外文翻译----设计加工螺杆式压缩机的内摆线
螺杆式制冷压缩机精简介绍PPT课件
④由于螺杆式制冷压缩机采用喷油方式，需要喷入大量油而
必须配置相应的辅助设备，从而使整个机组的体积和质量加
大。
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(4) 螺杆式制冷压缩机的性能参数及其计算
1)输气量和容积效率
①输气量
螺杆式制冷压缩机输气量的概念与活塞式相同，也是指压 缩机在单位时间内排出的气体，换算到吸气状态下的容积。
型号标记示例 示例1 LG16ⅡTA，表示转子名义直径为160mm，以R717 为制冷剂、特长导程、第二次改型的开启螺杆式单级制冷压缩 机。 示例2 BLG14-45G，表示转子名义直径为140mm、配用电 动机额定功率为45kW、用于高温名义工况的半封闭螺杆式单 级制冷压缩机。
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开启螺杆式制冷压缩机
机壳
螺杆式制冷 压缩机的机 壳一般为剖 分式。它由 机体（气缸 体）、吸气 端座、排气 端座及两端 端盖组成。
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机壳部件 1.吸气端盖；2.吸气端座；3.机体；4.排气端座；5.排气端盖
转子
转子结构 1.阴螺杆；2.阳螺杆
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转子是螺杆式 制冷压缩机的主要 部件，常采用整体 式结构，将螺杆与 轴做成一体。
长，维修简单，使用可靠，有利于实现操作自动化；
④螺杆式制冷压缩机对进液不敏感，可采用喷油或喷液冷却，
故在相同的压力比下，排气温度比活塞式制冷压缩机低得多，
因此单级压力比高；
⑤与离心式制冷压缩机相比，螺杆式制冷压缩机具有强制输气
的特点。即输气量几乎不受排气压力的影响。在较宽的工况范
围内，仍可保持较高的效率。
依靠啮合运动着的 一对阴阳转子，借助 它们的齿、齿槽与机 壳内壁所构成的呈 “V”字形的一对齿间 容积呈周期性大小变 化，来完成制冷剂气 体吸入—压缩—排出 的工作过程。
制冷压缩机教学第五章螺杆式制冷压缩机
2.效率
绝热效a率 dNNteh
制冷压缩机教学第五章螺杆式制冷 压缩机
指示效率 i
Nth Ni
机械效率 m
Ni Ne
绝热a效 dN N t e率 hN N t iN N heii m
3.输气系数
Va Vg
表征压缩机 的容积特性
制冷压缩机教学第五章螺杆式制冷 压缩机
制冷压缩机教学第五章螺杆式制冷 压缩机
滑阀的轴向移动，改变转子的有 效工作长度，达到输气量调节的目的。
制冷压缩机教学第五章螺杆式制冷 压缩机
螺杆压缩机的润滑系统
制冷压缩机教学第五章螺杆式制冷 压缩机
制冷压缩机教学第五章螺杆式制冷 压缩机
螺杆式压缩机的结构
一、螺杆式制冷压缩机的结构
1.机 壳
螺杆式制冷压 缩机的机壳一般为 剖分式。它由机体 （气缸体）、吸气 端座、排气端座及 两端端盖组成。
制冷压缩机教学第五章螺杆式制冷 压缩机
机体是连接各零部件的中心部件， 它为各零部件提供正确的装配位置，保 证阴、阳转子在气缸内啮合，可靠地进 行工作。
制冷压缩机教学第五章螺杆式制冷 压缩机
2.螺杆
转子是螺杆 式制冷压缩机的 主要部件，常采 用整体式结构， 将螺杆与轴做成 一体。
转子的毛坯常 为锻件，一般多采 用中碳钢 ，有特殊 要求时也有用40Cr 等合金材料。
制冷压缩机教学第五章螺杆式制冷 压缩机
3.轴承与油压平衡活塞
径向力大小与转子直径、长径比、内压力 比及运行工况有关。
圆周速度确定后，螺杆转速也随之 确定。喷油螺杆压缩机主动转子转速 范围为630～4400r/min。
制冷压缩机教学第五章螺杆式制冷 压缩机
螺杆压缩机机械外文文献翻译、中英文翻译、外文翻译
英文原文Screw CompressorThe Symmetric profile has a huge blow-hole area which excludes it from any compressor applicat -ion where a high or even moderate pressure ratio is involved. However, the symmetric profile per -forms surprisingly well in low pressure compressor applications.More details about the circular p -rofile can be found in Margolis, 1978.2.4.8 SRM “A” ProfileThe SRM “A” profile is shown in Fig. 2.11. It retains all the favourable features of the symmetric profile like its simplicity while avoiding its main disadvantage,namely, the large blow-hole area. The main goal of reducing the blow hole area was achieved by allowing the tip points of the main and gate rotors to generate their counterparts, trochoids on the gate and main rotor respectively. T -he “A” profile consists mainly of circles on the gate rotor and one line which passes through the gate rotor axis.The set of primary curves consists of: D2C2, which is a circle on the gate rotor with the centre on the gate pitch circle, and C2B2, which is a circle on the gate rotor, the centre of whi ch lies outside the pitch circle region.This was a new feature which imposed some problems in the generation of its main rotor counterpart, because the mathematics used for profile generation at tha -t time was insufficient for general gearing. This eccentricity ensured that the pressure angles on th -e rotor pitches differ from zero, resulting in its ease of manufacture. Segment BA is a circle on th -e gate rotor with its centre on the pitch circle. The flat lobe sides on the main and gate rotors weregenerated as epi/hypocycloids by points G on the gate and H on the main rotor respectively. GF2 is a radial line at the gate rotor. This brought the same benefits to manufacturing as the previously mentioned circle eccentricity onFig. 2.11 SRM “A” Profile2.4 Review of Most Popular Rotor Profiles 31 the opposite lobe side. F2E2 is a circle with the cent -re on the gate pitch and finally, E2D2 is a circle with the centre on the gate axis.More details on t -he “A” profile are published by Amosov et al., 1977 and by Rinder, 1979.The “A” profile is a go od example of how a good and simple idea evolved into a complicated result. Thus the “A” pro file was continuously subjected to changes which resulted in the “C” profile. This was mainly gen erated to improve the profile manufacturability. Finally, a completely new profile, the“D” profile was generated to introduce a new development in profile gearing and to increase the gate rotor tor -que.Despite the complexity o f its final form the “A” profile emerged to be the most popular scre -w compressor profile, especially after its patent expired.2.4.9 SRM “D” ProfileThe SRM “D” profile, shown in Fig. 2.12, is generated exclusively by circles with the centres off the rotor pitch circles.Similar to the Demonstrator, C2D2 is an eccentric circle of radius r3 onthe gate rotor. B1C1 is an eccentric circle of radius r1, which, together withthe small circular arc A1J1 of radius r2, is positioned on the main rotor. G2H2is a small circular arc on the gate rotor and E2F2 is a circular arc on the gaterotor. F2G2 is a relatively large circular arc on the gate rotor which produces a corresponding curve of the smallest possible curvature on the main rotor.Both circular arc, B2C2 and F2G2 ensure a large radius of curvature in the pitch circle area. This avoids high stresses in the rotor contact region.Fig. 2.12 SRM “D” ProfileThe “G” profile was introduced by SRM in the late nineteen nineties as a replacement for the “D” rotor and is shown in Fig. 2.13. Compared with the“D” rotor, the “G” rotor has the unique feature of two additional circles in the addendum area on both lobes of the main rotor, close to the pitch circle.This feature improves the rotor contact and, additionally, generates shorter sealing lines. This can be seen in Fig. 2.13, where a rotor featuring “G” profile characteristics only on its flat side through segment H1I1 is presented.Fig. 2.13 SRM “G” Profile2.4.11 City “N” Rack Generated Rotor Profile“N” rotors are calculated by a rack generation procedure. This distinguishes them from any others. In this case, the large blow-hole area, which is a characteristic of rack generated rotors, is overcome by generating the high pressure side of the rack by means of a rotor conjugate procedure. This undercuts the single appropriate curve on the rack. Such a rack is then used for profiling both the main and the gate rotors. The method and its extensions were used by the authors to create a number of different rotor profiles, some of them used by Stosic et al., 1986, and Hanjalic and Stosic, 1994. One of the applications of the rack generation procedure is described in Stosic, 1996.The following is a brief description of a rack generated “N” rotor profile,typical of a family of rotor profiles designed for the efficient compression of air,common refrigerants and a number of process gases. The rotors are generated by the combined rack-rotor generation procedure whose features are such that it may be readily modified further to optimize performance for any specific application.2.4 Review of Most Popular Rotor Profiles 33The coordinates of all primary arcs on the rack are summarized here relative to the rack coordinate system. The lobe of the rack is divided into several arcs. The divisions between the profile arcs are denoted by capital letters and each arc is defined separately, as shown in the Figs.2.14 and 2.15 where the rack and the rotors are shown.Fig. 2.14 Rack generated “N” ProfileFig. 2.15 “N” rotor primary curves g iven on rack34 2 Screw Compressor GeometryAll curves are given as a “general arc” expressed as: axp + byq = 1. Thus straight lines, circles, parabolae, ellipses and hyperbolae are all easily described by selecting appropriate values for parameters a, b, p and q.Segment DE is a straight line on the rack, EF is a circular arc of radius r4,segment FG is a straight line for the upper involute, p = q = 1, while segment GH on the rack is a meshing curve generated by the circular arc G2H2 on the gate rotor. Segment HJ on the rack is a meshing curve generated by the circular arc H1J1 of radius r2 on the main rotor. Segment JA is a circular arc of radius r on the rack, AB is an arc which can be either a circle or a parabola, a hyperbola or an ellipse, segment BC is a straight line on the rack matching the involute on the rotor round lobe and CD is a circular arc on the rack, radius r3.More details of the “N” profile can be found in Stosic, 1994.2.4.12 Characteristics of “N” ProfileSample illustrations of the “N” profile in 2-3, 3-5, 4-5, 4-6, 5-6, 5-7 and 6-7 configurations are given in Figs. 2.16 to Fig. 2.23. It should be noted that all rotors considered were obtained automatically from a computer code by simply specifying the number of lobes in the main and gate rotors, and the lobe curves in the general form.A variety of modified profiles is possible. The “N” profile design is a compromise between full tightness, small blow-hole area, large displacement.Fig. 2.16 “N” Rotors in 2-3 configurationFig. 2.17 “N” Rotors in 3-5 configurationFig. 2.18 “N” Rotors in 4-5 configurationFig. 2.19 “N” Rotors in 4-6 configurationFig. 2.20 “N” Rotors compared with “Sigma”, SRM “D” and “Cyclon” rotorsFig. 2.21 “N” Rotors in 5-6 configurationFig. 2.22 “N” Rotors in 5-7 configurationFig. 2.23 “N” rotors in 6/7 configurationsealing lines, small confined volumes, involute rotor contact and proper gate rotor torque distribution together with high rotor mechanical rigidity.The number of lobes required varies according to the designated compressor duty. The 3/5 arrangement is most suited for dry air compression, the 4/5 and 5/6 for oil flooded compressors with a moderate pressure difference and the 6/7 for high pressure and large built-in volume ratio refrigeration applications.Although the full evaluation of a rotor profile requires more than just a geometric assessment, some of the key features of the “N” profile may be readily appreciated by comparing it with three of the most popular screw rotor profiles already described here, (a) The “Sigma” profile by Bammert,1979, (b) the SRM “D” profile by Astberg 1982, and (c) the “Cyclon” profile by Hough and Morris, 1984. All these rotors are shown in Fig. 2.20 where it can be seen that the “N” profiles have a grea ter throughput and a stiffer gate rotor for all cases when other characteristics such as the blow-hole area, confined volume and high pressure sealing line lengths are identical.Also, the low pressure sealing lines are shorter, but this is less important because the corresponding clearance can be kept small.The blow-hole area may be controlled by adjustment of the tip radii on both the main and gate rotors and also by making the gate outer diameter equal to or less than the pitch diameter. Also the sealing lines can be kept very short by constructing most of the rotor profile from circles whose centres are close to the pitch circle. But, any decrease in the blow-hole area will increasethe length of the sealing line on the flat rotor side. A compromise betweenthese trends is therefore required to obtain the best result.2.4 Review of Most Popular Rotor Profiles 39Rotor instability is often caused by the torque distribution in the gate rotor changing direction during a complete cycle. The profile generation procedure described in this paper makes itpossible to control the torque on the gate rotor and thus avoid such effects. Furthermore, full involute contact between the “N” rotors enables any additional contact load to be absorbed more easily than with any other type of rotor. Two rotor pairs are shown in Fig. 2.24 the first exhibits what is described as “negative” gate rotor torque while the second shows the more usual “positive” torque.Fig. 2.24 “N” with negative torque, left and positive torque, right2.4.13 Blower Rotor ProfileThe blower profile, shown in Fig. 2.25 is symmetrical. Therefore only one quarter of it needs to be specified in order to define the whole rotor. It consists of two segments, a very small circle on the rotor lobe tip and a straight line. The circle slides and generates cycloids, while the straight line generates involutes.Fig. 2.25 Blower profile中文译文螺杆压缩机螺杆压缩机的几何形状对称分布有一个巨大的吹孔面积不包括它任何压缩机应用在高或中等压力比参与。
喷油螺杆压缩机的流量分析外文文献翻译、中英文翻译、外文翻译
中文译文4.3 在喷油螺杆压缩机的流量4.3.1 网格生成的油润滑压缩机阳极和阴极的转子有40个数值细胞沿各叶片间的圆周方向，6细胞在径向和轴向方向上的112。
这些形式为转子和壳体444830细胞总数。
为了避免需要增加网格点的数量，如果一个更精确的计算是必需的，一个适应的方法已应用于边界的定义。
时间变化的数量为25，在这种情况下，一个内部循环。
的对阳极的转子转一圈所需的时间步骤的总数是那么125。
在转子中的细胞数为每个时间步长保持相同。
以实现这一目标，一个特殊的网格移动程序开发中的时间通过压缩机转速的确定步骤，正如4章解释。
对于初始时间步长的数值网格图4-15提出。
图4数值网格喷油螺杆压缩机444830细胞4.3.2数学模型的油润滑压缩机数学模型的动量，能量，质量和空间方程问题，如第2.2节所描述的，但一个额外的方程的标量属性油的浓度的增加使石油对整个压缩机性能的影响进行计算。
本构关系是一样的前面的例子。
石油是一种被动的物种在模型处理，这不混合液体-空气的背景。
对空气的影响占通过物质和能量的来源是加上或减去的主要流模型相应的方程。
在这种情况下，动量方程通过拖曳力的影响如前所述。
建立工作条件和从吸气开始全方位1巴压力获得6，7压力的增加，8和9条近450000细胞放电，数值网格对于每一种情况下只有25时间步骤来获得所需的工作条件，其次是进一步的25的时间的步骤来完成一个完整的压缩机循环。
每个时间步所需的约30分钟的运行时间在一个800 MHz的AMD 速龙处理器计算机内存需要约450 MB。
4.3.3对油的数值模拟和实验结果的比较—淹没式压缩机在压缩机中的腔室，在压缩机内的循环的实验得到的压力历史和测得的空气流量和压缩机功率的情况下，测量的速度场担任了宝贵的基础，以验证CFD计算的结果。
要获得这些值，5/6喷油压缩机中，已经描述的，测试安装在压缩机实验室在城市大学伦敦，如图4-16上的钻机。
4-16喷油螺杆空气压缩机5 / 6-128mm（= 90mm）在测试床4.3流的喷油螺杆压缩机该试验台满足螺杆压缩机的接受所有pneurop /程序的要求试验。
螺杆压缩机的英文翻译
摘要】双螺杆压缩机是一种比较新颖的压缩机，因其可靠性高、操作维修方便、动力平衡性好、适应性强等优点，而广泛地应用于矿山、化工、动力、冶金、建筑、机械、制冷等工业部门。
双螺杆压缩机已经超过所有工业压缩机的50 ％，其市场份额超过80 ％，今后其市场份额还将继续扩大。
可见，研究双螺杆压缩机具有十分重要的意义。
本课题主要是设计通用的喷油双螺杆空气压缩机，采用单边不对称摆线-销齿圆弧型型线，阴、阳转子齿数比为6：4。
设计新型转子型线，目的是使接触线长度、泄漏三角形面积和封闭余隙容积3者达到最优化设计，以进一步提高双螺杆压缩机的机械性能。
重点研究的是双螺杆压缩机的转子型线设计、几何特性、受力分析、热力学计算。
【关键词】双螺杆压缩机转子型线啮合线齿间容积[Abstract] The twin-screw compressor is a kind of newly emerging compressor. Because of its high reliability, easy repair, good balance and good adaptability etc, and widely applied to such industrial departments as mine, chemical industry, power, metallurgy, architecture, machinery, refrigeration, etc. By designing the project, the volumetric efficiency is 70%, the compressed temperature is more 80℃。
It is very important to design and research a twin-screw compressor in industrial. The project is to design a universaltwin-screw air compressor, and to adopt single side asymmetric swept line unilaterally and dowel tooth circular rotor profile. There are six lobes on the female rotor and four lobes on the male rotor. The aim of designing a new rotor profile is to optimize the contact line length, blowhole area and clearance volume. That can improve the mechanical performance of a twin-screw compressor further. The project is mainly to research a twin-screw compressor rotor profile, geometry characteristic, mechanics analysis, thermodynamics calculation[Keywords] A twin-screw compressor, rotor profile, mesh curve, tooth space volume。
螺杆式制冷压缩机的工作原理及结构
螺杆式制冷压缩机的⼯作原理及结构螺杆式制冷压缩机的⼯作原理及结构第⼀节螺杆式制冷压缩机的⼯作原理1、螺杆式制冷压缩机的特点与活塞压缩机的往复容积式不同，螺杆式压缩机是⼀种回转容积式压缩机。
与活塞压缩机相⽐，螺杆式制冷压缩机有以下优点：a.体积⼩重量轻，结构简单，零部件少，只相当于活塞压缩机的1/3～1/2；b.转速⾼，单机制冷量⼤；c.易损件少，使⽤维护⽅便；d.运转平稳，振动⼩；e.单级压⽐⼤，可以在较低蒸发温度下使⽤；f.排⽓温度低，可以在⾼压⽐下⼯作；g.对湿⾏程不敏感；h.制冷量可以在10%～100%之间⽆级调节；i.操作⽅便，便于实现⾃动控制；j.体积⼩，便于实现机组化。
缺点：转⼦、机体等部件加⼯精度要求⾼，装配要求⽐较严格；油路系统及辅助设备⽐较复杂；因为转速⾼，所以噪声⽐较⼤。
2、螺杆式制冷压缩机⼯作原理双螺杆（压缩机）是由⼀对相互啮合、旋向相反的阴、阳转⼦，阴转⼦为凹型，阳转⼦为凸型。
随着转⼦按照⼀定的传动⽐旋转，转⼦基元容积由于阴阳转⼦相继侵⼊⽽发⽣改变。
侵⼊段（啮合线）向排⽓端推移，于是封闭在沟槽内的⽓体容积逐渐缩⼩，压⼒逐渐升⾼，压⼒升⾼到⼀因为⼀台压缩机的内压⽐⼀般都是固定的，⽽⼯况的变化会导致内、外压⽐不⼀致。
所以在选⽤压缩机时，应选⽤内压⽐与使⽤⼯况对应的外压⽐相同或接近的，才能获得节能。
常⽤的调节内压⽐的办法有：更换具有不同开⼝位置的滑阀（滑阀上开有径向排⽓⼝），通过改变排⽓⼝位置来改变内压⽐；采⽤具有可以调节内容积⽐的压缩机（可调内容积⽐螺杆压缩机）。
第⼆节螺杆式压缩机的结构螺杆制冷压缩机⼀般可分为机体部件、转⼦部件、滑阀部件、轴封部件和联轴器部件。
1)机体部件机体部件主要是由机体、吸⽓端座、吸⽓端盖排⽓端座、排⽓端盖及轴封压盖等零件组成。
机体：机体内设有∞字形空腔，容纳转⼦，是压缩机的⼯作汽缸。
机体内腔上部设有径向吸⽓⼝。
机体下部有⼀部分缸壁被镗掉⽤于放置滑阀。
压缩机专业词汇中英文对照大全，你不收藏算我输！
压缩机专业词汇中英文对照大全，你不收藏算我输！中英对照压缩机分类及配件词汇容积式压缩机 positive displacement compressor往复式压缩机（活塞式压缩机） reciprocating compressor 回转式压缩机 rotary compressor滑片式压缩机 sliding vane compressor单滑片回转式压缩机 single vane rotary compressor滚动转子式压缩机 rolling rotor compressor三角转子式压缩机 triangle rotor compressor多滑片回转式压缩机 multi-vane rotary compressor滑片 blade旋转活塞式压缩机 rolling piston compressor涡旋式压缩机 scroll compressor涡旋盘 scroll固定涡旋盘 stationary scroll, fixed scroll驱动涡旋盘 driven scroll, orbiting scroll斜盘式压缩机（摇盘式压缩机） swash plate compressor 斜盘 swash plate摇盘 wobble plate螺杆式压缩机 screw compressor单螺杆压缩机 single screw compressor阴转子 female rotor阳转子 male rotor主转子 main rotor闸转子 gate rotor无油压缩机 oil free compressor膜式压缩机 diaphragm compressor活塞式压缩机 reciprocating compressor单作用压缩机 single acting compressor双作用压缩机 double acting compressor双效压缩机 dual effect compressor双缸压缩机 twin cylinder compressor闭式曲轴箱压缩机 closed crankcase compressor开式曲轴箱压缩机 open crankcase compressor顺流式压缩机 uniflow compressor逆流式压缩机 return flow compressor干活塞式压缩机 dry piston compressor双级压缩机 compound compressor多级压缩机 multistage compressor差动活塞式压缩机stepped piston compound compressor, differential piston compressor串轴式压缩机 tandem compressor, dual compressor截止阀 line valve, stop valve排气截止阀 discharge line valve吸气截止阀 suction line valve部分负荷旁通口 partial duty port能量调节器 energy regulator容量控制滑阀 capacity control slide valve容量控制器 capacity control消声器 muffler联轴节 coupling曲轴箱 crankcase曲轴箱加热器 crankcase heater轴封 crankcase seal, shaft seal填料盒 stuffing box轴封填料 shaft packing机械密封 mechanical seal波纹管密封 bellows seal转动密封 rotary seal迷宫密封 labyrinth seal轴承 bearing滑动轴承 sleeve bearing偏心环 eccentric strap滚珠轴承 ball bearing滚柱轴承 roller bearing滚针轴承 needle bearing止推轴承 thrust bearing外轴承 pedestal bearing臼形轴承 footstep bearing轴承箱 bearing housing止推盘 thrust collar偏心销 eccentric pin曲轴平衡块crankshaft counterweight, crankshaft balance weight曲柄轴 crankshaft偏心轴 eccentric type crankshaft曲拐轴 crank throw type crankshaft连杆 connecting rod连杆大头 crank pin end连杆小头 piston pin end曲轴 crankshaft主轴颈 main journal曲柄 crank arm, crank shaft曲柄销 crank pin曲拐 crank throw曲拐机构 crank-toggle阀盘 valve disc阀杆 valve stem阀座 valve seat阀板 valve plate阀盖 valve cage阀罩 valve cover阀升程限制器valve lift guard阀升程 valve lift阀孔 valve port吸气口 suction inlet压缩机气阀 compressor valve吸气阀 suction valve排气阀 delivery valve圆盘阀 disc valve环片阀 ring plate valve簧片阀 reed valve舌状阀 cantilever valve条状阀 beam valve提升阀 poppet valve菌状阀 mushroom valve杯状阀 tulip valve缸径 cylinder bore余隙容积 clearance volume附加余隙（补充余隙） clearance pocket活塞排量 swept volume, piston displacement理论排量 theoretical displacement实际排量 actual displacement实际输气量 actual displacement, actual output of gas气缸工作容积 working volume of the cylinder活塞行程容积 piston displacement气缸 cylinder气缸体 cylinder block气缸壁 cylinder wall水冷套 water cooled jacket气缸盖（气缸头） cylinder head安全盖（假盖） safety head假盖 false head活塞环 piston ring气环 sealing ring刮油环 scraper ring油环 scrape ring活塞销 piston pin活塞 piston活塞行程 piston stroke吸气行程 suction stroke膨胀行程 expansion stroke压缩行程 compression stroke排气行程 discharge stroke升压压缩机 booster compressor立式压缩机 vertical compressor卧式压缩机 horizontal compressor角度式压缩机 angular type compressor对称平衡型压缩机 symmetrically balanced type compressor 压缩机参数词汇1. performance parameter 性能参数——表征压缩机主要性能的诸参数，如：气量、压力、温度、功率及噪声、振动等2. constructional parameter 结构参数——表征压缩机结构特点的诸参数，如：活塞力、行程、转速、列数、各级缸径、外形尺寸等3. inlet pressure/suction pressure 吸气压力（吸入压力）——在标准吸气位置气体的平均绝对全压力。
中英文文献翻译-螺杆式压缩机
英文原文Screw CompressorsThe direction normal to the helicoids, can be used to calculate the coordinates of the rotorhelicoids n x and n y from x and y to which the clearance is added as:dt dyD p x x n δ+=， dt dxD p y y n δ-=， ⎪⎭⎫ ⎝⎛+=dt dy y dt dx x D z n δ （2.19） where the denominator D is given as ：22222⎪⎭⎫ ⎝⎛++⎪⎭⎫ ⎝⎛+⎪⎭⎫ ⎝⎛=dt dy y dt dx x dt dy p dt dx p x D （2.20） n x and n y serve to calculate new rotor end plane coordinates, x 0n and y 0n ,with the clearances obtained for angles θ = n z /p and τ respectively. These on x and on y now serve to calculate the transverse clearance δ0 as the difference between them, as well as the original rotor coordinates o x and o y .If by any means, the rotors change their relative position, the clearance distribution at one end of the rotors may be reduced to zero on the flat side of the rotor lobes. In such a case, rotor contact will be prohibitively long on the flat side of the profile, where the dominant relative rotor motion is sliding, as shown in Fig. 2.29. This indicates that rotor seizure will almost certainly occur in that region if the rotors come into contact with each other.Fig. 2.29. Clearance distribution between the rotors: at suction, mid rotors, and discharge withpossible rotor contact at the dischargeFig. 2.30. Variable clearance distribution applied to the rotors It follows that the clearance distribution should be non-uniform to avoid hard rotor contact in rotor areas where sliding motion between the rotors is dominant.In Fig. 2.30, a reduced clearance of 65 μm is presented, which is now applied in rotor regions close to the rotor pitch circles, while in other regions it is kept at 85 μm, as was done by Edstroem, 1992. As can be seen in Fig. 2.31, the situation regarding rotor contact is now quite different. This is maintained along the rotor contact belt close to the rotor pitch circles and fully avoided at other locations. It follows that if contact occurred, it would be of a rolling character rather than a combination of rolling and sliding or even pure sliding. Such contact will not generate excessive heat and could therefore be maintained for a longer period without damaging the rotors until contact ceases or the compressor is stopped.2.6 Tools for Rotor ManufactureThis section describes the generation of formed tools for screw compressor hobbing, milling and grinding based on the envelope gearing procedure.2.6.1 Hobbing ToolsA screw compressor rotor and its formed hobbing tool are equivalent to a pair of meshing crossed helical gears with nonparallel and nonintersecting axes. Their general meshing condition is given in Appendix A. Apart from the gashes forming the cutter faces, the hob is simply a helical gear in which.Fig. 2.31. Clearance distribution between the rotors: at suction, mid of rotor and discharge with apossible rotor contact at the dischargeEach referred to as a thread, Colburne, 1987. Owing to their axes not being parallel, there is only point contact between them whereas there is line contact between the screw machine rotors. The need to satisfy the meshing equation given in Appendix A, leads to the rotor – hob meshing requirement for the given rotor transverse coordinate points 1o x and 1o y and their first derivative 0101dx dy .The hob transverse coordinate points 2o x and 2o y can then be calculated. These are sufficient to obtain the coordinate 2012012y x R +=The axial coordinate 2z , calculated directly, and 2R are hob axial plane coordinates which define the hob geometry.The transverse coordinates of the screw machine rotors, described in the previous section, are used as an example here to produce hob coordinates. he rotor unit leads 1P are 48.754mm for the main and −58.504mm for the ate rotor. Single lobe hobs are generated for unit leads 2P :6.291mm for the m ain rotor and −6.291mm for the gate rotor. The corresponding hob helix a ngles ψ are 85◦ and 95◦. The same rotor-to-hob centre distance C = 110mm a nd the shaft angle Σ = 50◦ are given for both rotors. Figure 2.32 contains a view to the hob.Reverse calculation of the hob – screw rotor transformation, also given in Appendix Apermits the determination of the transverse rotor profile coordinates which will be obtained as a result of the manufacturing process. These ay be compared with those originally specified to determine the effect ofFig. 2.32. Rotor manufacturing: hobbing tool left , right milling toolmanufacturing errors such as imperfect tool setting or tool and rotor deformation upon the final rotor profile.For the purpose of reverse transformation, the hob longitudinal plane coordinates 2R and 2z and 22dz dR should be given. The axial coordinate 2z is used to calculate 22P Z T =, which is then used to calculate the hob transverse coordinates:τcos 202R x =， τs i n 202R y = （2.21）These are then used as the given coordinates to produce a meshing criterionand the transverse plane coordinates of the “manufactured” rotors.A comparison between the original rotors and the manufactured rotors is given in Fig. 2.33 with the difference between them scaled 100 times. Two types of error are considered. The left gate rotor, is produced with 30um offset in the centre distance between the rotor and the tool, and the main rotor withFig. 2.33. Manufacturing imperfections0.2◦ of fset in the tool shaft angle Σ. Details of this particular meshing method are given by Stosic 1998.2.6.2 Milling and Grinding ToolsFormed milling and grinding tools may also be generated by placing 02=P in the general meshing equation, given in Appendix A, and then following the procedure of this section. The resulting meshing condition now reads as:[]0cot cot 1111111111=⎥⎦⎤⎢⎣⎡∂∂-∂∂+⎪⎭⎫ ⎝⎛∂∂+∂∂∑+-∑t x C t y p p t y y t x x p x C θ (2.22) However in this case, when one expects to obtain screw rotor coordinates from the tool coordinates, the singularity imposed does not permit the calculation of the tool transverse plane coordinates. The main meshing condition cannot therefore be applied. For this purpose another condition is derived for the reverse milling tool to rotor transformation from which the meshing angle τ is calculated:()0cot sin cot cos 12212222=-∑+∑++⎪⎪⎭⎫ ⎝⎛+C p dR dz C p dR dz z R ττ (2.23) Once obtained, τ will serve to calculate the rotor coordinates after the “manufacturing” process. The obtained rotor coordinates will contain all manufacturing imperfections, like mismatch of the rotor – tool centre distance, error in the rotor – tool shaft angle, axial shift of the tool or tool deformation during the process as they are input to the calculation process. A full account of this useful procedure is given by Stosic 1998.2.6.3 Quantification of Manufacturing ImperfectionsThe rotor – tool transformation is used here for milling tool profile generation. The reverse procedure is used to calculate the “manufactured” rotors. The rack generated 5-6 128mm rotors described by Stosic, 1997a are used as given profiles: x (t ) and y (t ). Then a tool – rotor transformation is used to quantify the influence of manufacturing imperfections upon the qualityof the produced rotor profile. Both, linear and angular offset were considered.Figure 2.33 presents the rotors, the main manufactured with the shaft angle offset 0.5◦and the gate with the centre distance offset 40 μm from that of the original rotors given by the dashed line on the left. On the right, the rotors are manufactured with imperfections, the main with a tool axial offset of 40 μm and the gate with a certain tool body deformation which resulted in 0.5◦offset of the relative motion angle θ. The original rotors are given by the dashed line.3Calculation of Screw Compressor Performance Screw compressor performance is governed by the interactive effects of thermodynamic and fluid flow processes and the machine geometry and thus can be calculated reliably only by their simultaneous consideration. This may be chieved by mathematical modelling in one or more dimensions. For most applications, a one dimensional model is sufficient and this is described in full. 3-D modelling is more complex and is presented here only in outline. A more detailed presentation of this will be made in a separate publication.3.1 One Dimensional Mathematical ModelThe algorithm used to describe the thermodynamic and fluid flow processes in a screw compressor is based on a mathematical model. This defines the instantaneous volume of the working chamber and its change with rotational angle or time, to which the conservation equations of energy and mass continuity are applied, together with a set of algebraic relationships used to define various phenomena related to the suction, compression and discharge of the working fluid. These form a set of simultaneous non-linear differential equations which cannot be solved in closed form.The solution of the equation set is performed numerically by means of the Runge-Kutta 4th order method, with appropriate initial and boundary conditions.The model accounts for a number of “real-life” effects, which may significantly influence the performance of a real compressor. These make it suitable for a wide range of applications and include the following:– The working fluid compressed can be any gas or liquid-gas mixture for which an equation of state and internal energy-enthalpy relation is known, i.e. any ideal or real gas or liquid-gas mixture of known properties.–The model accounts for heat transfer between the gas and the compressor rotors or its casing in a form, which though approximate, reproduces the overall effect to a good first order level of accuracy.– The model accounts for leakage of the working medium through the clearances between the two rotors and between the rotors and the stationary parts of the compressor.– The process equations and the subroutines for their solution are independent of those which define the compressor geometry. Hence, the model can be readily adapted to estimate the performance of any geometry or type of positive displacement machine.– The effects of liquid injection, including that of oil, water, or refrigerant can be accounted for during the suction, compression and discharge stages.– A set of subroutines to estimate the thermodynamic properties and changes of state of the working fluid during the entire compressor cycle of operations completes the equation set and thereby enables it to be solved.Certain assumptions had to be introduced to ensure efficient computation.These do notimpose any limitations on the model nor cause significant departures from the real processes and are as follows:– The fluid flow in the model is assumed to be quasi one-dimensional.–Kinetic energy changes of the working fluid within the working chamber are negligible compared to internal energy changes.–Gas or gas-liquid inflow to and outflow from the compressor ports is assumed to be isentropic.– Leakage flow of the fluid through the clearances is assumed to be adiabatic.3.1.1 Conservation EquationsFor Control Volume and Auxiliary RelationshipsThe working chamber of a screw machine is the space within it that contains the working fluid. This is a typical example of an open thermodynamic system in which the mass flow varies with time. This, as well as the suction and discharge plenums, can be defined by a control volume for which the differential equations of the conservation laws for energy and mass are written. These are derived in Appendix B, using Reynolds Transport Theorem.A feature of the model is the use of the non-steady flow energy equation to compute the thermodynamic and flow processes in a screw machine in terms of rotational angle or time and how these are affected by rotor profile modifications. Internal energy, rather than enthalpy, is then the derived variable. This is computationally more convenient than using enthalpy as the derived Variable since, even in the case of real fluids, it may be derived, without reference to pressure. Computation is then carried out through a series of iterative cycles until the solution converges. Pressure, which is the desired output variable, can then be derived directly from it, together with the remaining required thermodynamic properties.The following forms of the conservation equations have been employed in the model:中文翻译螺杆式压缩机几何的法线方向的螺旋，可以用来计算的坐标转子螺旋n x 和n y 的从x 和y 的间隙加入如：dt dyD p x x n δ+=， dt dxD p y y n δ-=， ⎪⎭⎫ ⎝⎛+=dt dy y dt dx x D z n δ (2.19) 其中分母D 被给定为：22222⎪⎭⎫ ⎝⎛++⎪⎭⎫ ⎝⎛+⎪⎭⎫ ⎝⎛=dt dy y dt dx x dt dy p dt dx p x D (2.20) n x ，n y 服务来计算新的转子端的平面的坐标，on x 和on y ，得到的间隙角θ =锌/ p 和τ 。
螺杆式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1 IntroductionThe 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 potentialimprovements 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 ConceptsThermodynamic 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 rotorFig. 1.1. Screw Compressor Rotorsadvances 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 iseffectively 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 housingOil injected Oil FreeFig. 1.3. Oil Injected and Oil Free CompressorsScrew 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 CompressorsScrew 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 MachineThis 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. Consequently it is widely used as a compressor in both the compressed air and refrigeration industries.1.2.2 The Oil Free MachineHere, there is no mixing of the working fluid with oil and contact between the rotors is prevented by timing gears which mesh outside the working chamber and are lubricated externally. In addition, to prevent lubricant entering the working chamber, internal seals are required on each shaft between the working chamber and the bearings. In the case of process gas compressors, double mechanical seals are used. Even with elaborate and costly systems such as these, successful internal sealing is still regarded as a problem by established process gas compressor manufacturers. It follows that such machines are considerably more expensive to manufacture than those that are oil injected.Both types require an external heat exchanger to cool the lubricating oil before it is readmitted to the compressor. The oil free machine requires an oil tank, filters and a pump to return the oil to the bearings and timing gear.The oil injected machine requires a separator to remove the oil from the high pressure discharged gas but relies on the pressure difference between suction and discharge to return the separated oil to the compressor. Theseadditional components increase the total cost of both types of machine but the add on cost is greater for the oil free compressor.1.3 Screw Machine DesignSerious efforts to develop screw machines began in the nineteen thirties, when turbomachines were relatively inefficient. At that time, Alf Lysholm, a talented Swedish engineer, required a high speed compressor, which could be coupled directly to a turbine to form a compact prime mover, in which the motion of all moving parts was purely rotational. The screw compressor appeared to him to be the most promising device for this purpose and all modern developments in these machines stem from his pioneering work. Typical screw compressor designs are presented in Figs. 1.4 and 1.5. From then until the mid nineteen sixties, the main drawback to their widespread use was the inability to manufacture rotors accurately at an acceptable cost. Two developments then accelerated their adoption. The first was the development of milling machines for thread cutting. Their use for rotor manufacture enabled these components to be made far more accurately at an acceptable cost. The second occurred in nineteen seventy three, when SRM, in Sweden, introduced the “A” profile, which reduced the internal leakage path area, known as the blow hole, by 90%. Screw compressors could then be built with efficiencies approximately equal to those of reciprocating machines and, in their oil flooded form, could operate efficiently with stage pressure ratios of up to 8:1. This was unattainable with reciprocating machines. The use of screw compressors, especially of the oil flooded type, then proliferated.Fig. 1.4. Screw compressor mechanical partsFig. 1.5. Cross section of a screw compressor with gear boxTo perform effectively, screw compressor rotors must meet the meshing requirements of gears while maintaining a seal along their length to minimise leakage at any position on the band of rotor contact. It follows that the compressor efficiency depends on both the rotor profile and the clearances between the rotors and between the rotors and the compressor housing.Screw compressor rotors are usually manufactured on pecialized machines by the use of formed milling or grinding tools. Machining accuracy achievable today is high and tolerances in rotor manufacture are of the order of 5 μm around the rotor lobes. 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 betweenthe rotors canbe as small as 12 μm.中文译文1 引言螺杆式压缩机是一种最常见的用来压缩气体的机器。
螺杆式制冷压缩机使用说明
夸夸我的爸爸的范文我得好好夸一夸我那超级棒的爸爸！我爸呀，就像一个全能超人。
先说他那双手，简直是神奇的“魔法手”。
家里什么东西坏了，只要到了他手里，就像被施了魔法一样，立马恢复正常。
记得有一次，家里的老台灯不亮了，我正打算把它扔了，爸爸却一把拿过来。
他打开台灯底座，这儿瞅瞅，那儿看看，拿着工具捣鼓了几下，然后“啪”的一声，台灯就亮了起来，那光芒好像都比以前更亮堂了呢。
我当时就觉得我爸比那些维修师傅都厉害。
他还是个“知识宝库”。
不管我在学习上遇到什么难题，只要去问他，就没有解决不了的。
有次我被一道数学题难住了，那题目就像一团乱麻，我绞尽脑汁也没个头绪。
我把题目拿给爸爸看，他眼睛一扫，就开始给我讲解。
他讲得可有趣了，把那些枯燥的数字和公式都变成了一个个有趣的故事。
原本让我头疼的数学题，在他的讲解下就像解开了谜题一样简单。
而且我爸知道的可不止是数学，历史、地理、科学，他都能说上一通，感觉他脑袋里装着整个世界的知识呢。
我爸的厨艺也是一绝。
他做的菜那叫一个香，每次他一进厨房，我就像小馋猫一样在厨房门口等着。
他最拿手的红烧肉，那色泽红亮，一口咬下去，肥而不腻，瘦肉部分鲜嫩可口，那味道在嘴里散开的时候，就像一场味蕾的狂欢派对。
还有他炒的青菜，绿油油的，又脆又嫩，吃起来有一股清甜的味道。
他总是说，做饭就像做人，得用心，我看他就是把满满的爱都放进了饭菜里。
爸爸还是个很幽默的人呢。
有一次我考试没考好，心情特别低落，感觉整个世界都变得灰暗了。
爸爸看到我这个样子，就模仿起小丑来逗我。
他把衣服拉起来当成裙子，扭着屁股，脸上还做出各种滑稽的表情。
我“扑哧”一声就笑了出来，他就趁机鼓励我，说一次失败不算什么，就像他做饭偶尔也会盐放多了一样，下次注意就好啦。
他的幽默就像阳光一样，能把我心里的乌云都驱散。
在我心中，爸爸就是一座高大的山，给我依靠；又是一盏明亮的灯，为我照亮前行的路。
我爱我的爸爸，他就是这个世界上最酷、最棒的爸爸！。
空压机相关专业术语中英文对照表
空压机相关专业术语中英文对照表序号中文英文1 空压机Air compressor2 压缩机Compressor3 活塞式空压机Piston compressor4 螺杆式空压机Screw compressor5 滑片式空压机Vane compressor6 离心式空压机Centrifugal compressor7 涡旋式空压机Scroll compressor8 无油空压机Oil-free compressor9 移动式空压机Portable compressor10 喷油螺杆空压机Oil injected screw compressor11 储气罐Air Tank12 压力表Pressure gauge13 油管Oil Tube14 风冷却器Air Cooler15 油冷却器Oil Cooler16 水冷却器Water Cooler17 电机Motor18 机头Air end19 进气阀Inlet valve20 温控阀Thermostatic valve21 压力传感器Pressure sensor22 吸附式干燥机Adsorption drier23 冷冻式干燥机Freeze drier24 吸附剂Adsorbent25 冷煤Refrigerant26 过滤器Filter27 管路过滤器Pipeline filter28 高效过滤器High efficiency air filter29 精密过滤器Fine filter30 滤芯Filter element31 除尘滤芯Dust removal filter32 除油滤芯Oil removal filter33 活性碳滤芯Activated carbon filter34 灭菌滤芯Sterilization filter35 露点Dew point36 压力露点Pressure dew point37 自动排水器Automatic drain valve38 电子排水器Electronic drain valve39 智能排水器Intelligent drain valve40 机油Oil41 油滤Oil filter42 空滤Air filter43 空滤总成Air filter housing44 油分Air/Oil separator45 外置油分Spin-on Air/Oil separator46 法兰Flange47 外径Outside diameter48 内径Inside diameter49 高度Height50 跑油Run oil51 压差Differential pressure52 干燥机Air drier53 旋入式过滤器Spin-on Filter54 维修包service kit55 售后市场after market56 压缩空气compressed air57 软管Tube58 汽水分离器steam separator59 油水分离器oil-water separator60 安全芯safe filter61 联轴器flex62 卸荷阀unloading valve63 最小压力阀Min pressure valve64 旁通阀by pass valve65 单向阀check valve66 替代品replacement67 检测阀check valve68 垫圈 washer69 库存store70 止动板Lock plate71 密封垫圈Seal washer。
压缩机-中英对照
4.7压缩机机组安装Installation of compressor4.7.1就位前应做好下列工作：the following work should be done beforecompressor emplacement——将汽轮机的凝汽器初步就位，其安装标高宜比设计标高低20mm～25mm；Place the turbine condenser in position, let the installation elevation20-25mm lower than the design data.——预先安装好机组下部在机组就位后无法安装的管道及管件等；Pre-install the pipes and fittings which cannot be installed after emplacement.——仔细清除底座底面的油污及铁锈，并涂刷一层水泥浆；Carefully clean the underside of the base plate, remove oil and rust, and one layer of cement mortarwill be coated on the plate's underside.4.7.2 Installation and alignment of the compressor shall generally be carried out inorder as follows:(1). Carry out treatments, such as chipping and fitting, on foundation furface.(2). Set line.(3). Place the frame of crankshaft and cross-guides on the liner provisionally and setanchor bolts.(4). Align the frame provisionally.(5). Mount crankshaft and confirm the deflection of crankshaft arms.(6). Pour mortar into anchor boxes.(7). After the mortar solidification, align the frame finally by tightening the anchorbolts.(8). Align compressor shaft with motor shaft.(9). Reconfirm the deflection of crankshaft arms.(10). Mount cross-heads and connection-rod.(11). Set other parts and accessories.Explanation:① Leveling of the frame in the crankshaft direction shall be carried out on the upper finished surfaces of both sides of the frame paralled to crankshaft, and leveling in the cross-guide direction shall be carried out the sliding surfaces of cross-guides.②Main bearings and crankshaft shall be mounted on the frame in accordance with the following procedures:a. Adjust main bearings to keep deflection of crankshaft arms within the specified toloerance.b. Confirm that the clearance between main bearing metal and crankshaft journal is as specified by the munufacturer, and contacting traces on the sliding surface of main bearings metal is uniform.c.4.7.3 Insallation and alignment requirements(1)4.7.2各缸体整体就位，穿好地脚螺栓，地脚螺栓的光杆部分应无油污，螺纹部分应涂抹油脂。
螺杆压缩机外文文献翻译、中英文翻译、外文翻译
螺杆压缩机外文文献翻译、中英文翻译、外文翻译英文原文Screw CompressorsN. Stosic I. Smith A. KovacevicScrew CompressorsMathematical Modellingand Performance CalculationWith 99 FiguresABCProf. Nikola StosicProf. Ian K. SmithDr. Ahmed KovacevicCity UniversitySchool of Engineering and Mathematical SciencesNorthampton SquareLondonEC1V 0HBU.K.e-mail:n.stosic@/doc/d6433edf534de518964bcf 84b9d528ea81c72f87.htmli.k.smith@/doc/d6433edf534de51896 4bcf84b9d528ea81c72f87.htmla.kovacevic@/doc/d6433edf534de51 8964bcf84b9d528ea81c72f87.htmlLibrary of Congress Control Number: 2004117305ISBN-10 3-540-24275-9 Springer Berlin Heidelberg New York ISBN-13 978-3-540-24275-8 Springer Berlin Heidelberg New YorkThis work is subject to copyright. All rights are reserved, whether the whole or part of the material is concerned, specifically the rights of translation, reprinting, reuse of illustrations, recitation, broadcasting, reproduction on microfilm or in any other way, and storage in data banks. Duplication of this publication or parts thereof is permitted only under the provisions of the German Copyright Law of September 9, 1965, in its current version, and permission for use must always be obtained from Springer. Violations are liable for prosecution under the GermanCopyright Law.Springer is a part of Springer Science+Business Media/doc/d6433edf534de518964bcf84b9d 528ea81c72f87.html_c Springer-Verlag Berlin Heidelberg 2005Printed in The NetherlandsThe use of general descriptive names, registered names, trademarks, etc. in this publication does not imply,even in the absence of a specific statement, that such names are exempt from the relevant protective laws and regulations and therefore free for general use.Typesetting: by the authors and TechBooks using a Springer LATEX macro packageCover design: medio, BerlinPrinted on acid-free paper SPIN: 11306856 62/3141/jl 5 4 3 2 1 0PrefaceAlthough the principles of operation of helical screw machines, as compressors or expanders, have been well known for more than 100 years, it is only during the past 30 years thatthese machines have become widely used. The main reasons for the long period before they were adopted were their relatively poor efficiency and the high cost of manufacturing their rotors. Two main developments led to a solution to these difficulties. The first of these was the introduction of the asymmetric rotor profile in 1973. This reduced the blowhole area, which was the main source of internal leakage by approximately 90%, and thereby raised the thermodynamic efficiency of these machines, to roughly the same level as that of traditional reciprocating compressors. The second was the introduction of precise thread milling machine tools at approximately the same time. This made it possible to manufacture items of complex shape, such as the rotors, both accurately and cheaply.From then on, as a result of their ever improving efficiencies, high reliability and compact form, screw compressors have taken an increasing share of the compressor market, especially in the fields of compressed air production, and refrigeration and air conditioning, and today, a substantial proportion of compressors manufactured for industry are of this type.Despite, the now wide usage of screw compressors and the publication of many scientific papers on their development, only a handful of textbooks have been published to date, which give a rigorous exposition of the principles of their operation and none of these are in English.The publication of this volume coincides with the tenth anniversary of the establishment of the Centre for Positive Displacement Compressor Technology at City University, London, where much, if not all, of the material it contains was developed. Its aim is to give an up to date summary of the state of the art. Its availability in a single volume should then help engineers inindustry to replace design procedures based on the simple assumptions of the compression of a fixed mass of ideal gas, by more up to date methods. These are based on computer models, which simulate real compression and expansion processes more reliably, by allowing for leakage, inlet and outlet flow and other losses, VI Preface and the assumption of real fluid properties in the working process. Also, methods are given for developing rotor profiles, based on the mathematical theory of gearing, rather than empirical curve fitting. In addition, some description is included of procedures for the three dimensional modelling of heat and fluid flow through these machines and how interaction between the rotors and the casing produces performance changes, which hitherto could not be calculated. It is shown that only a relatively small number of input parameters is required to describe both the geometry and performance of screw compressors. This makes it easy to control the design process so that modifications can be cross referenced through design software programs, thus saving both computer resources and design time, when compared with traditional design procedures.All the analytical procedures described, have been tried and proven on machines currently in industrial production and have led to improvements in performance and reductions in size and cost, which were hardly considered possible ten years ago. Moreover, in all cases where these were applied, the improved accuracy of the analytical models has led to close agreement between predicted and measured performance which greatly reduced development time and cost. Additionally, the better understanding of the principles of operation brought about by such studies has led to an extension of the areas of application of screw compressors and expanders.It is hoped that this work will stimulate further interest in an area, where, though much progress has been made, significant advances are still possible.London, Nikola StosicFebruary 2005 Ian SmithAhmed KovacevicNotationA Area of passage cross section, oil droplet total surfacea Speed of soundC Rotor centre distance, specific heat capacity, turbulence model constantsd Oil droplet Sauter mean diametere Internal energyf Body forceh Specific enthalpy h = h(θ), convective heat transfer coefficient betweenoil and gasi Unit vectorI Unit tensork Conductivity, kinetic energy of turbulence, time constant m Massm˙ Inlet or exit mass flow rate m˙ = m˙ (θ)p Rotor lead, pressure in the working chamber p = p(θ)P Production of kinetic energy of turbulenceq Source term˙Q Heat transfer rate between the fluid and the compressor surroundin gs˙Q= ˙Q(θ)r Rotor radiuss Distance between the pole and rotor contact points, control volume surfacet TimeT Torque, Temperatureu Displacement of solidU Internal energyW Work outputv Velocityw Fluid velocityV Local volume of the compressor working chamber V = V (θ)˙VVolume flowVIII Notationx Rotor coordinate, dryness fraction, spatial coordinatey Rotor coordinatez Axial coordinateGreek Lettersα Temperature dilatation coefficientΓ Diffusion coefficientε Dissipation of kinetic energy of turbulenceηi Adiabatic efficiencyηt Isothermal efficiencyηv Volumetric efficiencySpecific variableφ Variableλ Lame coefficientμ Viscosityρ Densityσ Prand tl numberθ Rotor angle of rotationζ Compound, local and point resistance coefficientω Angular speed of rotationPrefixesd differentialΔ IncrementSubscriptseff Effectiveg Gasin Inflowf Saturated liquidg Saturated vapourind Indicatorl Leakageoil Oilout Outflowp Previous step in iterative calculations SolidT Turbulentw pitch circle1 main rotor, upstream condition2 gate rotor, downstream conditionContents1Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ………………………. . . . . . . . . . . . . . . 1 1.1 Basic Concepts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .. .. . . . . . . . . 4 1.2 Types of Screw Compressors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .. . . . . ….. . . . . . . .7 1.2.1 The Oil Injected Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . …... . .71.2.2 The Oil Free Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .. . . . . . . . . . . ….... .7 1.3 Screw Machine Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .. . . . . . . . . . . . . . . .8 1.4 Screw Compressor Practice . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .. . . . . . . . . . . . . . . . .101.5RecentDevelopments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12 1.5.1RotorProfiles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .. . . . . 13 1.5.2CompressorDesign . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17 2ScrewCompressorGeometry. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 192.1 The Envelope Method as a Basis for the Profiling of Screw CompressorRotors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ………………………….. . . . . ….. . . . . . . . 19 2.2 Screw Compressor Rotor Profile s . . . . . . . . . . . . . . . . . . . . …. . . . . . . . . . . . . . . . . . . ….. . . 20 2.3 Rotor ProfileCalculation . . . . . . . . . . . . . . . . . . . . . . . . . . . …………………………. . . . . .23 2.4 Review of Most Popular Rotor Profiles . . . . . . . . . . . . . . . ………………………….. . . . . . 23 2.4.1 Demonstrator Rotor Profile (“N” Rotor Generated) . . ………………………………….. . 24 2.4.2 SKBK Profile . . . . . . . . . . . . . . . . . . . . . . . . . . . ……………………………... . . . . . . . . .26 2.4.3 Fu Sheng Profile . . . . . . . . . . . . . . . . . . . . . . . . . ………………………………. . . . . . . . .27 2.4.4 “Hyper”Profile . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ………………………………. . .27 2.4.5 “Sigma” Profile . . . . . . . . . . . . . . . . . . . . . . .. . . . . . ………………………………. . . . . .28 2.4.6 “Cyclon” Profile . . . . . . . . . . . . . . . . . . . . . . . . . . . . ………………………………. . . . . .28 2.4.7 Symmetric Prof ile . . . . . . . . . . . . . . . . . . . . . . . . . . . ……………………………… . . . . .29 2.4.8 SRM “A” Profile . . . . . . . . . . . . . . . . . . . . . . . . . . ……………………………… . . . . . . .30 2.4.9 SRM “D” Profile . . . . . . . . . . . . . . . . . . . . . . . . . . . ……………………………… . . . . . .31 2.4.10 SRM “G” Profile . . . . . . . . . . . . . . . .. . . . . . . . …………………………….. . . . . . . . . .32 2.4.11 City “N” Rack Generated Rotor Profile . . . . . . . . . . . ………………………………… . . 32 2.4.12 Characteristics of “N” Profile . . . . . . . . . . . . . . . . . . . ………………………………. . . . 34 2.4.13 Blower Rot or Profile . . . . . . . . . . . . . . . . . . . . …………………………….. . . . . . . . . . . 39 X Contents2.5 Identification of Rotor Positionin Compressor Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . …………………………….. . . . . . . .40 2.6 Tools for Rotor Manufacture . . . . . . . . . . . . . . . . . . . . . . …………………………. . . . . . . .45 2.6.1 Hobbing Tools . . . . . . . . . . . . . . . . . . . . . . . . . . ………….…..………………. . . . . . . . . .45 2.6.2 Milling and Grinding Tools . . . . . . . . . . . . . . . . . . . ……………………………….... . . . . 482.6.3 Quantification of ManufacturingImperfections . . . . . ……………………………….... . . 483 Calculation of Screw Compressor Performance . . . . . . . . . . ………………………………. . . 49 3.1 One Dimensional Mathematical Model . . . . . . . . . . . . . . …………………………... . . . . . .49 3.1.1 Conservation Equationsfor Control Volume and Auxiliary Relationships . . . . ............................................... . . 50 3.1.2 Suction and Discharge Ports . . . . . . . . . . . . . . . . . . . ....................................... . . . . 53 3.1.3 Gas Leakages . . . . . . . . . . . . . . . . . . . . . . . . . . .................................... . . . . . . . . . .54 3.1.4 Oil or Liquid Injection . . . . . . . . . . . ...................................... . . . . . . . . . . . . . . . . . 55 3.1.5 Computation of Fluid Properties . . . . . . . . ........................................ . . . . . . . . . . . 57 3.1.6 Solution Procedure for Compressor Thermodynamics . (58)3.2 Compressor Integral Parameters . . . . . . . . . . . . . . . . . . . ………………………….. . . . . . . . 59 3.3 Pressure Forces Actingon Screw Compressor Rotors . . . . . . . . . . . . . . . . . . . . . . ................................... . . . . . . . 61 3.3.1 Calculation of Pressure Radial Forces and Torque . . . . .. (61)3.3.2 Rotor Bending Deflections . . . . . . . . . . . . . . . . . . . . . ……………………………….. . . . 64 3.4 Optimisation of the Screw Compressor Rotor Profile,Compressor Design and Operating Parameters . . . . . . . . . . ……………………………….. . . . 65 3.4.1 OptimisationRationale . . . . . . . . . . . . . . . . . . . . . . . . ……………………………….. . . . 65 3.4.2 Minimisation Method Usedin Screw CompressorOptimisation . . . . . . . . . . . ……………………………………… . . . . . . 67 3.5 Three Dimensional CFD and Structure Analysisof a Screw Compressor . . . . . . . . . . . . . . . . . . . . . . . . . …………………………….. . . . . . . . .71 4 Principles of Screw Compressor Design. . . . . . . . . . . …………………………… . . . . . . . . 77 4.1 Clearance Management. . . . . . . . . . . . . . . . . . . . . . . . ………….….………… . . . . . . . . . .78 4.1.1 Load Sustainability . . . . . . . . . . . . . . . . . . . . . . . . . . . . ………….………………….. . . .79 4.1.2 Compressor Size and Scale . . . . . . . . . . . . . . ………………………………. . . . . . . . . . . 80 4.1.3 RotorConfiguration . . . . . . . . . . . . . . . . . . . . . . . ……………………………... . . . . . . .82 4.2 Calculation Example:5-6-128mm Oil-Flooded Air Compressor . . . . . . . . . . . . . . . ……………………………... . . . 824.2.1 Experimental Verification of the Model . . . . . . . . . . . ………………………………. . . . 845 Examples of Modern Screw Compressor Designs . . . . . . . ……………………………… . . . 89 5.1 Design of an Oil-Free Screw CompressorBased on 3-5 “N” Rotors . . . . . . . . . . . . . . . . . . . . . . . . . . ……………………………. . . . . . . 90 5.2 The Design of Familyof Oil-Flooded Screw Compressors Basedon 4-5 “N” Rotors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . …………………………… . . . . . . .93 Contents XI.5.3 Design of Replacement Rotorsfor Oil-FloodedCompressors . . . . . . . . . . . . . . . . . . . . . . . . . . . ................................. . .96 5.4 Design of Refrigeration Compressors . . . . . . . . . . . . . . . .............................. . . . . . . 100 5.4.1 Optimisation of Screw Compressors for Refrigeration . . . (102)5.4.2 Use of New Rotor Profiles . . . . . . . . . . . . . . . . . . . . . . . . . . (103)5.4.3 Rotor Retrofits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ……………………………. . . 103 5.4.4 Motor Cooling Through the Superfeed Port in Semihermetic Compressors . . . . . . . . . . . . . . . . . . . …………………………………… . . . 103 5.4.5 Multirotor Screw Compressors . . . . . . . . . . . . . . . . . …………………………….... . . . . 104 5.5 Multifunctional Screw Machines . . . . . . . . . . . . . . . . . . ……………………….. . . . . . . . . 108 5.5.1 Simultaneous Compression and Expansionon One Pair of Rotors . . . . . . . . . . . . . . . . . . . . . . . . . . ............................................ . 108 5.5.2 Design Characteristics of Multifunctional Screw Rotors .. (109)5.5.3 Balancing Forces on Compressor-Expander Rotors . …………………..……………. . . 1105.5.4 Examples of Multifunctional Screw Machines . . . . . . . . (111)6Conclusions. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . …………………… . . . . . . . . . 117A Envelope Method of Gearin g . . . . . . . . . . . . . . . . . . . . . . . . ………………………… . . . . . 119B Reynolds TransportTheorem. . . . . . . . . . . . . . . . . . . . . . . …………………………. . . . . . . 123C Estimation of Working Fluid Propertie s . . . . . . . . . . . . . . . …………………………….. . . . 127 Re ferences. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ………………… . . . . . . . . . . 133中文译文螺杆压缩机N.斯托西奇史密斯先生A科瓦切维奇螺杆压缩机计算的数学模型和性能尼古拉教授斯托西奇教授伊恩史密斯博士艾哈迈德科瓦切维奇工程科学和数学北安普敦广场伦敦城市大学英国电子邮件：n.stosic@/doc/d6433edf534de518964bcf 84b9d528ea81c72f87.htmli.k.smith@/doc/d6433edf534de51896 4bcf84b9d528ea81c72f87.htmla.kovacevic@/doc/d6433edf534de51 8964bcf84b9d528ea81c72f87.html国会图书馆控制号：2004117305isbn-10 3-540-24275-9 纽约施普林格柏林海德堡isbn-13 978-3-540-24275-8 纽约施普林格柏林海德堡这项工作是受版权保护,我们保留所有权利。
压缩机英语术语
压缩机英语术语Type of compressor 压缩机类型1. positive displacement compressor 容积式(正位移)压缩机1) reciprocating compressor往复式压缩机2) piston compressor活塞式压缩机3) plunger compressor柱塞式压缩机4) diaphragm compressor 隔膜式压缩机5) rotary compressor回转式压缩机6) double screw compressor双螺杆压缩机7) single screw compressor单螺杆压缩机8) rolling piston compressor滚动活塞式压缩机9) sliding vane compressor 滑片式压缩机10) liquid ring compressor 液环式压缩机11) triangle rotor compressor 三角转子压缩机12) scroll compressor 涡旋式压缩机13) Roots blower 罗茨鼓风机2.dynamic compressor 动力式压缩机1) turbo compressor 透平压缩机2)centrifugal compressor 离心式压缩机3) axial flow compressor 轴流式压缩机Reciprocating compressor 往复式压缩机基本概念1) compressed medium 压缩介质(介质、工质)——被压缩的工质(气体) 2) stage 级——完成压缩循环的基本单元3) 段——在一台工艺流程用压缩机中，相邻各级的气量和组分相同时，称为段 4) line (row) 列——在同一气缸轴线上的单个气缸或串联气缸，结构上组成一列 5) bare compressor 主机——压缩机的机体部分和压缩部分的总称6) driver 驱动机(原动机)——驱动压缩机的动力机械或装置7) auxiliary equipments 附属设备(辅机)——除主机和驱动机外，其余设备的总称 8) frame and moving parts 机体部分(基础部分)——压缩机的机身或曲轴箱和运动部件等的总称9) compressing components 压缩部分(气缸部分)——压缩机的气缸、活塞、气阀和填料函等部件的总称10) variable-type 变型——对原型机的结构作局部变更，在其型号中标有“结构差异”者，称为原型机的变型11) variant 派生——原型机的机体部分基本不变或其型号中的“结构”不变，而“特征”和(或)性能参数及特性或介质改变者，称为原型机的派生12) inter-cooling 中间冷却(级间冷却)——导走级间压缩介质的热量 13) after coling 后冷却——导走完成压缩后的压缩介质的热量14) liquid injection cooling 喷液冷却——向压缩介质中喷液以降低介质温度 15 ideal compressor 理想压缩机——无余隙容积、无泄漏、亦无压力损失现象的压缩机 types 分类1. crankshaft piston compressor 曲轴活塞压缩机(活塞压缩机)——具有曲轴旋转运动的压缩机2. reciprocating compressor without crankshaft 无曲轴压缩机(无轴压缩机)——没有曲轴或轴的旋转运动的压缩机3. shaft piston compressor 轴活塞压缩机(斜盘压缩机)——活塞轴线平行于动力输入轴轴线，且均布于其周围的压缩机4. diaphragm compressor 隔膜压缩机(膜片压缩机)——机械直接或液压驱动膜片，完成压缩循环的压缩机5. gas compressor 气体压缩机(压气机)——压缩介质除空气外的其他气体，也可包括空气的压缩机之统称6. air compressor 空气压缩机(压风机)——压缩机介质为空气的压缩机7. general purpose compressor/all purpose compressor/compressor for pressed air/pressed–air compressor 动力用压缩机——为气动机械和气动工具提供动力气源的压缩机8. process compressor 工艺流程用压缩机(流程压缩机)——石油、化工等工艺流程用的压缩机9. compressor for ship purpose 船用压缩机——专用于舰船的压缩机10. filling compressor 充瓶用压缩机(充瓶压缩机)——用于压缩介质装瓶的压缩机 11. multi–purpose compressor/multi–service compressor 联合压缩机——同一台压缩机中，各气缸分别压缩多种工质且非前后级关系的压缩机12. combined compressor/combination compressor复合压缩机(串联压缩机)——同一压缩机中，分别采用不同类型的压缩机，形成前后级关系，达到提高介质压力的压缩机 13. motor compressor/gas–engine compressor 摩托压缩机14. integral compressor/motor compressor 整体压缩机(摩托压缩机)——与往复式原动机部分地共用一个运动机构的压缩机。
压缩 机 中英词汇
Foundations 基础
Guide vane 导叶
Impeller 叶轮
Intercooling 中间冷却器
Coupling 联轴器
Diaphragm 隔板
Diffuser 扩压器
Discharge nozzle 排气接管
Discharge volute 排气蜗壳
Elastohydrodynamic 弹性流体动压
Elliptical 椭圆形
Film thickness 膜厚
Flow 流量
Fluid film 流体膜
Gas 气体
Hydrostatic 流体静压
Journal 轴颈
Liner 衬套
Materials 材料
Stress distribution 应力分布
Tie rods 拉杆
Damped systems 阻尼系统
Dewhirl vanes 破涡片
Inlet nozzle 进气（接）管
Inlet volute 进气蜗壳
Multistage 多级
Off design operation 非设计工况操作
Oil system 密封油系统
Absolute pressure 绝对压力
Absolute temperature 绝对温度
Adiabatic compression 绝热压缩过程
Air padding (压缩)空气（填充）输送
Aluminalkyle 烷基铝
Performance 性能
Slope 倾斜
Splitter vane 分流叶片
机械专业相关英文词汇集锦—中英文对照
机械专业相关英文词汇集锦—中英文对照整理：小笨猪（sqchtolzy）阿基米德蜗杆 Archimedes worm 安全系数 safety factor; factor of safety安全载荷 safe load 凹面、凹度 concavity扳手 wrench 板簧 flat leaf spring半圆键 woodruff key 变形 deformation摆杆 oscillating bar 摆动从动件 oscillating follower摆动从动件凸轮机构 cam with oscillating follower 摆动导杆机构 oscillating guide-bar mechanism摆线齿轮 cycloidal gear 摆线齿形 cycloidal tooth profile摆线运动规律 cycloidal motion 摆线针轮 cycloidal-pin wheel包角 angle of contact 保持架 cage背对背安装 back-to-back arrangement 背锥 back cone ；normal cone背锥角 back angle 背锥距 back cone distance比例尺 scale 比热容 specific heat capacity闭式链 closed kinematic chain 闭链机构 closed chain mechanism臂部 arm 变频器 frequency converters变频调速 frequency control of motor speed 变速 speed change变速齿轮 change gear ; change wheel 变位齿轮 modified gear变位系数 modification coefficient 标准齿轮 standard gear标准直齿轮 standard spur gear 表面质量系数 superficial mass factor表面传热系数 surface coefficient of heat transfer 表面粗糙度 surface roughness并联式组合 combination in parallel 并联机构 parallel mechanism并联组合机构 parallel combined mechanism 并行工程 concurrent engineering并行设计 concurred design, CD 不平衡相位 phase angle of unbalance不平衡 imbalance (or unbalance) 不平衡量 amount of unbalance不完全齿轮机构 intermittent gearing 波发生器 wave generator波数 number of waves 补偿 compensation参数化设计 parameterization design, PD 残余应力 residual stress操纵及控制装置 operation control device 槽轮 Geneva wheel槽轮机构 Geneva mechanism ；Maltese cross 槽数 Geneva numerate槽凸轮 groove cam 侧隙 backlash差动轮系 differential gear train 差动螺旋机构 differential screw mechanism差速器 differential 常用机构 conventional mechanism; mechanism in common use 车床 lathe 承载量系数 bearing capacity factor承载能力 bearing capacity 成对安装 paired mounting尺寸系列 dimension series 齿槽 tooth space齿槽宽 spacewidth 齿侧间隙 backlash齿顶高 addendum 齿顶圆 addendum circle齿根高 dedendum 齿根圆 dedendum circle齿厚 tooth thickness 齿距 circular pitch齿宽 face width 齿廓 tooth profile齿廓曲线 tooth curve 齿轮 gear齿轮变速箱 speed-changing gear boxes 齿轮齿条机构 pinion and rack齿轮插刀 pinion cutter; pinion-shaped shaper cutter 齿轮滚刀 hob ,hobbing cutter齿轮机构 gear 齿轮轮坯 blank齿轮传动系 pinion unit 齿轮联轴器 gear coupling齿条传动 rack gear 齿数 tooth number齿数比 gear ratio 齿条 rack齿条插刀 rack cutter; rack-shaped shaper cutter 齿形链、无声链 silent chain齿形系数 form factor 齿式棘轮机构 tooth ratchet mechanism插齿机 gear shaper 重合点 coincident points重合度 contact ratio 冲床 punch传动比 transmission ratio, speed ratio 传动装置 gearing; transmission gear传动系统 driven system 传动角 transmission angle传动轴 transmission shaft 串联式组合 combination in series串联式组合机构 series combined mechanism 串级调速 cascade speed control创新 innovation ; creation 创新设计 creation design垂直载荷、法向载荷 normal load 唇形橡胶密封 lip rubber seal磁流体轴承 magnetic fluid bearing 从动带轮 driven pulley从动件 driven link, follower 从动件平底宽度 width of flat-face从动件停歇 follower dwell 从动件运动规律 follower motion从动轮 driven gear 粗线 bold line粗牙螺纹 coarse thread 大齿轮 gear wheel打包机 packer 打滑 slipping带传动 belt driving 带轮 belt pulley带式制动器 band brake 单列轴承 single row bearing单向推力轴承 single-direction thrust bearing 单万向联轴节 single universal joint单位矢量 unit vector 当量齿轮 equivalent spur gear; virtual gear当量齿数 equivalent teeth number; virtual number of teeth当量摩擦系数 equivalent coefficient of friction当量载荷 equivalent load 刀具 cutter导数 derivative 倒角 chamfer导热性 conduction of heat 导程 lead导程角 lead angle 等加等减速运动规律 parabolic motion; constant acceleration and deceleration motion等速运动规律 uniform motion; constant velocity motion 等径凸轮 conjugate yoke radial cam等宽凸轮 constant-breadth cam 等效构件 equivalent link等效力 equivalent force 等效力矩 equivalent moment of force等效量 equivalent 等效质量 equivalent mass等效转动惯量 equivalent moment of inertia 等效动力学模型 dynamically equivalent model底座 chassis 低副 lower pair点划线 chain dotted line （疲劳）点蚀 pitting垫圈 gasket 垫片密封 gasket seal碟形弹簧 belleville spring 动力学 dynamics顶隙 bottom clearance 定轴轮系 ordinary gear train; gear train with fixed axes 动密封 kinematical seal 动能 dynamic energy动力粘度 dynamic viscosity 动力润滑 dynamic lubrication动平衡 dynamic balance 动平衡机 dynamic balancing machine动态特性 dynamic characteristics 动态分析设计 dynamic analysis design动压力 dynamic reaction 动载荷 dynamic load端面 transverse plane 端面参数 transverse parameters端面齿距 transverse circular pitch 端面齿廓 transverse tooth profile端面重合度 transverse contact ratio 端面模数 transverse module端面压力角 transverse pressure angle 锻造 forge对称循环应力 symmetry circulating stress 对心滚子从动件 radial (or in-line ) roller follower 对心直动从动件 radial (or in-line ) translating follower对心移动从动件 radial reciprocating follower对心曲柄滑块机构 in-line slider-crank (or crank-slider) mechanism多列轴承 multi-row bearing多楔带 poly V-belt 多项式运动规律 polynomial motion多质量转子 rotor with several masses 惰轮 idle gear额定寿命 rating life 额定载荷 load ratingII 级杆组 dyad 发生线 generating line发生面 generating plane 法面 normal plane法面参数 normal parameters 法面齿距 normal circular pitch法面模数 normal module 法面压力角 normal pressure angle法向齿距 normal pitch 法向齿廓 normal tooth profile法向直廓蜗杆 straight sided normal worm 法向力 normal force反馈式组合 feedback combining 反向运动学 inverse ( or backward) kinematics反转法 kinematic inversion 反正切 Arctan范成法 generating cutting 仿形法 form cutting方案设计、概念设计 concept design, CD 防振装置 shockproof device飞轮 flywheel 飞轮矩 moment of flywheel非标准齿轮 nonstandard gear 非接触式密封 non-contact seal非周期性速度波动 aperiodic speed fluctuation 非圆齿轮 non-circular gear粉末合金 powder metallurgy 分度线 reference line; standard pitch line分度圆 reference circle; standard (cutting) pitch circle分度圆柱导程角 lead angle at reference cylinder分度圆柱螺旋角 helix angle at reference cylinder 分母 denominator分子 numerator 分度圆锥 reference cone; standard pitch cone分析法 analytical method 封闭差动轮系 planetary differential复合铰链 compound hinge 复合式组合 compound combining复合轮系 compound (or combined) gear train 复合平带 compound flat belt复合应力 combined stress 复式螺旋机构 Compound screw mechanism复杂机构 complex mechanism 杆组 Assur group干涉 interference 刚度系数 stiffness coefficient刚轮 rigid circular spline 钢丝软轴 wire soft shaft刚体导引机构 body guidance mechanism 刚性冲击 rigid impulse (shock)刚性转子 rigid rotor 刚性轴承 rigid bearing刚性联轴器 rigid coupling 高度系列 height series高速带 high speed belt 高副 higher pair格拉晓夫定理 Grashoff`s law 根切 undercutting公称直径 nominal diameter 高度系列 height series功 work 工况系数 application factor工艺设计 technological design 工作循环图 working cycle diagram工作机构 operation mechanism 工作载荷 external loads工作空间 working space 工作应力 working stress工作阻力 effective resistance 工作阻力矩 effective resistance moment公法线 common normal line 公共约束 general constraint公制齿轮 metric gears 功率 power功能分析设计 function analyses design 共轭齿廓 conjugate profiles共轭凸轮 conjugate cam 构件 link鼓风机 blower 固定构件 fixed link; frame固体润滑剂 solid lubricant 关节型操作器 jointed manipulator惯性力 inertia force 惯性力矩 moment of inertia ,shaking moment 惯性力平衡 balance of shaking force 惯性力完全平衡 full balance of shaking force 惯性力部分平衡 partial balance of shaking force 惯性主矩 resultant moment of inertia惯性主失 resultant vector of inertia 冠轮 crown gear广义机构 generation mechanism 广义坐标 generalized coordinate轨迹生成 path generation 轨迹发生器 path generator滚刀 hob 滚道 raceway滚动体 rolling element 滚动轴承 rolling bearing滚动轴承代号 rolling bearing identification code 滚针 needle roller滚针轴承 needle roller bearing 滚子 roller滚子轴承 roller bearing 滚子半径 radius of roller滚子从动件 roller follower 滚子链 roller chain滚子链联轴器 double roller chain coupling 滚珠丝杆 ball screw滚柱式单向超越离合器 roller clutch 过度切割 undercutting函数发生器 function generator 函数生成 function generation含油轴承 oil bearing 耗油量 oil consumption耗油量系数 oil consumption factor 赫兹公式 H. Hertz equation合成弯矩 resultant bending moment 合力 resultant force合力矩 resultant moment of force 黑箱 black box横坐标 abscissa 互换性齿轮 interchangeable gears花键 spline 滑键、导键 feather key滑动轴承 sliding bearing 滑动率 sliding ratio滑块 slider 环面蜗杆 toroid helicoids worm环形弹簧 annular spring 缓冲装置 shocks; shock-absorber灰铸铁 grey cast iron 回程 return回转体平衡 balance of rotors 混合轮系 compound gear train积分 integrate 机电一体化系统设计 mechanical-electrical integration system design机构 mechanism 机构分析 analysis of mechanism机构平衡 balance of mechanism 机构学 mechanism机构运动设计 kinematic design of mechanism 机构运动简图 kinematic sketch of mechanism机构综合 synthesis of mechanism 机构组成 constitution of mechanism机架 frame, fixed link 机架变换 kinematic inversion机器 machine 机器人 robot机器人操作器 manipulator 机器人学 robotics技术过程 technique process 技术经济评价 technical and economic evaluation技术系统 technique system 机械 machinery机械创新设计 mechanical creation design, MCD 机械系统设计 mechanical system design, MSD机械动力分析 dynamic analysis of machinery 机械动力设计 dynamic design of machinery机械动力学 dynamics of machinery 机械的现代设计 modern machine design机械系统 mechanical system 机械利益 mechanical advantage机械平衡 balance of machinery 机械手 manipulator机械设计 machine design; mechanical design 机械特性 mechanical behavior机械调速 mechanical speed governors 机械效率 mechanical efficiency机械原理 theory of machines and mechanisms 机械运转不均匀系数 coefficient of speed fluctuation 机械无级变速 mechanical stepless speed changes 基础机构 fundamental mechanism基本额定寿命 basic rating life 基于实例设计 case-based design,CBD基圆 base circle 基圆半径 radius of base circle基圆齿距 base pitch 基圆压力角 pressure angle of base circle基圆柱 base cylinder 基圆锥 base cone急回机构 quick-return mechanism 急回特性 quick-return characteristics急回系数 advance-to return-time ratio 急回运动 quick-return motion棘轮 ratchet 棘轮机构 ratchet mechanism棘爪 pawl 极限位置 extreme (or limiting) position极位夹角 crank angle between extreme (or limiting) positions计算机辅助设计 computer aided design, CAD 计算机辅助制造 computer aided manufacturing, CAM计算机集成制造系统 computer integrated manufacturing system, CIMS计算力矩 factored moment; calculation moment 计算弯矩 calculated bending moment加权系数 weighting efficient 加速度 acceleration加速度分析 acceleration analysis 加速度曲线 acceleration diagram尖点 pointing; cusp 尖底从动件 knife-edge follower间隙 backlash 间歇运动机构 intermittent motion mechanism减速比 reduction ratio 减速齿轮、减速装置 reduction gear减速器 speed reducer 减摩性 anti-friction quality渐开螺旋面 involute helicoids 渐开线 involute渐开线齿廓 involute profile 渐开线齿轮 involute gear渐开线发生线 generating line of involute 渐开线方程 involute equation渐开线函数 involute function 渐开线蜗杆 involute worm渐开线压力角 pressure angle of involute 渐开线花键 involute spline简谐运动 simple harmonic motion 键 key键槽 keyway 交变应力 repeated stress交变载荷 repeated fluctuating load 交叉带传动 cross-belt drive交错轴斜齿轮 crossed helical gears 胶合 scoring角加速度 angular acceleration 角速度 angular velocity角速比 angular velocity ratio 角接触球轴承 angular contact ball bearing角接触推力轴承 angular contact thrust bearing 角接触向心轴承 angular contact radial bearing角接触轴承 angular contact bearing 铰链、枢纽 hinge校正平面 correcting plane 接触应力 contact stress接触式密封 contact seal 阶梯轴 multi-diameter shaft结构 structure 结构设计 structural design截面 section 节点 pitch point节距 circular pitch; pitch of teeth 节线 pitch line节圆 pitch circle 节圆齿厚 thickness on pitch circle节圆直径 pitch diameter 节圆锥 pitch cone节圆锥角 pitch cone angle 解析设计 analytical design紧边 tight-side 紧固件 fastener径节 diametral pitch 径向 radial direction径向当量动载荷 dynamic equivalent radial load 径向当量静载荷 static equivalent radial load径向基本额定动载荷 basic dynamic radial load rating径向基本额定静载荷 basic static radial load tating 径向接触轴承 radial contact bearing 径向平面 radial plane径向游隙 radial internal clearance 径向载荷 radial load径向载荷系数 radial load factor 径向间隙 clearance静力 static force 静平衡 static balance静载荷 static load 静密封 static seal局部自由度 passive degree of freedom 矩形螺纹 square threaded form锯齿形螺纹 buttress thread form 矩形牙嵌式离合器 square-jaw positive-contact clutch绝对尺寸系数 absolute dimensional factor 绝对运动 absolute motion绝对速度 absolute velocity 均衡装置 load balancing mechanism抗压强度 compression strength 开口传动 open-belt drive开式链 open kinematic chain 开链机构 open chain mechanism可靠度 degree of reliability 可靠性 reliability可靠性设计 reliability design, RD 空气弹簧 air spring空间机构 spatial mechanism 空间连杆机构 spatial linkage空间凸轮机构 spatial cam 空间运动副 spatial kinematic pair空间运动链 spatial kinematic chain 框图 block diagram空转 idle 宽度系列 width series雷诺方程Reynolds‘s equation 离心力 centrifugal force离心应力 centrifugal stress 理论廓线 pitch curve离合器 clutch 离心密封 centrifugal seal理论啮合线 theoretical line of action 隶属度 membership 力 force力多边形 force polygon 力封闭型凸轮机构 force-drive (or force-closed) cam mechanism 力矩 moment 力平衡 equilibrium力偶 couple 力偶矩 moment of couple连杆 connecting rod, coupler 连杆机构 linkage连杆曲线 coupler-curve 连心线 line of centers链 chain 链传动装置 chain gearing链轮 sprocket ; sprocket-wheel ; sprocket gear ; chain wheel 联组V 带 tight-up V belt联轴器 coupling ; shaft coupling 两维凸轮 two-dimensional cam临界转速 critical speed 六杆机构 six-bar linkage龙门刨床 double Haas planer 轮坯 blank轮系 gear train 螺杆 screw螺距 thread pitch 螺母 screw nut螺旋锥齿轮 helical bevel gear 螺钉 screws螺栓 bolts 螺纹导程 lead螺纹效率 screw efficiency 螺旋传动 power screw螺旋密封 spiral seal 螺纹 thread (of a screw)螺旋副 helical pair 螺旋机构 screw mechanism螺旋角 helix angle 螺旋线 helix ,helical line绿色设计 green design ; design for environment 马耳他机构 Geneva wheel ; Geneva gear马耳他十字 Maltese cross 脉动无级变速 pulsating stepless speed changes脉动循环应力 fluctuating circulating stress 脉动载荷 fluctuating load铆钉 rivet 迷宫密封 labyrinth seal密封 seal 密封带 seal belt密封胶 seal gum 密封元件 potted component密封装置 sealing arrangement 面对面安装 face-to-face arrangement面向产品生命周期设计 design for product`s life cycle, DPLC名义应力、公称应力 nominal stress模块化设计 modular design, MD 模块式传动系统 modular system模幅箱 morphology box 模糊集 fuzzy set模糊评价 fuzzy evaluation 模数 module摩擦 friction 摩擦角 friction angle摩擦力 friction force 摩擦学设计 tribology design, TD摩擦阻力 frictional resistance 摩擦力矩 friction moment摩擦系数 coefficient of friction 摩擦圆 friction circle磨损 abrasion ;wear; scratching 末端执行器 end-effector目标函数 objective function 耐腐蚀性 corrosion resistance耐磨性 wear resistance 内齿轮 internal gear挠性机构 mechanism with flexible elements 挠性转子 flexible rotor内齿圈 ring gear内力 internal force 内圈 inner ring能量 energy 能量指示图 viscosity逆时针 counterclockwise (or anticlockwise) 啮出 engaging-out啮合 engagement, mesh, gearing 啮合点 contact points啮合角 working pressure angle 啮合线 line of action啮合线长度 length of line of action 啮入 engaging-in牛头刨床 shaper 凝固点 freezing point; solidifying point扭转应力 torsion stress 扭矩 moment of torque扭簧 helical torsion spring 诺模图 NomogramO 形密封圈密封 O ring seal 盘形凸轮 disk cam盘形转子 disk-like rotor 抛物线运动 parabolic motion疲劳极限 fatigue limit 疲劳强度 fatigue strength偏置式 offset 偏( 心) 距 offset distance偏心率 eccentricity ratio 偏心质量 eccentric mass偏距圆 offset circle 偏心盘 eccentric偏置滚子从动件 offset roller follower 偏置尖底从动件 offset knife-edge follower偏置曲柄滑块机构 offset slider-crank mechanism 拼接 matching评价与决策 evaluation and decision 平底宽度 face width频率 frequency 平带 flat belt平带传动 flat belt driving 平底从动件 flat-face follower平分线 bisector 平均应力 average stress平均中径 mean screw diameter 平均速度 average velocity平衡 balance 平衡机 balancing machine平衡品质 balancing quality 平衡平面 correcting plane平衡质量 balancing mass 平衡重 counterweight平衡转速 balancing speed 平面副 planar pair, flat pair平面机构 planar mechanism 平面运动副 planar kinematic pair平面连杆机构 planar linkage 平面凸轮 planar cam平面凸轮机构 planar cam mechanism 平面轴斜齿轮 parallel helical gears普通平键 parallel key 其他常用机构 other mechanism in common use起动阶段 starting period 启动力矩 starting torque气动机构 pneumatic mechanism 奇异位置 singular position起始啮合点 initial contact , beginning of contact 气体轴承 gas bearing千斤顶 jack 嵌入键 sunk key强迫振动 forced vibration 切齿深度 depth of cut曲柄 crank 曲柄存在条件 Grashoff`s law曲柄导杆机构 crank shaper (guide-bar) mechanism 曲柄滑块机构 slider-crank (or crank-slider) mechanism 曲柄摇杆机构 crank-rocker mechanism 曲齿锥齿轮 spiral bevel gear曲率 curvature 曲率半径 radius of curvature曲面从动件 curved-shoe follower 曲线拼接 curve matching曲线运动 curvilinear motion 曲轴 crank shaft驱动力 driving force 驱动力矩 driving moment (torque)全齿高 whole depth 权重集 weight sets球 ball 球面滚子 convex roller球轴承 ball bearing 球面副 spheric pair球面渐开线 spherical involute 球面运动 spherical motion球销副 sphere-pin pair 球坐标操作器 polar coordinate manipulator燃点 spontaneous ignition 热平衡 heat balance; thermal equilibrium人字齿轮 herringbone gear 冗余自由度 redundant degree of freedom柔轮 flexspline 柔性冲击 flexible impulse; soft shock柔性制造系统 flexible manufacturing system; FMS 柔性自动化 flexible automation润滑油膜 lubricant film 润滑装置 lubrication device润滑 lubrication 润滑剂 lubricant三角形花键 serration spline 三角形螺纹 V thread screw三维凸轮 three-dimensional cam 三心定理 Kennedy`s theorem砂轮越程槽 grinding wheel groove 砂漏 hour-glass少齿差行星传动 planetary drive with small teeth difference设计方法学 design methodology设计变量 design variable 设计约束 design constraints深沟球轴承 deep groove ball bearing 生产阻力 productive resistance升程 rise 升距 lift实际廓线 cam profile 十字滑块联轴器double slider coupling; Oldham‘s coupling 矢量 vector 输出功 output work输出构件 output link 输出机构 output mechanism输出力矩 output torque 输出轴 output shaft输入构件 input link 数学模型 mathematic model实际啮合线 actual line of action 双滑块机构 double-slider mechanism, ellipsograph双曲柄机构 double crank mechanism 双曲面齿轮 hyperboloid gear双头螺柱 studs 双万向联轴节 constant-velocity (or double) universal joint 双摇杆机构 double rocker mechanism 双转块机构 Oldham coupling双列轴承 double row bearing 双向推力轴承 double-direction thrust bearing松边 slack-side 顺时针 clockwise瞬心 instantaneous center 死点 dead point四杆机构 four-bar linkage 速度 velocity速度不均匀( 波动) 系数 coefficient of speed fluctuation速度波动 speed fluctuation速度曲线 velocity diagram 速度瞬心 instantaneous center of velocity塔轮 step pulley 踏板 pedal台钳、虎钳 vice 太阳轮 sun gear弹性滑动 elasticity sliding motion 弹性联轴器 elastic coupling ; flexible coupling弹性套柱销联轴器 rubber-cushioned sleeve bearing coupling 套筒 sleeve梯形螺纹 acme thread form 特殊运动链 special kinematic chain特性 characteristics 替代机构 equivalent mechanism调节 modulation, regulation 调心滚子轴承 self-aligning roller bearing调心球轴承 self-aligning ball bearing 调心轴承 self-aligning bearing调速 speed governing 调速电动机 adjustable speed motors调速系统 speed control system 调压调速 variable voltage control调速器 regulator, governor 铁磁流体密封 ferrofluid seal停车阶段 stopping phase 停歇 dwell同步带 synchronous belt 同步带传动 synchronous belt drive凸的，凸面体 convex 凸轮 cam凸轮倒置机构 inverse cam mechanism 凸轮机构 cam , cam mechanism凸轮廓线 cam profile 凸轮廓线绘制 layout of cam profile凸轮理论廓线 pitch curve 凸缘联轴器 flange coupling图册、图谱 atlas 图解法 graphical method推程 rise 推力球轴承 thrust ball bearing推力轴承 thrust bearing 退刀槽 tool withdrawal groove退火 anneal 陀螺仪 gyroscopeV 带 V belt 外力 external force外圈 outer ring 外形尺寸 boundary dimension万向联轴器 Hooks coupling ; universal coupling 外齿轮 external gear弯曲应力 beading stress 弯矩 bending moment腕部 wrist 往复移动 reciprocating motion往复式密封 reciprocating seal 网上设计 on-net design, OND微动螺旋机构 differential screw mechanism 位移 displacement位移曲线 displacement diagram 位姿 pose , position and orientation稳定运转阶段 steady motion period 稳健设计 robust design蜗杆 worm 蜗杆传动机构 worm gearing蜗杆头数 number of threads 蜗杆直径系数 diametral quotient蜗杆蜗轮机构 worm and worm gear 蜗杆形凸轮步进机构 worm cam interval mechanism蜗杆旋向 hands of worm 蜗轮 worm gear涡圈形盘簧 power spring 无级变速装置 stepless speed changes devices无穷大 infinite 系杆 crank arm, planet carrier现场平衡 field balancing 向心轴承 radial bearing向心力 centrifugal force 相对速度 relative velocity相对运动 relative motion 相对间隙 relative gap象限 quadrant 橡皮泥 plasticine细牙螺纹 fine threads 销 pin消耗 consumption 小齿轮 pinion小径 minor diameter 橡胶弹簧 balata spring修正梯形加速度运动规律 modified trapezoidal acceleration motion修正正弦加速度运动规律 modified sine acceleration motion斜齿圆柱齿轮 helical gear 斜键、钩头楔键 taper key泄漏 leakage 谐波齿轮 harmonic gear谐波传动 harmonic driving 谐波发生器 harmonic generator斜齿轮的当量直齿轮 equivalent spur gear of the helical gear心轴 spindle 行程速度变化系数 coefficient of travel speed variation行程速比系数 advance-to return-time ratio 行星齿轮装置 planetary transmission行星轮 planet gear 行星轮变速装置 planetary speed changing devices行星轮系 planetary gear train 形封闭凸轮机构 positive-drive (or form-closed) cam mechanism 虚拟现实 virtual reality 虚拟现实技术 virtual reality technology, VRT虚拟现实设计 virtual reality design, VRD 虚约束 redundant (or passive) constraint许用不平衡量 allowable amount of unbalance许用压力角 allowable pressure angle 许用应力 allowable stress; permissible stress悬臂结构 cantilever structure 悬臂梁 cantilever beam循环功率流 circulating power load 旋转力矩 running torque旋转式密封 rotating seal 旋转运动 rotary motion选型 type selection 压力 pressure压力中心 center of pressure 压缩机 compressor压应力 compressive stress 压力角 pressure angle牙嵌式联轴器 jaw (teeth) positive-contact coupling雅可比矩阵 Jacobi matrix 摇杆 rocker液力传动 hydrodynamic drive 液力耦合器 hydraulic couplers液体弹簧 liquid spring 液压无级变速 hydraulic stepless speed changes液压机构 hydraulic mechanism 一般化运动链 generalized kinematic chain移动从动件 reciprocating follower 移动副 prismatic pair, sliding pair移动关节 prismatic joint 移动凸轮 wedge cam盈亏功 increment or decrement work 应力幅 stress amplitude应力集中 stress concentration 应力集中系数 factor of stress concentration应力图 stress diagram 应力—应变图 stress-strain diagram优化设计 optimal design 油杯 oil bottle油壶 oil can 油沟密封 oily ditch seal有害阻力 useless resistance 有益阻力 useful resistance有效拉力 effective tension 有效圆周力 effective circle force有害阻力 detrimental resistance余弦加速度运动 cosine acceleration (or simple harmonic) motion预紧力 preload 原动机 primer mover圆带 round belt 圆带传动 round belt drive圆弧齿厚 circular thickness 圆弧圆柱蜗杆 hollow flank worm圆角半径 fillet radius 圆盘摩擦离合器 disc friction clutch圆盘制动器 disc brake 原动机 prime mover原始机构 original mechanism 圆形齿轮 circular gear圆柱滚子 cylindrical roller 圆柱滚子轴承 cylindrical roller bearing圆柱副 cylindric pair 圆柱式凸轮步进运动机构 barrel (cylindric) cam圆柱螺旋拉伸弹簧 cylindroid helical-coil extension spring圆柱螺旋扭转弹簧 cylindroid helical-coil torsion spring圆柱螺旋压缩弹簧 cylindroid helical-coil compression spring圆柱凸轮 cylindrical cam 圆柱蜗杆 cylindrical worm圆柱坐标操作器 cylindrical coordinate manipulator圆锥螺旋扭转弹簧 conoid helical-coil compression spring圆锥滚子 tapered roller 圆锥滚子轴承 tapered roller bearing圆锥齿轮机构 bevel gears 圆锥角 cone angle原动件 driving link 约束 constraint约束条件 constraint condition 约束反力 constraining force跃度 jerk 跃度曲线 jerk diagram运动倒置 kinematic inversion 运动方案设计 kinematic precept design运动分析 kinematic analysis 运动副 kinematic pair运动构件 moving link 运动简图 kinematic sketch运动链 kinematic chain 运动失真 undercutting运动设计 kinematic design 运动周期 cycle of motion运动综合 kinematic synthesis 运转不均匀系数 coefficient of velocity fluctuation 运动粘度 kenematic viscosity 载荷 load载荷—变形曲线 load—deformation curve载荷—变形图 load—deformation diagram窄V 带 narrow V belt 毡圈密封 felt ring seal展成法 generating 张紧力 tension张紧轮 tension pulley 振动 vibration振动力矩 shaking couple 振动频率 frequency of vibration振幅 amplitude of vibration 正切机构 tangent mechanism正向运动学 direct (forward) kinematics 正弦机构 sine generator, scotch yoke织布机 loom 正应力、法向应力 normal stress制动器 brake 直齿圆柱齿轮 spur gear直齿锥齿轮 straight bevel gear 直角三角形 right triangle直角坐标操作器 Cartesian coordinate manipulator 直径系数 diametral quotient直径系列 diameter series 直廓环面蜗杆 hindley worm直线运动 linear motion 直轴 straight shaft质量 mass 质心 center of mass执行构件 executive link; working link 质径积 mass-radius product智能化设计 intelligent design, ID 中间平面 mid-plane中心距 center distance 中心距变动 center distance change中心轮 central gear 中径 mean diameter终止啮合点 final contact, end of contact 周节 pitch周期性速度波动 periodic speed fluctuation 周转轮系 epicyclic gear train肘形机构 toggle mechanism 轴 shaft轴承盖 bearing cup 轴承合金 bearing alloy轴承座 bearing block 轴承高度 bearing height轴承宽度 bearing width 轴承内径 bearing bore diameter轴承寿命 bearing life 轴承套圈 bearing ring轴承外径 bearing outside diameter 轴颈 journal轴瓦、轴承衬 bearing bush 轴端挡圈 shaft end ring轴环 shaft collar 轴肩 shaft shoulder轴角 shaft angle 轴向 axial direction轴向齿廓 axial tooth profile 轴向当量动载荷 dynamic equivalent axial load 轴向当量静载荷 static equivalent axial load轴向基本额定动载荷 basic dynamic axial load rating轴向基本额定静载荷 basic static axial load rating轴向接触轴承 axial contact bearing 轴向平面 axial plane轴向游隙 axial internal clearance 轴向载荷 axial load轴向载荷系数 axial load factor 轴向分力 axial thrust load主动件 driving link 主动齿轮 driving gear主动带轮 driving pulley 转动导杆机构 whitworth mechanism转动副 revolute (turning) pair 转速 swiveling speed ; rotating speed转动关节 revolute joint 转轴 revolving shaft转子 rotor 转子平衡 balance of rotor装配条件 assembly condition 锥齿轮 bevel gear锥顶 common apex of cone 锥距 cone distance锥轮 bevel pulley; bevel wheel 锥齿轮的当量直齿轮 equivalent spur gear of the bevel gear 锥面包络圆柱蜗杆 milled helicoids worm 准双曲面齿轮 hypoid gear子程序 subroutine 子机构 sub-mechanism自动化 automation 自锁 self-locking自锁条件 condition of self-locking 自由度 degree of freedom, mobility总重合度 total contact ratio 总反力 resultant force总效率 combined efficiency; overall efficiency 组成原理 theory of constitution组合齿形 composite tooth form 组合安装 stack mounting组合机构 combined mechanism 阻抗力 resistance最大盈亏功 maximum difference work between plus and minus work纵向重合度 overlap contact ratio 纵坐标 ordinate组合机构 combined mechanism 最少齿数 minimum teeth number最小向径 minimum radius 作用力 applied force坐标系 coordinate Piping work: 铺管工程Steam trace: 加热蒸汽管道Cutting: 切割socket weld承插焊接fillet weld角焊,填角焊branch connection分支接续fabrication tolerance.制造容差local heat treatment 局部热处理threaded pipe螺纹管seal welding.密封焊接flange joint 凸缘接头undercut 底切feeder馈电线conduit outlet电线引出口seal fitting 密封接头, 密封配件Screw thread lubricant螺纹润滑剂Seal: 绝缘层weld reinforcement 焊缝补强lock washer 锁紧[止动, 防松]垫圈electrical panel.配电板,配电盘nipple螺纹接头zinc plated.镀锌的ring joint 环接, 围缘接合bolt 螺栓control: 控制器National Electrical Code 全国电气规程master schedule 主要图表, 综合图表, 设计任务书, 主要作业表torque wrench 转矩扳手job site 施工现场flange connection.凸缘联接Hard hat：安全帽Goggles：护目镜stockpile贮存packing list装箱单crate: 柳条箱purchased material list原材料进货单back-feed反馈wire coil线盘，线卷，NPT thread. 美国标准锥管螺纹cable gland 电缆衬垫terminal block线弧, 接头排接线盒, 接线板, 线夹power drill机械钻connector. 接线器insulated sleeve绝缘套管wire connector接线器wire terminal电线接头control wiring控制线路motor lead电动机引出线power wiring电力布线tender documents提供证件orifice plate.挡板nut 螺母flange gasket 法兰垫片dimensional inspection 尺寸检验burn through 烧蚀piping system.管道系统reinforcement of weld加强焊缝fabrication.制造dye penetrant examination染料渗透试验法magnetic particle examination 磁粉检验girth weld环形焊缝cement lined piping 水泥衬里weld joint 焊缝, 焊接接头spool drawing 管路图, 管路详图spot test 抽查, 当场测试butt weld 对接焊缝Random Radiography随机射线照相检查radiographic examination 射线照相检查assembly.装配erection 架设examination试验cable tray.电缆盘rigid steel conduit 钢制电线管power control 功率控制arc welding 电弧焊control cable控制电缆操纵索normal bend 法向[法线]弯管cable glands: 电缆衬垫exfoliation剥落power receptacle 电力插座grounding conductor 接地导体lighting fixture照明器材junction box 分线箱race way 电缆管道terminal box接线盒distribution board配电盘, 配电屏receptacle 插座tumble switch.翻转开关,拨动式开关cathodic protection system 阴极保护系统Assembly line 组装线Layout 布置图Conveyer 流水线物料板Rivet table 拉钉机Rivet gun 拉钉枪Screw driver 起子Pneumatic screw driver 气动起子worktable 工作桌OOBA 开箱检查fit together 组装在一起fasten 锁紧(螺丝) fixture 夹具(治具)pallet 栈板barcode 条码barcode scanner 条码扫描器fuse together 熔合fuse machine热熔机repair修理operator作业员QC品管supervisor 课长ME 制造工程师MT 制造生技cosmetic inspect 外观检查inner parts inspect 内部检查thumb screw 大头螺丝lbs. inch 镑、英寸EMI gasket 导电条front plate 前板rear plate 后板chassis 基座bezel panel 面板power button 电源按键reset button 重置键Hi-pot test of SPS 高源高压测试V oltage switch of SPS 电源电压接拉键sheet metal parts 冲件plastic parts 塑胶件SOP 制造作业程序material check list 物料检查表work cell 工作间trolley 台车carton 纸箱sub-line 支线left fork 叉车personnel resource department 人力资源部production department生产部门planning department企划部QC Section品管科stamping factory冲压厂painting factory烤漆厂molding factory成型厂common equipment常用设备uncoiler and straightener整平机punching machine 冲床robot机械手hydraulic machine油压机lathe车床planer |plein|刨床miller铣床grinder磨床linear cutting线切割electrical sparkle电火花welder电焊机staker=reviting machine铆合机position职务president董事长general manager总经理special assistant manager特助factory director厂长department director部长deputy manager | =vice manager副理section supervisor课长deputy section supervisor =vice section superisor副课长group leader/supervisor组长line supervisor线长assistant manager助理to move, to carry, to handle搬运。
螺杆式压缩机.
关于螺杆压缩机的润滑
螺杆压缩机绝大部分采取喷油式润滑。通过 润滑油与节支混合进入压缩机，随着介质的流动 对机器进行润滑和密封。其优点如下： 1)降低排气温度。 2)减少工质泄漏，提高密封效果。 3)增强对零部件的润滑，提高零部件寿命。 4)对声波有吸收和阻尼作用，可以降低噪声。 5)冲洗掉机械杂质，减少磨损。
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5. 具有正排量压缩的特点，即排气量几乎不受 排气压力的影响，在小排气量时不发生喘振 现象，在宽广的工况范围内，仍可保持较高 的效率。 6. 机组的运行自动化程度高，可连续无级调节 ； 7. 对介质不敏感，可以采用喷油冷却，故在相 同的压力比下，排温比活塞式低得多，因此 单级压力比高； 8. 没有余隙容积，因而容积效率高；
面相适应，使转子精确地装入机体内。在机体内壁面设有符
合转子转角要求的径向吸气孔口，保证转子在旋转中顺利实 现吸气过程。 吸、排气端座是位于机体前后两端的密封连接件，它除 作机体的端面密封外，更重要的是提供了阴、阳转子和支承
转子的轴承装配位置。
转子
实现变容式压缩的主要部件，由阴、阳转子组成 。转子齿形是用高精度的专用机床、专用刀具加工而 成，是压缩机的关键零件之一。转子型线常为单边非 对称摆线——圆弧型线。阴、阳转子在结构上有以下 两种方式： 1. 阳转子与电动机联接为主动转子，传递转矩，同时、 通过啮合关系带动阴转子(从动转子)旋转。 2. 阴、阳转子通过各自的从动齿轮与电机带动的主动齿 轮啮合，子、机 体、轴承、轴封及流量调节装置等，有的 在压缩机出口还带有消声器。
机壳
由机体、吸气端座和排气端座组成，是压缩机的主要组
成部分。机体是连接各零部件的中心部件，它为各零部件提
供正确的装配位置，保证阴、阳转子在气缸内啮合，可靠地 进行工作。其端面形状为∞形，这与两个啮合转子的外圆柱
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附录1 The Original EnglishTHE KEY TECHNOLOGY OF DESIGN HOB FORHOBBING SCREW COMPRESSOR ROTORSWITH CUCLOID-ARC PROFILEABSTRCTThe profile of cycloid-arc screw compressor rotors is not a smooth profile; it has a tip on it. When design the hob cutter used for machining this kind of rotors, the profile of hob edge will appear separation. In this paper, the author made researches on the design theory of hob cutter for hobbing the cycloid-arc rotor with tip profile, and got the best way for design this kind of hob cutter with a separate edge. It is good practice to design the hob cutter and hob the cycloid-arc rotor according to practical design, manufacture and test.(1) INTRODUCTIONThe efficiency and reliability of screw compressor mainly depend on manufacturing technology of screw rotors. At present, the machining method of our country for machining screw rotors is milling the shortcoming of milling is low productivity and machining accuracy. Hobbing characteristic is high productivity and machining accuracy, so the machining method for hobbing instead of milling screw compressor rotors is now becoming more and more popular.Hobbing instead of milling for machining screw compressor rotors has much more advantage, but the key problem for carrying out hobbing the screw compressor rotors is that the profile of screw compressor rotors must be suited to hobbing. Our national standard profile for screw compressor rotors have no-symmetric cycloid-arc profile and symmetric are profile [1], since no-symmetric cycloid-arc profile screw compressor has much more advantage than symmetric are profile screw compressor, our national factory all adopt the former at present. The property of no-symmetriccycloid-arc profile is that the conjoint curve of profile isn’t slick curve, it has a tip on the profile, it is still a great difficult for hobbing instead of milling this kind of screw rotors in our cou ntry as the design problem of hob cutter. In this paper, we’ll make researches on the design theory of hob cutter for hobbing the no-symmetric cycloid-arc rotor with tip profile.(2 ) EXISTING PROBLEMFig.1 shows the end section of no-symmetric cycloid-arc rotors, its end profile is composed of radial line ab, arc bc, prolonged cycloid cd and radial line de. The point of intersection of prolonged cycloid cd and radial line de exist a tip d, that is, the d point of intersection hasn’t common tangent. As we calculate the corresponding axial profile of hob cutter according to cutting tool design handbook or other cutting tool design data, we’ll find that the axial profile of hob cutter becomes two separate curves, like the one shown in Fig.2.Fig.1 The end profile of screw rotor Fig.2 The axial profile of hobIn order to machining the required rotor profile and insure the tip not being cut out, people can usually take following two ways to solve this problem. One way is to prolong curve cd and radial line de as Fig.3 shows, this way can avoid appearing separate curve of hob edge, but hob profile will become Fig.4 shows, this kind of hob edge can neither be machined nor be used. Another way is to make a concave curve to link the separate hob edge as Fig.5 shows.Fig.3 The end profile of screw rotor Fig.4 The axial profile of hob Fig.5 The concave curve This way can avoid the tip being cut out, but it will produce two new tips on hob edge. This kind of hob is not only difficult to be machined but also easy to be worn on the tips. Form above discussing we can see that above two ways is not the best way to solve this problem. The best way to solve this kind of problem is to figure out the intermediate curve between separate edge curves accurately.(3) THE BEST WAY FOR CALCULATING INTERMEDIATE CURVE ACCURATELYHere we make use of the intermediate rack to calculate the intermediate curve between separate edge curves. That is, in the first place, we figure out the intermediate profile of rack according to the mesh of intermediate rack and rotor, in the second place, we figure out the intermediate profile of hob edge curve according to the mesh of intermediate rack and hob worm.According to gear mesh theorem, we can figure out the profile of intermediate rack mesh with rotor easily. As the tip exists on the profile of rotor, calculated profile of rotor will be two separate curves as Fig.2 shows. The two coordinates points d1 and d2 can easily figure out as following d1(x1, y1) and d2(x2, y2), obviously, the formation of separate curve of rack profile is that the tip d on rotor profile move around the rack to form when rack meshes with rotor. According to Fig.6 we can see, the mesh of rack and rotor is equal to pitch circle of rotor rolling on the pitch line of rack, the point d on rotor formed moving track is the intermediate curve of rack.Fig.6 The formation of separate curve on rackThe separate curve on rack can easily be given by the following equation:11sin()cos()t t x r y r θρθφρθφ=-+⎧⎨=-++⎩ (1) Where r is the pitch circle radius of rotor, ρ is the length of radial line od, Ф is the angle included between the radial line oe and the coordinate axis Y .θ is a variable, its bound is θ1≤θ≤θ2, θ1 and θ2 value can be calculated by the equation (1) according to the coordinate value of d1(x1, y1) and d2(x2, y2).As we know the separate curve equation on rack, we can figure out the separate curve equation of rack at the end of hob worm by the Fig.7 as following:1211cos /cos t t t t x x y y ββ=⎧⎨=⎩ (2) Where β1 is the spiral angle of rotor, β2 is the spiral angle of hob worm.According to gear mesh theorem [2], we can figure out the separate curve equation of hob worm mesh with the separate curve equation on rack by the Fig.7 as following:Fig.7 The common rack mesh with the rotor and the hob worm3111111311111111()cos ()sin ()cos ()sin ()//t t t t t t tt t x r x y r y y r x r y tgu x r u tg dy dx ϕϕϕϕϕϕϕ-=-+-⎧⎪=-+-⎪⎨=+⎪⎪=⎩ (3) According to formula (1), (2) and (3), we can accurately figure out the separate curve between the two separate profiles on hob edge.We can insure to hob the right profile of cycloid-arc rotor according to above formula to design the hob. It is good practice to design the hob cutter and hob the cycloid-arc rotor according to practical design, manufacture and test.4 REFERENCES[1] Li Wenling. Rotary Compressor for Refrigeration.Beijing: Mechanical Industry Press, 1992. 110~122. (in Chinese)[2] Li Rusheng. Design Principle of Cutting Tools. Nanjin: Science & Technology Press, 1985. 475~485. (in Chinese)2 中文翻译设计加工螺杆式压缩机的内摆线—弧轮廓所用滚刀的关键技术摘要螺杆式压缩机的内摆线—弧部分的轮廓并不是光滑的，它存在一个尖端。