螺杆压缩机的设计外文文献翻译、中英文翻译、外文翻译

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.

中文译文

螺杆压缩机的设计

在1999年福尔摩斯就报道,建立在新的霍尔德玻璃化螺纹磨床上的准确性甚至可以更

μ之内。这意味着,就转高，这样即使在大批量生产时也可以保持螺纹磨削机公差范围在3m

μ。

子单独生产而言,转子之间的间隙可以小至12m

如今，螺杆机被使用在诸如压缩机和扩展器等应用上，为了能从它们获取最优性能，每个应用是需要特定的设计和操作的。因此，对于一个限制类机器而言，它不可能通过通用转子配置的规范或者工作参数的设置来生产高效率的机器。

图1.4 螺杆压缩机机械零件

图1.5 螺杆压缩机与齿轮箱剖视图

压缩空气、制冷剂和工艺气体需要工业压缩机，而对于每种应用，它们的设计必须由为获取最理想的结果而有所不同。通常，那些长时间工作的制冷和制程气体压缩机必须具有高效；至于空气压缩机，尤其是在移动型压缩机上,效率可能相比大小和成本而言不那么重要。

无油压缩的空气几乎完全是由螺杆压缩机传输的。在工艺气体压缩方面，情况也正越来越相似。在制冷领域，往复式和叶片式压缩机正不断被螺杆机取代，而且可以预计在未来见年内，制冷压缩机的需求会有一个戏剧性的上升。

目前螺杆压缩机的尺寸制造范围覆盖在以凸轮转子直径75—620mm和允许传递的压缩气体流量在0.6—600m3/min之间。在众多螺杆压缩机中，对于单级无油润滑压缩机压力比可以达到3.5，对于单级油润滑压缩机压力比可以高达15，正常的压力差有15巴，但最大压力差有时超过40巴。通常,对于油润滑空气压缩机的应用,这些机器的容积效率现在已经超过90%,特定的输入功率已经减少到就几年前被视为高不可攀的数值。

1.1 螺杆压缩机的实践

瑞典的SRM技术公司是一个先驱，至今他们仍然在螺杆压缩机的实践领域处于领导地位。诸如英国的Compair公司、比利时的Atlas-Copco公司、美国的Ingersol-Rand和Gardner Denver公司以及德国的GHH公司等其他公司都密切的跟随他们，且York、Trane 以及Carrier 公司在将螺杆式压缩机应用于制冷和空调方面处于领先地位，日本螺杆压缩机制造商，如Hitachi, Mycom 和Kobe-Steel公司也都很有名。在中东和远东，许多相关的螺杆压缩机新公司已建立。在中国、印度以及其他发展中国家的新兴市场上开发新螺杆压缩机工厂。英国公司Holroyd虽然没有直接参与压缩机的生产，但它仍然是螺杆转子的最大的制造商，并且也是在工具设计和螺杆式压缩机转子机床生产的世界领导者。

尽管日益普及的螺杆压缩机,但它们的基本知识和理解仍然是有限的。在1960年早期，关于三螺杆压缩机的教科书已发表在俄罗斯。1960年Sakun做出了一个完整的关于圆形、椭圆形和摆线形齿廓生成的描述以及一个关于俄罗斯的名为SKBK的非对称剖视图的可重复性演示。在他的书中这个轮廓的生成是基于包络理论。

Andreev在1961年又阐述了关于螺杆轮廓理论，并且为转子刀具轮廓生成理论做出了贡献。而1964年Golovintsov教科书的内容朴实无华，但其关于螺杆压缩机的章节既有趣又详实。1977年同样在俄罗斯的Asomov给出了一个关于SRM技术的循环演示，这在5年之后连同Lysholm的经典概论一起作为项专利技术。

两套教材已在德国出版，1979年Rinder在SRM技术成为专利之后的第七年提出了一种基于齿轮理论重建的SRM非对称型材料轮廓生成方法。而1988年Konka出版了一些关于螺杆压缩机的工程方面的技术资料。

直到最近，许多涉及螺杆压缩机的教科书已发表在英国。诸如1993年奥尼尔开创工业压缩机和1994年Arbon开创旋转双轴式压缩机，一些压缩机制造商在螺杆压缩机方面的制造手册和大量的制造小册子为他们提供有用的信息,但这些要么是机密要么不是在公共领域中的。它们中的一些,比如SRM数据书,虽然可用但只有得到SRM专利文献许可才可引用在螺杆压缩机。

螺杆压缩机的专利上有一个非常大的数量。在过去三十年中已经出现数以千计的专利，

其中开关磁阻电机独自一种却持有750项。这些专利涉及这些机器的各个方面，尤其是涉及到它们的转子轮廓。关于1952年Nilson的对称剖面、1979年Shibbie的非对称剖面以及1982年Astberg的“D”剖面的SRM专利技术是在这个方面最广泛引用的参考文献，在1999年Ohoman又为SRM技术引入了“G”剖面。其他一些成功的剖面专利案例也可能提到，像1984年Atlas-Copco,Compair 和Hough、1974年Gardner Denver和Edstroem 、1983年Hitachi和Kasuya以及同年的Ingersoll-Rand和Bowman两人。最近,几个非常成功的专利被授予相对较小的公司如在1988年授予了由李后藤先生创立复盛公司以及在1995年授予了由Chia-Hsing创立的Hanbel企业。由Rinder, 1987, 和Stosic, 1996提出来一种新型用齿条作为主要曲线的基础的轮廓生成方法。

所有的剖面专利都是有程序生成的，但是无论是在专利或者陪同出版物中使用方法的信息几乎是不公开的，因此，要花很多年才能掌握这些程序。很多这样的例子，比如Margolis 在1977年发表了他的关于对称圆形剖面的推导,32年后这项技术已经申请了专利；还有Rinder 在1979年用齿轮啮合标准来再现SRM不对称剖面，9年后这也作为项专利出版。还有这样一个例子，唐，1995，推导了SRM的“D”剖面分析作为他博士论文的一部分，13年后这也成为一项专利出版物。螺杆压缩机的其他许多方面也获得了专利。这些包括几乎所有最著名的特征，如：驱油后，转子叶尖螺旋口的吸入和排出，轴向力补偿，卸载，滑阀和省煤器端口，其中大部分是由SRM提出的，然而,其它公司也热衷于专利申请。总的印象是，专利专家在螺杆压缩机的发展方面和工程师一样重要。

在技术工艺上螺杆压缩机出版物有一个令人惊讶的缺陷。1942和1966的Lysholm文件是一个第二十世纪中期的例外，但他并没有包括任何关于他介绍的降低气孔面积的细节分析的细节，因此，像Stosic，1997，1998，···等人一样的论文可以被看作是一个例外。近年来，通过1992年的Stosic国际制冷机构、1995年的Fujiwara、1996年的IMechE以及Smith的论文、1994和1998年的Fleming以及1998年的Stosic，螺杆压缩机材料的在期刊出版物已越来越普遍。这些使的人们可以获得比历届总发表的多更多的信息，1998年Stosic的论文是及时出版现代实践的一个典型的例子。

有三个专门或部分处理螺杆压缩机的压缩机的会议。它们分别是在美国普渡大学举办的两年一次的压缩机技术会议、在英国举办的关于压缩机及其系统的IMechE会议和在德国多特蒙德举办的“VDI空压机动力会议”。尽管在这些事件上发表的螺杆压缩机论文的数量很多,但只有其中的几个包含有用关于分析转子和压缩机设计的信息。典型的可以被应用的出版物有92年Edstroem的论文、94年Stosic的论文、84年和90年Singh的论文以及92年Zhang的论文，这些论文可以从读者从中获取这方面信息，而且这些论文表明他们使用包络理论来计算它们的转子的一些几何特征。多特蒙德会议给出了一些有趣的论文，比如84年Rider的论文，他提出了螺杆转子轮廓的机架的成形方法，其中包括一个基于齿轮理论的可完全重复式模式。Hanjalic,1994 和Holmes, 1994给出了关于轮廓、制造和控制的更多的细节分析。Kauder, 1994, 1998和Stosic, 1998的都是典型的成功的关于适用于实际工程的大学研究案例。Sauls, 1994, 1998的可能会视为一个好工程案例。伦敦压缩机会议会有一些有趣的论文，像是89年的Edstroem的以及99年的Stosic的等等。

许多齿轮的参考教材为螺杆转子轮廓提供了有用的基础信息，然而他们都局限于古典齿轮共轭作用条件。在1968 以及1956–1994期间Litvin的论文可以看成是一个在给出可直接用于螺杆压缩机剖析的啮合理论方面的例外做法。

1.2 近期的发展

螺杆压缩机的高效运行主要依赖于适当的转子设计，另外对于所有类型的压缩机的成功设计有一个重要的需求，这种需求就是能够准确预测对于任何参数变化所导致的性能的影响，这些变化涉及间隙的大小、转子轮廓形状、油或液体喷射位置和速度以及转子直径、比例和速度。

如今，当之间间隙很紧凑并且内部泄漏率缩小时，通过更多的引入改良后的设计原则是进一步的改进的唯一可能。而实现这个目标主要的要求就是提高转子线型，从而使得通过压缩机的内部流动面积最大化而泄漏路径最小化；与此同时，由于转子表面之间相对运动而产生的内部摩擦也尽可能减小。

尽管似乎看起来转子仿形切削正处于一个全面发展的状态，但这是不现实的，它在事实上确有实质性提高的空间。最有前景的似乎就是通过能够生成强有力的但质量却很轻的转子的齿条轮廓成形技术，这些生成的转子有着高产量和较小的转子接触应力，而且也能使用低粘度的润滑剂。

有着优良端口外形的转子外壳可以使用多变量设计优化技术来进行设计，这降低了流动损失，从而也允许有更高的转子速度和更紧凑的机器结构。

压缩机轴承设计的改善近年来也取得了显著成效。如今在某些情况下，它也能处理流体润滑，密封性能也越来越好。所有这些使得使用范围越广，螺杆压缩机的性能也越高效。

1.2.1 转子型线

主要用于螺杆压缩机转子型线生成的实践就是在一个实际螺杆转子上创建主曲线，以及用一些合理的共轭运动准则来在另一个螺杆转子上生成相应的二次曲线。任何曲线都可以作为基本曲线，但是传统的圆是最常用的。在节圆上所有圆的圆心在另一个转子上生成一个相似的圆，类似的，如果圆心在转子轴上也一样。

圆心距离节圆和其他曲线的偏置有精确的相似，这些其他曲线有椭圆线、抛物线以及双曲线。他们在另外的转子上生成既成曲线，也即所谓的外摆线。无独有偶，位于某一转子上的点会破坏在另一个转子上的内摆线或者EPI。再过几十年，生产转子所需的技术对于一个主弧线得选择是有限的，这个主弧线能够推导出一个合理的二次型线的。

只包含圆线型的对称圆弧型线、Lysholm的非对称型线、除去节圆圆形的圆线介绍了在高压侧的一套摆线，并且形成了第一非对称螺杆转子型线。SRM非对称型线在门转子的低压侧配置了一个偏置圆，其后的SKBK型线在主要的转子上置备了相同的偏置圆。在这两种情况下，一种分析演变曲线作为内摆线或者EPI被提出，而SRM的“D”型线专门由圆线构成，几乎所有的这些偏心都位于主要的或者门转子上。所有专利后,在一次和二次转子上赋予的主曲线,在其他转子上的既成曲线,所有的可能都基于经典传动装置或其他一些类似的条件的衍生。最近,这个圆线已经逐渐被其他曲线取代，如在富盛公司中的椭圆线型、在康普艾和日立公司的抛物线型以及在“hyper”型线中的双曲线。在最新的资料中双曲线似乎是

给密封线长度以最佳转子配比的最合适的替换。

另一个用于生产螺杆压缩机转子型线的曲线的实践就是使用虚构的或者“非物质的”转子。由于所有的啮合方程式独立于它们所表达的坐标系统中，对于给定的使用坐标系统的曲线来定义主弧线是有可能的，而这种坐标系统独立于者两种转子的。这就意味着在许多情况下，这些定义方程可以简化，同时定义所有曲线的坐标系统的使用进一步简化了使用设计过程。在转子独立坐标系统中模版是规定了的，同样对于无限半径的机架转子的情况类似。由此，通过一个程序可以获得一些转子上的二次弧线，而这些程序就被称为“机架生成”技术。第一次由Menssen，1977，出版的关于机架生成技术的专利缺乏实用性，但是便捷的使用了Rinder，1987的理论，而且最近Stosic，1996，提出了一个更好的型线生成技术。

一个高效的螺杆压缩机需要有一个大流量的转子型线，这种型线有大横截面积的、一个短密封线和一个小的吹孔面积。横截面积越大，在相同转子尺寸和速度时流速就越快，较短的密封线和一个较小的气孔减少泄漏，较高的流量和较小的泄漏率增加压缩机容积效率，也就是流速传递的作为总流量加上泄露量总和的一小部分，这反过来又增加了绝热效率因为在内压循环气体上浪费得能量就小了。

吹孔面积和流动区域大小的最佳选择取决于压缩机的工作，由于压力差低，泄漏率会比较小，因此由一个大的横截面面积实现的收益可能超过一个大的气孔面积所带来的相关损失。确定凸角多的最佳数量的情况也是类似的，因为更少的凸角数意味着更大的过流面积，但却增加了它们之间的压力差。

精确制造允许转子间隙减少，尽管驱油，直接转子接触的可能性就增加了。硬质转子的接触导致门转子变形、接触力增加而最终转子会失效。因此，轮廓的设计应使失效发生的风险最小化。

新的配置文件的搜索已经刺激和促进了近年来数学建模和计算机模拟技术，这些分析方法可以被组合来形成过程分析和优化的有力工具，从而消除了反应直观变话的旧方法，验证了繁琐的试验和错误的测试。因此，这种对螺杆转子凸角型线优化设计的方法基本上是在过去的几年中设计的，可能会导致在不久的将来机器性能的进一步改进。然而，压缩机的几何形状和在它内部的进化过程是如此复杂，以至于对于要得到成功的模型，大量的近似建模很需要。因此，在公开文献报道的计算机模型和编码往往在方法和各种现象进行建模的数学水平不同的，缺乏对比实验验证仍然阻碍了建模各概念的一个全面验证。尽管如此，计算机建模和优化正在稳步获得信用，并且也越来越多地采用改进设计。

大多数的螺杆压缩机仍在主转子制造了4叶片和外径相同的两个转子的门转子上制造6叶片，这种结构是使干燥和油润滑压缩机应用具有有利特性的一种折中，而且也常用于空气制冷或工艺气体压缩机中。然而其他配置，如5：6和5：7和最近的4:5和3：5系类正变得越来越流行。在主转子上有5个叶片对于高压压缩机的压力比是合适的，特别是如果结合了大螺旋角时。4：5的配置安排对于适度压力比的油润滑应用而言已经成为最佳比例。3：5的配置在干燥应用下受到青睐，因为它在门转子和主转子间提供了一个高传动比，这样也可用来减少所需的驱动轴之间的速度。

图1.6显示了双螺杆压缩机转子画在一起进行比较，他们是由他们的商业名称或他们的

专利名称所描绘。

下一个是描绘的是1982Astberg的SRM“D”线型。

转子所提供最大尺寸是5：6的配置，因为它在一个小转子尺寸上将大排放口、良好的负载特性的大位移结合在，成为最受欢迎的转子组合。它在空气压缩制冷和空调是成功的。该集团从SRM技术的“D”型线，其次是“Sigma”，Bammert1979的型线，富盛公司的李1988的以及“Hyper”Chia-Hsig1995的型线。所有显示的型线是“转子产生的”，它们之间的主要区别是在主叶片上，它是一种在主转子上的偏置圆，一圆线分别接着一直线、椭圆线和双曲线。双曲线似乎是为了达到这个目的的最好的几何解法，最后两个是机架产生的Rinder的1984版的以及Stosic的1996版的转子。这些基本曲线是在机架上进行选择和分配来创造一个有着比任何其他已知的螺杆压缩机转子具有较强门转子叶片的更大的横截面面积。