外文翻譯--螺桿壓縮機的設(shè)計

1、<p><b>　　英文原文</b></p><p>　　Screw Machine Design</p><p>　　Holmes, 1999 reported that even higher accuracy was achieved on the new Holroyd vitrifying thread-grinding machine, thus

2、 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.</p><p&g

3、t;　　Screw machines are used today for different applications both as compressors and expanders. </p><p>　　Fig. 1.4. Screw compressor mechanical parts</p><p>　　Fig. 1.5. Cross section of a screw

4、compressor with gear box</p><p>　　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 specifi

5、cation of a universal rotor configuration or set of working parameters, even for a restricted class of machines.</p><p>　　Industrial compressors are required to compress air, refrigerants and process gases.

6、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 com

7、pressors, especially for mobile applications, efficiency may be less important than size and cost.</p><p>　　Oil free compressed air is delivered almost exclusively by screw compressors. The situation is beco

8、ming 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 comp

9、ressors is expected in the next few years.</p><p>　　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 g

10、as 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

11、 pressure differences sometimes exceed 40 bars. Typically, for oil flooded air compression applications, the volumetric effic</p><p>　　1.1 Screw Compressor Practice</p><p>　　The Swedish company

12、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 t

13、hem 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 rel

14、atively new screw compressor companies h</p><p>　　and tool machine production for screw compressor rotors. </p><p>　　Despite the increasing popularity of screw compressors, public knowledge and

15、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 generatio

16、n and a reproducible presentation of a Russian asymmetric profile named SKBK. The</p><p>　　profile generation in his book was based on an envelope approach. Andreev, 1961 repeats the theory of screw profiles

17、 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, ga

18、ve a reproducible presentation of the SRM asymmetric profile, five years after it was patented, together with the classic Lysholm Profile.</p><p>　　Two textbooks have been published in German. Rinder, 1979,

19、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

20、 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 manufacture

21、r</p><p>　　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

22、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 mos

23、t widely quoted in reference literature on this topic. Ohman, 1999, introduced the “</p><p>　　All patented profiles were generated by a procedure but information on the methods used is hardly disclosed eithe

24、r 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 y

25、ears 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 al</p><p>　　There is a surprising lack of

26、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 redu

27、ce 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

28、the International Institution </p><p>　　There are three compressor conferences which deal exclusively or partly with screw compressors. These are the biennial compressor technology conference, held at Purdue

29、 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</p><p>　　of papers

30、 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 informatio

31、n 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

32、 papers such as that by Rinder, 1984, who presented</p><p>　　Many reference textbooks on gears give useful background for screw rotor profiling. However all of them are limited to the classical gear conjugat

33、e 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.</p><p>　　1.2 Recent Deve

34、lopments</p><p>　　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 a

35、bility 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.<

36、;/p><p>　　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 im

37、prove 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

38、 as small as possible.</p><p>　　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

39、 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.</p><p>　　Rotor housing

40、s 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.</p><p>　　Improvements in compr

41、essor 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.</p&g

42、t;<p>　　1.2.1 Rotor Profiles</p><p>　　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 g

43、enerate 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 ci

44、rcles with centres on the pitch circles generate a similar circle on another rotor. It is the same if the circle centr</p><p>　　Circles with centres offset from the pitch circles and other curves, like ellip

45、ses, 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.

46、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.</p><p>　　The symmetric circular profile consists o

47、f 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

48、 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 hyp

49、ocycloids. The SRM </p><p>　　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 whi

50、ch 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

51、 use of one coordinate system to define all the curves further simplifies the design process. Typicall</p><p>　　An efficient screw compressor needs a rotor profile which has a large flow cross section area,

52、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

53、 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 </p><p>　　The optimum

54、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

55、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.</p>

56、<p>　　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 c

57、ontact forces and ultimately rotor seizure. Hence the profile should be designed so that the risk of seizure is minimised.</p><p>　　The search for new profiles has been both stimulated and facilitated by rec

58、ent 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 intuiti

59、ve 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 le</p><

60、;p>　　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 favo

61、urable 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 becomin

62、g increasingly popular. Five lobes in the main rotor are suitable fo</p><p>　　Figure 1.6 shows pairs of screw compressor rotors plotted together for comparison. They are described by their commercial name or

63、 by a name which denotes their patent.</p><p>　　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

64、 application. The SRM asymmetric profile Shibbie, 1979, which historically appears to be the most successful screw compressor profile is near the top.</p><p>　　The next one is Astberg’s SRM “D” profiles 1982

65、.</p><p>　　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 favourabl

66、e 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 profi

67、le, the FuSheng, Lee, 1988 and the “Hyper”, Chia-Hsing, 1995 profile. All the profil</p><p><b>　　中文譯文</b></p><p>　　螺桿壓縮機的設(shè)計 </p><p>　　在1999年福爾摩斯就報道,建立在新的霍爾德玻璃化螺紋磨床上的準確

68、性甚至可以更高，這樣即使在大批量生產(chǎn)時也可以保持螺紋磨削機公差范圍在3之內(nèi)。這意味著,就轉(zhuǎn)子單獨生產(chǎn)而言,轉(zhuǎn)子之間的間隙可以小至12。</p><p>　　如今，螺桿機被使用在諸如壓縮機和擴展器等應(yīng)用上，為了能從它們獲取最優(yōu)性能，每個應(yīng)用是需要特定的設(shè)計和操作的。因此，對于一個限制類機器而言，它不可能通過通用轉(zhuǎn)子配置的規(guī)范或者工作參數(shù)的設(shè)置來生產(chǎn)高效率的機器。</p><p>　　圖1.4

69、 螺桿壓縮機機械零件</p><p>　　圖1.5 螺桿壓縮機與齒輪箱剖視圖 </p><p>　　壓縮空氣、制冷劑和工藝氣體需要工業(yè)壓縮機，而對于每種應(yīng)用，它們的設(shè)計必須由為獲取最理想的結(jié)果而有所不同。通常，那些長時間工作的制冷和制程氣體壓縮機必須具有高效；至于空氣壓縮機，尤其是在移動型壓縮機上,效率可能相比大小和成本而言不那么重

70、要。</p><p>　　無油壓縮的空氣幾乎完全是由螺桿壓縮機傳輸?shù)?。在工藝氣體壓縮方面，情況也正越來越相似。在制冷領(lǐng)域，往復(fù)式和葉片式壓縮機正不斷被螺桿機取代，而且可以預(yù)計在未來見年內(nèi)，制冷壓縮機的需求會有一個戲劇性的上升。</p><p>　　目前螺桿壓縮機的尺寸制造范圍覆蓋在以凸輪轉(zhuǎn)子直徑75—620mm和允許傳遞的壓縮氣體流量在0.6—600m³/min之間。在眾多螺桿壓

71、縮機中，對于單級無油潤滑壓縮機壓力比可以達到3.5，對于單級油潤滑壓縮機壓力比可以高達15，正常的壓力差有15巴，但最大壓力差有時超過40巴。通常,對于油潤滑空氣壓縮機的應(yīng)用,這些機器的容積效率現(xiàn)在已經(jīng)超過90%,特定的輸入功率已經(jīng)減少到就幾年前被視為高不可攀的數(shù)值。</p><p>　　1.1 螺桿壓縮機的實踐</p><p>　　瑞典的SRM技術(shù)公司是一個先驅(qū)，至今他們?nèi)匀辉诼輻U壓縮機

72、的實踐領(lǐng)域處于領(lǐng)導(dǎo)地位。諸如英國的Compair公司、比利時的Atlas-Copco公司、美國的Ingersol-Rand和Gardner Denver公司以及德國的GHH公司等其他公司都密切的跟隨他們，且York、Trane 以及 Carrier公司在將螺桿式壓縮機應(yīng)用于制冷和空調(diào)方面處于領(lǐng)先地位，日本螺桿壓縮機制造商，如 Hitachi, Mycom 和Kobe-Steel公司也都很有名。在中東和遠東，許多相關(guān)的螺桿壓縮機新公司已建

73、立。在中國、印度以及其他發(fā)展中國家的新興市場上開發(fā)新螺桿壓縮機工廠。英國公司Holroyd雖然沒有直接參與壓縮機的生產(chǎn)，但它仍然是螺桿轉(zhuǎn)子的最大的制造商，并且也是在工具設(shè)計和螺桿式壓縮機轉(zhuǎn)子機床生產(chǎn)的世界領(lǐng)導(dǎo)者。</p><p>　　盡管日益普及的螺桿壓縮機,但它們的基本知識和理解仍然是有限的。在1960年早期，關(guān)于三螺桿壓縮機的教科書已發(fā)表在俄羅斯。1960年Sakun做出了一個完整的關(guān)于圓形、橢圓形和擺線形齒

74、廓生成的描述以及一個關(guān)于俄羅斯的名為SKBK的非對稱剖視圖的可重復(fù)性演示。在他的書中這個輪廓的生成是基于包絡(luò)理論。</p><p>　　Andreev在1961年又闡述了關(guān)于螺桿輪廓理論，并且為轉(zhuǎn)子刀具輪廓生成理論做出了貢獻。而1964年Golovintsov教科書的內(nèi)容樸實無華，但其關(guān)于螺桿壓縮機的章節(jié)既有趣又詳實。1977年同樣在俄羅斯的Asomov給出了一個關(guān)于SRM技術(shù)的循環(huán)演示，這在5年之后連同Lysh

75、olm的經(jīng)典概論一起作為項專利技術(shù)。</p><p>　　兩套教材已在德國出版，1979年Rinder在SRM技術(shù)成為專利之后的第七年提出了一種基于齒輪理論重建的SRM非對稱型材料輪廓生成方法。而1988年Konka出版了一些關(guān)于螺桿壓縮機的工程方面的技術(shù)資料。</p><p>　　直到最近，許多涉及螺桿壓縮機的教科書已發(fā)表在英國。諸如1993年奧尼爾開創(chuàng)工業(yè)壓縮機和1994年Arbon開

76、創(chuàng)旋轉(zhuǎn)雙軸式壓縮機，一些壓縮機制造商在螺桿壓縮機方面的制造手冊和大量的制造小冊子為他們提供有用的信息,但這些要么是機密要么不是在公共領(lǐng)域中的。它們中的一些,比如SRM數(shù)據(jù)書,雖然可用但只有得到SRM專利文獻許可才可引用在螺桿壓縮機。</p><p>　　螺桿壓縮機的專利上有一個非常大的數(shù)量。在過去三十年中已經(jīng)出現(xiàn)數(shù)以千計的專利，其中開關(guān)磁阻電機獨自一種卻持有750項。這些專利涉及這些機器的各個方面，尤其是涉及到它

77、們的轉(zhuǎn)子輪廓。關(guān)于1952年Nilson的對稱剖面、1979年 Shibbie的非對稱剖面以及1982年Astberg的“D”剖面的SRM專利技術(shù)是在這個方面最廣泛引用的參考文獻，在1999年Ohoman又為SRM技術(shù)引入了“G”剖面。其他一些成功的剖面專利案例也可能提到，像1984年Atlas-Copco,Compair 和Hough、1974年 Gardner Denver和Edstroem 、1983年Hitachi和Kasuya

78、以及同年的Ingersoll-Rand和Bowman兩人。最近,幾個非常成功的專利被授予相對較小的公司如在1988年授予了由李后藤先生創(chuàng)立復(fù)盛公司以及在1995年授予了由Chia-Hsing創(chuàng)立的Hanbel企業(yè)。 由 Rinder, 1987, 和 Stosic, 1996提出來一種新型用齒條作為主要曲線的基礎(chǔ)的輪廓生成方法。</p><p>　　所有的剖面專利都是有程序生成的，但是無論是在專利或者陪同出版物中

79、使用方法的信息幾乎是不公開的，因此，要花很多年才能掌握這些程序。很多這樣的例子，比如Margolis在1977年發(fā)表了他的關(guān)于對稱圓形剖面的推導(dǎo),32年后這項技術(shù)已經(jīng)申請了專利；還有Rinder在1979年用齒輪嚙合標準來再現(xiàn)SRM不對稱剖面，9年后這也作為項專利出版。還有這樣一個例子，唐，1995，推導(dǎo)了SRM的“D”剖面分析作為他博士論文的一部分，13年后這也成為一項專利出版物。螺桿壓縮機的其他許多方面也獲得了專利。這些包括幾乎所有

80、最著名的特征，如：驅(qū)油后，轉(zhuǎn)子葉尖螺旋口的吸入和排出，軸向力補償，卸載，滑閥和省煤器端口，</p><p>　　其中大部分是由SRM提出的，然而,其它公司也熱衷于專利申請。總的印象是，專利專家在螺桿壓縮機的發(fā)展方面和工程師一樣重要。</p><p>　　在技術(shù)工藝上螺桿壓縮機出版物有一個令人驚訝的缺陷。1942和1966的 Lysholm文件是一個第二十世紀中期的例外，但他并沒有包括任何關(guān)

81、于他介紹的降低氣孔面積的細節(jié)分析的細節(jié)，因此，像Stosic，1997，1998，···等人一樣的論文可以被看作是一個例外。近年來，通過1992年的Stosic國際制冷機構(gòu)、1995年的Fujiwara、1996年的 IMechE以及Smith的論文、1994和1998年的Fleming以及1998年的Stosic，螺桿壓縮機材料的在期刊出版物已越來越普遍。這些使的人們可以獲得比歷屆總發(fā)表的多更多的信息，1

82、998年Stosic的論文是及時出版現(xiàn)代實踐的一個典型的例子。</p><p>　　有三個專門或部分處理螺桿壓縮機的壓縮機的會議。它們分別是在美國普渡大學舉辦的兩年一次的壓縮機技術(shù)會議、在英國舉辦的關(guān)于壓縮機及其系統(tǒng)的IMechE會議和在德國多特蒙德舉辦的“VDI空壓機動力會議”。盡管在這些事件上發(fā)表的螺桿壓縮機論文的數(shù)量很多,但只有其中的幾個包含有用關(guān)于分析轉(zhuǎn)子和壓縮機設(shè)計的信息。典型的可以被應(yīng)用的出版物有92

83、年Edstroem的論文、94年Stosic的論文、84年和90年Singh的論文以及92年Zhang的論文，這些論文可以從讀者從中獲取這方面信息，而且這些論文表明他們使用包絡(luò)理論來計算它們的轉(zhuǎn)子的一些幾何特征。多特蒙德會議給出了一些有趣的論文，比如84年Rider的論文，他提出了螺桿轉(zhuǎn)子輪廓的機架的成形方法，其中包括一個基于齒輪理論的可完全重復(fù)式模式。Hanjalic,1994 和Holmes, 1994給出了關(guān)于輪廓、制造和控制的更

84、多的細節(jié)分析。Kauder, 1994, 1998和Stosic, 1998的都是典型的成功的關(guān)于適用于實際工程的大學研究案例。 Sauls, 1994, 1998的可能會視為一個好工程案例。倫敦壓縮機會議會有一些有趣的論文，像</p><p>　　許多齒輪的參考教材為螺桿轉(zhuǎn)子輪廓提供了有用的基礎(chǔ)信息，然而他們都局限于古典齒輪共軛作用條件。在1968 以及1956–1994期間Litvin的論文可以看成是一個在給

85、出可直接用于螺桿壓縮機剖析的嚙合理論方面的例外做法。</p><p><b>　　1.2 近期的發(fā)展</b></p><p>　　螺桿壓縮機的高效運行主要依賴于適當?shù)霓D(zhuǎn)子設(shè)計，另外對于所有類型的壓縮機的成功設(shè)計有一個重要的需求，這種需求就是能夠準確預(yù)測對于任何參數(shù)變化所導(dǎo)致的性能的影響，這些變化涉及間隙的大小、轉(zhuǎn)子輪廓形狀、油或液體噴射位置和速度以及轉(zhuǎn)子直徑、比例和速

86、度。</p><p>　　如今，當之間間隙很緊湊并且內(nèi)部泄漏率縮小時，通過更多的引入改良后的設(shè)計原則是進一步的改進的唯一可能。而實現(xiàn)這個目標主要的要求就是提高轉(zhuǎn)子線型，從而使得通過壓縮機的內(nèi)部流動面積最大化而泄漏路徑最小化；與此同時，由于轉(zhuǎn)子表面之間相對運動而產(chǎn)生的內(nèi)部摩擦也盡可能減小。</p><p>　　盡管似乎看起來轉(zhuǎn)子仿形切削正處于一個全面發(fā)展的狀態(tài)，但這是不現(xiàn)實的，它在事實上確有

87、實質(zhì)性提高的空間。最有前景的似乎就是通過能夠生成強有力的但質(zhì)量卻很輕的轉(zhuǎn)子的齒條輪廓成形技術(shù)，這些生成的轉(zhuǎn)子有著高產(chǎn)量和較小的轉(zhuǎn)子接觸應(yīng)力，而且也能使用低粘度的潤滑劑。</p><p>　　有著優(yōu)良端口外形的轉(zhuǎn)子外殼可以使用多變量設(shè)計優(yōu)化技術(shù)來進行設(shè)計，這降低了流動損失，從而也允許有更高的轉(zhuǎn)子速度和更緊湊的機器結(jié)構(gòu)。</p><p>　　壓縮機軸承設(shè)計的改善近年來也取得了顯著成效。如今在某

88、些情況下，它也能處理流體潤滑，密封性能也越來越好。所有這些使得使用范圍越廣，螺桿壓縮機的性能也越高效。</p><p>　　1.2.1 轉(zhuǎn)子型線</p><p>　　主要用于螺桿壓縮機轉(zhuǎn)子型線生成的實踐就是在一個實際螺桿轉(zhuǎn)子上創(chuàng)建主曲線，以及用一些合理的共軛運動準則來在另一個螺桿轉(zhuǎn)子上生成相應(yīng)的二次曲線。任何曲線都可以作為基本曲線，但是傳統(tǒng)的圓是最常用的。在節(jié)圓上所有圓的圓心在另一個轉(zhuǎn)子上

89、生成一個相似的圓，類似的，如果圓心在轉(zhuǎn)子軸上也一樣。</p><p>　　圓心距離節(jié)圓和其他曲線的偏置有精確的相似，這些其他曲線有橢圓線、拋物線以及雙曲線。他們在另外的轉(zhuǎn)子上生成既成曲線，也即所謂的外擺線。無獨有偶，位于某一轉(zhuǎn)子上的點會破壞在另一個轉(zhuǎn)子上的內(nèi)擺線或者EPI。再過幾十年，生產(chǎn)轉(zhuǎn)子所需的技術(shù)對于一個主弧線得選擇是有限的，這個主弧線能夠推導(dǎo)出一個合理的二次型線的。</p><p>

90、;　　只包含圓線型的對稱圓弧型線、Lysholm的非對稱型線、除去節(jié)圓圓形的圓線介紹了在高壓側(cè)的一套擺線，并且形成了第一非對稱螺桿轉(zhuǎn)子型線。SRM非對稱型線在門轉(zhuǎn)子的低壓側(cè)配置了一個偏置圓，其后的SKBK型線在主要的轉(zhuǎn)子上置備了相同的偏置圓。在這兩種情況下，一種分析演變曲線作為內(nèi)擺線或者EPI被提出，而SRM的“D”型線專門由圓線構(gòu)成，幾乎所有的這些偏心都位于主要的或者門轉(zhuǎn)子上。所有專利后,在一次和二次轉(zhuǎn)子上賦予的主曲線,在其他轉(zhuǎn)子上的

91、既成曲線,所有的可能都基于經(jīng)典傳動裝置或其他一些類似的條件的衍生。最近,這個圓線已經(jīng)逐漸被其他曲線取代，如在富盛公司中的橢圓線型、在康普艾和日立公司的拋物線型以及在“hyper”型線中的雙曲線。在最新的資料中雙曲線似乎是給密封線長度以最佳轉(zhuǎn)子配比的最合適的替換。</p><p>　　另一個用于生產(chǎn)螺桿壓縮機轉(zhuǎn)子型線的曲線的實踐就是使用虛構(gòu)的或者“非物質(zhì)的”轉(zhuǎn)子。由于所有的嚙合方程式獨立于它們所表達的坐標系統(tǒng)中，對

92、于給定的使用坐標系統(tǒng)的曲線來定義主弧線是有可能的，而這種坐標系統(tǒng)獨立于者兩種轉(zhuǎn)子的。這就意味著在許多情況下，這些定義方程可以簡化，同時定義所有曲線的坐標系統(tǒng)的使用進一步簡化了使用設(shè)計過程。在轉(zhuǎn)子獨立坐標系統(tǒng)中模版是規(guī)定了的，同樣對于無限半徑的機架轉(zhuǎn)子的情況類似。由此，通過一個程序可以獲得一些轉(zhuǎn)子上的二次弧線，而這些程序就被稱為“機架生成”技術(shù)。第一次由Menssen，1977，出版的關(guān)于機架生成技術(shù)的專利缺乏實用性，但是便捷的使用了Ri

93、nder，1987的理論，而且最近Stosic，1996，提出了一個更好的型線生成技術(shù)。</p><p>　　一個高效的螺桿壓縮機需要有一個大流量的轉(zhuǎn)子型線，這種型線有大橫截面積的、一個短密封線和一個小的吹孔面積。橫截面積越大，在相同轉(zhuǎn)子尺寸和速度時流速就越快，較短的密封線和一個較小的氣孔減少泄漏，較高的流量和較小的泄漏率增加壓縮機容積效率，也就是流速傳遞的作為總流量加上泄露量總和的一小部分，這反過來又增加了絕熱

94、效率因為在內(nèi)壓循環(huán)氣體上浪費得能量就小了。</p><p>　　吹孔面積和流動區(qū)域大小的最佳選擇取決于壓縮機的工作，由于壓力差低，泄漏率會比較小，因此由一個大的橫截面面積實現(xiàn)的收益可能超過一個大的氣孔面積所帶來的相關(guān)損失。確定凸角多的最佳數(shù)量的情況也是類似的，因為更少的凸角數(shù)意味著更大的過流面積，但卻增加了它們之間的壓力差。</p><p>　　精確制造允許轉(zhuǎn)子間隙減少，盡管驅(qū)油，直接轉(zhuǎn)子

95、接觸的可能性就增加了。硬質(zhì)轉(zhuǎn)子的接觸導(dǎo)致門轉(zhuǎn)子變形、接觸力增加而最終轉(zhuǎn)子會失效。因此，輪廓的設(shè)計應(yīng)使失效發(fā)生的風險最小化。</p><p>　　新的配置文件的搜索已經(jīng)刺激和促進了近年來數(shù)學建模和計算機模擬技術(shù)，這些分析方法可以被組合來形成過程分析和優(yōu)化的有力工具，從而消除了反應(yīng)直觀變話的舊方法，驗證了繁瑣的試驗和錯誤的測試。因此，這種對螺桿轉(zhuǎn)子凸角型線優(yōu)化設(shè)計的方法基本上是在過去的幾年中設(shè)計的，可能會導(dǎo)致在不久的

96、將來機器性能的進一步改進。然而，壓縮機的幾何形狀和在它內(nèi)部的進化過程是如此復(fù)雜，以至于對于要得到成功的模型，大量的近似建模很需要。因此，在公開文獻報道的計算機模型和編碼往往在方法和各種現(xiàn)象進行建模的數(shù)學水平不同的，缺乏對比實驗驗證仍然阻礙了建模各概念的一個全面驗證。盡管如此，計算機建模和優(yōu)化正在穩(wěn)步獲得信用，并且也越來越多地采用改進設(shè)計。</p><p>　　大多數(shù)的螺桿壓縮機仍在主轉(zhuǎn)子制造了4葉片和外徑相同的兩

97、個轉(zhuǎn)子的門轉(zhuǎn)子上制造6葉片，這種結(jié)構(gòu)是使干燥和油潤滑壓縮機應(yīng)用具有有利特性的一種折中，而且也常用于空氣制冷或工藝氣體壓縮機中。然而其他配置，如5：6和5：7和最近的4:5和3：5系類正變得越來越流行。在主轉(zhuǎn)子上有5個葉片對于高壓壓縮機的壓力比是合適的，特別是如果結(jié)合了大螺旋角時。4：5的配置安排對于適度壓力比的油潤滑應(yīng)用而言已經(jīng)成為最佳比例。3：5的配置在干燥應(yīng)用下受到青睞，因為它在門轉(zhuǎn)子和主轉(zhuǎn)子間提供了一個高傳動比，這樣也可用來減少所

98、需的驅(qū)動軸之間的速度。</p><p>　　圖1.6顯示了雙螺桿壓縮機轉(zhuǎn)子畫在一起進行比較，他們是由他們的商業(yè)名稱或他們的專利名稱所描繪。</p><p>　　下一個是描繪的是1982Astberg的 SRM“D”線型。</p><p>　　轉(zhuǎn)子所提供最大尺寸是5：6的配置，因為它在一個小轉(zhuǎn)子尺寸上將大排放口、良好的負載特性的大位移結(jié)合在，成為最受歡迎的轉(zhuǎn)子組合。它

99、在空氣壓縮制冷和空調(diào)是成功的。該集團從SRM技術(shù)的“D”型線，其次是“Sigma”，Bammert1979的型線，富盛公司的李1988的以及“Hyper”Chia-Hsig1995的型線。所有顯示的型線是“轉(zhuǎn)子產(chǎn)生的”，它們之間的主要區(qū)別是在主葉片上，它是一種在主轉(zhuǎn)子上的偏置圓，一圓線分別接著一直線、橢圓線和雙曲線。雙曲線似乎是為了達到這個目的的最好的幾何解法，最后兩個是機架產(chǎn)生的Rinder的1984版的以及Stosic的1996版的