英文原文 1 Introduction The screw compressor is one of the most common types of machine used to compress gases. Its construction is simple in that it essentially comprises only a pair of meshing rotors, with helical grooves machined in them, contained in a casing, which fits closely round them. The rotors and casing are separated by very small clearances. The rotors are driven by an external motor and mesh like gears in such a manner that, as they rotate, the space formed between them and the casing is reduced progressively. Thus, any gas trapped in this case is compressed. The geometry of such machines is complex and the flow of the gas being compressed within them occurs in three stages. Firstly, gas enters between the lobes, through an inlet port at one end of the casing during the start of rotation. As rotation continues, the space between the rotors no longer lines up with the inlet port and the gas is trapped and thus compressed. Finally, after further rotation, the opposite ends of the rotors pass a second port at the other end of the casing, through which the gas is discharged. The whole process is repeated between successive pairs of lobes to create a continuous but pulsating flow of gas from low to high pressure. These machines are mainly used for the supply of compressed air in the building industry, the food, process and pharmaceutical industries and, where required, in the metallurgical industry and for pneumatic transport. They are also used extensively for compression of refrigerants in refrigeration and air conditioning systems and of hydrocarbon gases in the chemical industry. Their relatively rapid acceptance over the past thirty years is due to their relatively high rotational speeds compared to other types of positive displacement machine, which makes them compact, their ability to maintain high efficiencies over a wide range of operating pressures and flow rates and their long service life and high reliability. Consequently, they constitute a substantial percentage of all positive displacement compressors now sold and currently in operation. The main reasons for this success are the development of novel rotor profiles, which have drastically reduced internal leakage, and advanced machine tools, which can manufacture the most complex shapes to tolerances of the order of 3 micrometers at an acceptable cost. Rotor profile enhancement is still the most promising means of further improving screw compressors and rational procedures are now being developed both to replace earlier empirically derived shapes and also to vary the proportions of the selected profile to obtain the best result for the application for which the compressor is required. Despite their wide usage, due to the complexity of their internal geometry and the non-steady nature of the processes within them, up till recently, only approximate analytical methods have been available to predict their performance. Thus, although it is known that their elements are distorted both by the heavy loads imposed by pressure induced forces and through temperature changes within them, no methods were available to predict the magnitude of these distortions accurately, nor how they affect the overall performance of the machine. In addition, improved modelling of flow patterns within the machine can lead to better porting design. Also, more accurate determination of bearing loads and how they fluctuate enable better choices of bearings to be made. Finally, if rotor and casing distortion, as a result of temperature and pressure changes within the compressor, can be estimated reliably, machining procedures can be devised to minimise their adverse effects. Screw machines operate on a variety of working fluids, which may be gases, dry vapour or multi-phase mixtures with phase changes taking place within the machine. They may involve oil flooding, or other fluids injected during the compression or expansion process, or be without any form of internal lubrication. Their geometry may vary depending on the number of lobes in each rotor, the basic rotor profile and the relative proportions of each rotor lobe segment. It follows that there is no universal configuration which would be the best for all applications. Hence, detailed thermodynamic analysis of the compression process and evaluation of the influence of the various design parameters on performance is more important to obtain the best results from these machines than from other types which could be used for the same application. A set of well defined criteria governed by an optimisation procedure is therefore a prerequisite for achieving the best design for each application. Such guidelines are also essential for the further improvement of existing screw machine designs and broadening their range of uses. Fleming et al., 1998 gives a good contemporary review of screw compressor modelling, design and application. A mathematical model of the thermodynamic and fluid flow processes within positive displacement machines, which is valid for both the screw compressor and expander modes of operation, is presented in this Monograph. It includes the use of the equations of conservation of mass, momentum and energy applied to an instantaneous control volume of trapped fluid within the machine with allowance for fluid leakage, oil or other fluid injection, heat transfer and the assumption of real fluid properties. By simultaneous solution of these equations, pressure-volume diagrams may be derived of the entire admission, discharge and compression or expansion process within the machine. A screw machine is defined by the rotor profile which is here generated by use of a general gearing algorithm and the port shape and size. This algorithm demonstrates the meshing condition which, when solved explicitly, enables a variety of rotor primary arcs to be defined either analytically or by discrete point curves. Its use greatly simplifies the design since only primary arcs need to be specified and these can be located on either the main or gate rotor or even on any other rotor including a rack, which is a rotor of infinite radius. The most efficient profiles have been obtained from a combined rotor-rack generation procedure. The rotor profile generation processor, thermofluid solver and optimizer,together with pre-processing facilities for the input data and graphical post processing and CAD interface, have been incorporated into a design tool in the form of a general computer code which provides a suitable tool for analysis and optimization of the lobe profiles and other geometrical and physical parameters. The Monograph outlines the adopted rationale and method of modelling, compares the shapes of the new and conventional profiles and illustrates potential improvements achieved with the new design when applied to dry and oil-flooded air compressors as well as to refrigeration screw compressors. The first part of the Monograph gives a review of recent developments in screw compressors. The second part presents the method of mathematical definition of the general case of screw machine rotors and describes the details of lobe shape specification. It focuses on a new lobe profile of a slender shape with thinner lobes in the main rotor, which yields a larger cross-sectional area and shorter sealing lines resulting in higher delivery rates for the same tip speed. The third part describes a model of the thermodynamics of the compression-expansion processes, discusses some modelling issues and compares the shapes of new and conventional profiles. It illustrates the potential improvements achievable with the new design applied to dry and oil-flooded air compressors as well as to refrigeration screw compressors. The selection of the best gate rotor tip radius is given as an example of how mathematical modelling may be used to optimise the design and the machines 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 effic iency. The prototypes were tested and compared with the best compressors currently on the market. The measured specific power input appeared to be lower than any published values for other equivalent compressors currently manufactured. Both the predicted advantages of the new rotor profile and the superiority of the design procedure were thereby confirmed. 1.1 Basic Concepts Thermodynamic machines for the compression and expansion of gases and vapours are the key components of the vast majority of power generation and refrigeration systems and essential for the production of compressed air and gases needed by industry. Such machines can be broadly classified by their mode of operation as either turbomachines or those of the positive displacement type. Turbomachines effect pressure changes mainly by dynamic effects, related to the change of momentum imparted to the fluids passing through them. These are associated with the steady flow of fluids at high velocities and hence these machines are compact and best suited for relatively large mass flow rates. Thus compressors and turbines of this type are mainly used in the power generation industry, where, as a result of huge investment in research and development programmes, they are designed and built to attain thermodynamic efficiencies of more than 90% in large scale power production plant. However, the production rate of machines of this type is relatively small and worldwide, is only of the order of some tens of thousands of units per annum. Positive displacement machines effect pressure changes by admitting a fixed mass of fluid into a working chamber where it is confined and then compressed or expanded and, from which it is finally discharged. Such machines must operate more or less intermittently. Such intermittent operation is relatively slow and hence these machines are comparatively large. They are therefore better suited for smaller mass flow rates and power inputs and outputs. A number of types of machine operate on this principle such as reciprocating, vane, scroll and rotary piston machines. In general, positive displacement machines have a wide range of application, particularly in the fields of refrigeration and compressed air production and their total world production rate is in excess of 200 million units per annum. Paradoxically, but possibly because these machines are produced by comparatively small companies with limited resources, relatively little is spent on research and development programmes on them and there are very few academic institutions in the world which are actively promoting their improvement. One of the most successful positive displacement machines currently in use is the screw or twin screw compressor. Its principle of operation, as indicated in Fig. 1.1, is based on volumetric changes in three dimensions rather than two. As shown, it consists, essentially, of a pair of meshing helical lobed rotors, contained in a casing. The spaces formed between the lobes on each rotor form a series of working chambers in which gas or vapour is contained. Beginning at the top and in front of the rotors, shown in the light shaded portion of Fig. 1.1a, there is a starting point for each chamber where the trapped volume is initially zero. As rotation proceeds in the direction of the arrows, the volume of that chamber then increases as the line of contact between the rotor with convex lobes, known as the main rotor, and the adjacent lobe of the gate rotor Fig. 1.1. Screw Compressor Rotors advances along the axis of the rotors towards the rear. On completion of one revolution i.e. 360 by the main rotor, the volume of the chamber is then a maximum and extends in helical form along virtually the entire length of the rotor. Further rotation then leads to reengagement of the main lobe with the succeeding gate lobe by a line of contact starting at the bottom and front of the rotors and advancing to the rear, as shown in the dark shaded portions in Fig. 1.1b. Thus, the trapped volume starts to decrease. On completion of a further 360 of rotation by the main rotor, the trapped volume returns to zero. The dark shaded portions in Fig. 1.1 show the enclosed region where therotors are surrounded by the casing, which fits closely round them, while the light shaded areas show the regions of the rotors, which are exposed to external pressure. Thus the large light shaded area in Fig. 1.1a corresponds to the low pressure port while the small light shaded region between shaft ends B and D in Fig. 1.1b corresponds to the high pressure port. Exposure of the space between the rotor lobes to the suction port, as their front ends pass across it, allows the gas to fill the passages formed between them and the casing until the trapped volume is a maximum. Further rotation then leads to cut off of the chamber from the port and progressive reduction in the trapped volume. This leads to axial and bending forces on the rotors and also to contact forces between the rotor lobes. The compression process continues until the required pressure is reached when the rear ends of the passages are exposed to the discharge port through which the gas flows out at approximately constant pressure. It can be appreciated from examination of Fig. 1.1, is that if the direction of rotation of the rotors is reversed, then gas will flow into the machine through the high pressure port and out through the low pressure port and it will act as an expander. The machine will also work as an expander when rotating in the same direction as a compressor provided that the suction and discharge ports are positioned on the opposite sides of the casing to those shown since this is effectively the same as reversing the direction of rotation relative to the ports. When operating as a compressor, mechanical power must be supplied to shaft A to rotate the machine. When acting as an expander, it will rotate automatically and power generated within it will be supplied externally through shaft A. The meshing action of the lobes, as they rotate, is the same as that of helical gears but, in addition, their shape must be such that at any contact position, a sealing line is formed between the rotors and between the rotors and the casing in order to prevent internal leakage between successive trapped passages. A further requirement is that the passages between the lobes should be as large as possible, in order to maximise the fluid displacement per revolution. Also, the contact forces between the rotors should be low in order to minimise internal friction losses. A typical screw rotor profile is shown in Fig. 1.2, where a configuration of 56 lobes on the main and gate rotors is presented. The meshing rotors are shown with their sealing lines, for the axial plane on the left and for the cross-sectional plane in the centre. Also, the clearance distribution between the two rotor racks in the transverse plane, scaled 50 times (6) is given above. Fig. 1.2. Screw rotor profile: (1) main, (2) gate, (3) rotor external and (4) pitch circles, (5) sealing line, (6) clearance distribution and (7) rotor flow area between the rotors and housing Oil injected Oil Free Fig. 1.3. Oil Injected and Oil Free Compressors Screw machines have a number of advantages over other positive displacement types. Firstly, unlike reciprocating machines, the moving parts all rotate and hence can run at much higher speeds. Secondly, unlike vane machines, the contact forces within them are low, which makes them very reliable. Thirdly, and far less well appreciated, unlike the reciprocating, scroll and vane machines, all the sealing lines of contact which define the boundaries of each cell chamber, decrease in length as the size of the working chamber decreases and the pressure within it rises. This minimises the escape of gas from the chamber due to leakage during the compression or expansion process. 1.2 Types of Screw Compressors Screw compressors may be broadly classified into two types. These are shown in Fig. 1.3 where machines with the same size rotors are compared: 1.2.1 The Oil Injected Machine This relies on relatively large masses of oil injected with the compressed gas in order to lubricate the rotor motion, seal the gaps and reduce the temperature rise during compression. It requires no internal seals, is simple in mechanical design, cheap to manufacture and highly efficient. Consequently it is widely used as a compressor in both the compressed air and refrigeration industries. 1.2.2 The Oil Free Machine Here, 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. These additional 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 Design Serious 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 parts Fig. 1.5. Cross section of a screw compressor with gear box To 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 m 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 m even in large batch production. This means that, as far as rotor production alone is concerned, clearances betweenthe rotors can be as small as 12 m. 中文譯文 1 引言 螺桿式壓縮機(jī)是一種最常見的用來壓縮氣體的機(jī)器。它的結(jié)構(gòu)簡單，因為它基本上只包括一對互相嚙合的螺旋形轉(zhuǎn)子，它們包含在一個殼體中，緊密地配合在周圍，轉(zhuǎn)子和機(jī)殼之間有非常小的空隙。轉(zhuǎn)子由外部的發(fā)動機(jī)和齒輪驅(qū)動，當(dāng)轉(zhuǎn)子旋轉(zhuǎn)時，轉(zhuǎn)子間的空隙以及轉(zhuǎn) 子與機(jī)殼之間的空間逐漸縮小。因此，在這種情況下，里面的氣體就被壓縮了。這種機(jī)器的幾何構(gòu)造是復(fù)雜的，流動的氣體被壓縮分三個階段。首先，通過啟動時轉(zhuǎn)子的旋轉(zhuǎn)，氣體通過機(jī)殼一端的進(jìn)氣口進(jìn)入齒間。其次，隨著旋轉(zhuǎn)的繼續(xù)，轉(zhuǎn)子之間的空間不再與進(jìn)氣口聯(lián)合，因此，里面的氣體就被壓縮。最后，進(jìn)一步旋轉(zhuǎn)之后，轉(zhuǎn)子的另一端經(jīng)過位于機(jī)殼另一端的第二個端口，通過這個端口氣體被排出。整個過程在齒的持續(xù)嚙合中重復(fù)著，形成了一個從低壓到高壓持續(xù)而脈動的氣流。 這些機(jī)器主要用于建筑業(yè)、食品業(yè)、加工業(yè)、制藥業(yè)以及在冶金工業(yè)和氣動輸送中需要的地 方。它們也廣泛地應(yīng)用于制冷和空調(diào)系統(tǒng)中所用的制冷劑的壓縮以及化學(xué)工業(yè)中烴氣的壓縮。在過去 30 年，它們之所以相對快速地被接受是由于相比其他類型的容積式機(jī)器，它們擁有相對高的轉(zhuǎn)速，這使得他們更加緊湊，它們能夠在大范圍的運(yùn)作壓力下保持高效率、高流速并且使用壽命長，依賴性強(qiáng)。因此，螺桿式壓縮機(jī)在當(dāng)前出售和使用的容積式壓縮機(jī)中占有很大比例。 螺桿式壓縮機(jī)成功的主要原因歸功于新型轉(zhuǎn)子，它已大大減少了內(nèi)部漏泄，以及先進(jìn)的機(jī)床，它能以一個可接受的成本制造出最復(fù)雜的形狀到 3 微米的公差。轉(zhuǎn)子齒形的改善仍是最有前景的手段來更好地 改進(jìn)螺桿式壓縮機(jī)。合理的程序正被研發(fā)，不僅用來代替早期經(jīng)驗得出的形狀，而且能多樣化地協(xié)調(diào)已選的輪廓，獲得最佳的應(yīng)用程序的壓縮機(jī)必需的。盡管螺桿式壓縮機(jī)大范圍使用，但是由于它們內(nèi)部構(gòu)造的復(fù)雜性以及工序本身的不穩(wěn)定性，直到現(xiàn)在，只有近似的分析方法能來預(yù)測它們的表現(xiàn)。因此，雖然眾所周知，它們的單體被重物所施加的壓力以及被自身的溫度變化所扭曲，目前沒法能精確預(yù)測這些扭曲的程度，也不能預(yù)測它們?nèi)绾斡绊憴C(jī)器的總性能。另外，機(jī)器內(nèi)部優(yōu)化的流型模式有助更好的排氣口設(shè)計。而且，更精確的測定軸承負(fù)載和它們波動的方式有助于更好的 選擇軸承。最后，由于壓縮機(jī)自身溫度和壓力變化而導(dǎo)致轉(zhuǎn)子和機(jī)殼扭曲，如果這個扭曲能被確實地預(yù)估，那么就能設(shè)計加工程序最大化地降低不利影響。 在螺桿機(jī)操作下的各種工作氣體，其可以是與機(jī)器內(nèi)的相位變化發(fā)生的氣體，干蒸汽或多相混合物。他們可能會涉及驅(qū)油或其他液體注入過程中的壓縮或膨脹過程中，沒有任何形式的內(nèi)部潤滑。它們的幾何形狀可以依據(jù)每個轉(zhuǎn)子上的齒，基本的轉(zhuǎn)子輪廓和每個轉(zhuǎn)齒片段上的相對部分來進(jìn)行變化。因此沒有通用的配置，適合各種應(yīng)用程序的。因此，壓縮過程詳細(xì)的熱力學(xué)分析和對各種性能設(shè)計參數(shù)的評估更能從這些機(jī)器上獲 得最好的結(jié)果，而非從能達(dá)到相同的用途的其他類型的機(jī)器上。所以，在優(yōu)化程序主導(dǎo)下設(shè)計一套詳細(xì)的標(biāo)準(zhǔn)是每個應(yīng)用程序?qū)崿F(xiàn)最好設(shè)計的前提。這樣的指導(dǎo)方針，為進(jìn)一步改善現(xiàn)有的螺桿機(jī)設(shè)計并擴(kuò)大其廣泛的用途也很重要。弗萊明等人在 1998 年對螺桿式壓縮機(jī)的建模、設(shè)計和應(yīng)用提供了一個很好的現(xiàn)代化評論。 熱力學(xué)和流體流動過程的數(shù)學(xué)模型在這個專著中被提出，它在容積式機(jī)的螺桿壓縮機(jī)和膨脹機(jī)的操作模式是有效的。專著中包括質(zhì)量守恒定律公式的使用，包括運(yùn)用于機(jī)器中的截面液即時控制卷的動力和能量，而這機(jī)器撥備液體滲漏、石油或其他流體注入、 傳熱和實際流體參數(shù)。通過求解這些方程組，能獲得機(jī)器中流體整體的吸收、釋放、壓縮或膨脹過程的壓力容積圖。 螺桿機(jī)的容量由轉(zhuǎn)子輪廓而定，而轉(zhuǎn)子輪廓根據(jù)普遍傳動裝置的使用以及氣門的形狀和大小。這個運(yùn)算法則當(dāng)解釋明確時，能論證嚙合條件，即能使多種轉(zhuǎn)子的主要弧線通過分析或者離散點曲線得出。它的使用大大簡化了設(shè)計，因為只有主要弧線需要被額定，而且這些能被設(shè)于主轉(zhuǎn)子或閘轉(zhuǎn)子或者其他任何包括齒條的擁有無限大半徑的轉(zhuǎn)子上。最有效的轉(zhuǎn)子輪廓圖已從一個聯(lián)合轉(zhuǎn)子架生成過程中得出。 轉(zhuǎn)子輪廓形成處理器，熱流體的解算器和優(yōu)化器，連同預(yù) 處理設(shè)施為輸入數(shù)據(jù)和圖形后處理和 CAD接口，已被納入一個設(shè)計工具，它以通過計算機(jī)編碼的形式為分析和齒形優(yōu)化以及其他幾何和物理參數(shù)提供一個適當(dāng)?shù)墓ぞ?。這個專著對比新的和傳統(tǒng)的輪廓，概述了采用的基本原理和建模方法，并闡明了當(dāng)應(yīng)用于有油或者無油的空氣壓縮機(jī)和制冷螺桿式壓縮機(jī)時，新的設(shè)計有潛在的改進(jìn)方面。 專著的第一部分綜述了螺旋式壓縮機(jī)當(dāng)前的發(fā)展。 第二部分介紹了用數(shù)學(xué)的方法定義在一般情況下螺桿機(jī)轉(zhuǎn)子的方法，還細(xì)節(jié)的描述了齒形的規(guī)范。它強(qiáng)調(diào)在主轉(zhuǎn)子上擁有更薄的新型的薄型齒，這樣的齒能形成一個更大的截面面積和更短的 密封線以致同等的轉(zhuǎn)速能產(chǎn)生更高的給料速度。 第三部分描述了壓縮膨脹過程的熱力學(xué)模型，討論了一些建模的問題，對比新的和傳統(tǒng)的剖面形狀。它闡明了隨著新的設(shè)計應(yīng)用于無油和有油式空氣壓縮機(jī)和制冷螺桿壓縮機(jī)時所能達(dá)到的潛在的改進(jìn)。最好的閘轉(zhuǎn)子齒頂圓半徑的選擇被作為一個例子來描述，說明數(shù)學(xué)模型可以用來優(yōu)化設(shè)計和機(jī)器的操作條件。 第四部分描述了設(shè)計一個高效螺桿壓縮機(jī)與新轉(zhuǎn)子配置。一個得到確鑿證實的容積式機(jī)器上壓縮過程的數(shù)學(xué)模型能夠決定轉(zhuǎn)子的最佳尺寸和速度、體積比率、油的注入位置和噴射口直徑。另外，像一個開放的進(jìn)氣孔、早期 曝露在外的排氣口，還有改良的軸承和密封件的規(guī)格等現(xiàn)代設(shè)計觀念都包含在內(nèi)，以最大限度地提高壓縮機(jī)的效率。這個原型被測試過并與當(dāng)前市場上最好的壓縮機(jī)作了對比。實測功率輸入似乎比當(dāng)前生產(chǎn)的其他任何等效壓縮機(jī)的公布值要低。預(yù)測的新轉(zhuǎn)子外形的優(yōu)勢和設(shè)計程序的優(yōu)越性因而得以證實。 1.1 基本概念 用于氣體和蒸汽壓縮和膨脹的熱力學(xué)機(jī)器是大多數(shù)發(fā)電和制冷系統(tǒng)的關(guān)鍵組分，而對工業(yè)上需壓縮空氣和氣體來說也是必不可少的。這種機(jī)器可以根據(jù)它們的運(yùn)行方式來廣泛地分類，分成渦輪機(jī)或者容積式機(jī)型。 渦輪機(jī)主要通過動態(tài)效果引起氣壓變化，而 這與傳遞給液體的動量的變化有關(guān)。這些與高速下液體的穩(wěn)流有關(guān)，因此，這些機(jī)器結(jié)構(gòu)緊湊，最適于相對大的流量。因此，這種類型的壓縮機(jī)和渦輪機(jī)主要是用于發(fā)電行業(yè)。由于研究和開發(fā)項目上巨大的投資，它們經(jīng)過設(shè)計和改造，能夠在大規(guī)模的電力生產(chǎn)工廠獲得 90%以上的熱力學(xué)效率。但是，在全世界范圍內(nèi)這種類型機(jī)器的生產(chǎn)率相對較低，而且每年的訂單只有數(shù)萬個單位。 容積式機(jī)器通過使定量的液體進(jìn)入工作腔，在那壓縮或膨脹再釋放，從而引起氣壓變化。這樣的機(jī)器必須或多或少間歇性地工作。這種間歇的運(yùn)轉(zhuǎn)相對緩慢，因而這些機(jī)器相對較大。因此，它們 更加適合小規(guī)模的流量和功率輸入與輸出。許多種機(jī)器采用這種工作原理，例如往復(fù)式，葉片，滾動和旋轉(zhuǎn)式活塞機(jī)器。 一般來說 ,容積式機(jī)器有廣泛的應(yīng)用 ,尤其在制冷、壓縮空氣領(lǐng)域，他們總世界產(chǎn)量超過 2 億輛每年。出乎意料的是，不過這可能是由于這些機(jī)器是由相對較小的公司生產(chǎn)的，資源有限所以研發(fā)項目上的投入比較少，很少有世界學(xué)術(shù)機(jī)構(gòu) ,積極促進(jìn)他們的進(jìn)步。 (a)從正面和頂部看 (b)從底部和后面看 圖 1.1 螺桿式壓縮機(jī)轉(zhuǎn)子 目前使用的最成功的容積式機(jī)器是螺桿式壓縮機(jī)和雙螺