search for new profiles has been both stimulated and facilitated by recent advances in mathematical modelling and computer simulation. These analytical methods may be combined to form a powerful tool for process analysis and optimisation and thereby eliminate the earlier approach of intuitive changes, verified by tedious trial and error testing. As a result, this approach to the optimum design of screw rotors lobe profiles has substantially evolved over the past few years and is likely to lead to further improvements in machine performance in the near future. However, the compressor geometry and the processes involved within it are so complex that numerous approximations are required for successful modelling. Consequently, the computer models and numerical codes reported in the open literature often differ in their approach and in the mathematical level at which various phenomena are modelled. A lack of comparative experimental verification still hinders a comprehensive validation of the various modelling concepts. In spite of this, computer modelling and optimization are steadily gaining in credibility and are increasingly employed for design improvement. The majority of screw compressors are still manufactured with 4 lobes in the main rotor and 6 lobes in the gate rotor with both rotors of the same outer diameter. This configuration is a compromise which has favourable features for both, dry and oil-flooded compressor applications and is used for air and refrigeration or process gas compressors. However, other configurations, like 5/6 and 5/7 and recently 4/5 and 3/5 are becoming increasingly popular. Five lobes in the main rotor are suitable for higher compressor pressure ratios, especially if combined with larger helix angles. The 4/5 arrangement has emerged as the best combination for oil-flooded applications of moderate pressure ratios. The 3/5 is favoured in dry applications, because it offers a high gear ratio between the gate and main rotors which may be taken advantage of to reduce the required drive shaft speed. Figure 1.6 shows pairs of screw compressor rotors plotted together for comparison. They are described by their commercial name or by a name which denotes their patent. The first group gives rotors with 4 lobes on the main and 6 lobes on the gate rotor. This rotor configuration is the most universally acceptable for almost any application. The SRM asymmetric profile Shibbie, 1979, which historically appears to be the most successful screw compressor profile is near the top. The next one is Astberg’s SRM “D” profiles 1982. The largest group of rotors presented is in 5/6 configuration which is becoming the most popular rotor combination because it combines a large displacement with large discharge ports and favourable load characteristics in a small rotor size. It is equally successful in air compression and in refrigeration and air-conditioning. The group starts with the SRM “D” profile, followed by the “Sigma”, Bammert, 1979 profile, the FuSheng, Lee, 1988 and the “Hyper”, Chia-Hsing, 1995 profile. All the profiles shown are “rotor-generated” and the main difference between them is in the leading lobe which is an offset circle on the main, a circle followed by a line, an ellipse and a hyperbola respectively. The hyperbola appears to be the best possible geometrical solution for that purpose. The last two are the rack generated rotors of Rinder, 1984 and Stosic, 1996. The primary curves for these were selected and distributed on a rack to create a larger cross section area with stronger gate rotor lobes than in any other known screw compressor rotor. 中文譯文 螺桿壓縮機(jī)的設(shè)計(jì) 在1999年福爾摩斯就報(bào)道,建立在新的霍爾德玻璃化螺紋磨床上的準(zhǔn)確性甚至可以更高，這樣即使在大批量生產(chǎn)時(shí)也可以保持螺紋磨削機(jī)公差范圍在3之內(nèi)。這意味著,就轉(zhuǎn)子單獨(dú)生產(chǎn)而言,轉(zhuǎn)子之間的間隙可以小至12。 如今，螺桿機(jī)被使用在諸如壓縮機(jī)和擴(kuò)展器等應(yīng)用上，為了能從它們獲取最優(yōu)性能，每個(gè)應(yīng)用是需要特定的設(shè)計(jì)和操作的。因此，對(duì)于一個(gè)限制類(lèi)機(jī)器而言，它不可能通過(guò)通用轉(zhuǎn)子配置的規(guī)范或者工作參數(shù)的設(shè)置來(lái)生產(chǎn)高效率的機(jī)器。 圖1.4 螺桿壓縮機(jī)機(jī)械零件 圖1.5 螺桿壓縮機(jī)與齒輪箱剖視圖 壓縮空氣、制冷劑和工藝氣體需要工業(yè)壓縮機(jī)，而對(duì)于每種應(yīng)用，它們的設(shè)計(jì)必須由為獲取最理想的結(jié)果而有所不同。通常，那些長(zhǎng)時(shí)間工作的制冷和制程氣體壓縮機(jī)必須具有高效；至于空氣壓縮機(jī)，尤其是在移動(dòng)型壓縮機(jī)上,效率可能相比大小和成本而言不那么重要。 無(wú)