螺杆压缩机机械外文文献翻译、中英文翻译、外文翻译

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单螺杆型使用的是机械外文文献翻译、中英文翻译、外文翻译

中国地质大学长城学院本科毕业设计外文资料翻译系别：工程技术系专业：机械设计制造及其自动化姓名：李江学号： 052083072012 年 4 月 20 日外文资料翻译译文塑料工业是与国民经济发展和社会文明建设息息相关的重要产业。

塑料工业的机械和装备的水平对该工业的发展起着关键作用。

比起传统的塑料挤出机，单螺杆塑料挤出机有他积极的优势，通常他能加工出高分子量、高粘度热塑性好的塑料。

其中有高生产率低熔点、高熔体强压力的塑料有利于化学降解，产品质量高，品，因此是最佳的塑料生产这使得单螺杆塑料挤出机生产经济，特别适用于稳定的挤压。

螺杆的回转航程和固定筒壁的相互作用是挤出机的泵出过程中必要的参数。

为了运输塑料材料，其摩擦在螺杆的表面要低，但在固定的筒壁要高。

如果达不到这个基本标准，塑料可能会随着螺杆旋转，而不是在轴向/输出方向上移动。

在输出区域，螺杆和机筒的表面通常都覆盖着的溶解物以及来自溶解物和螺杆通道之间的外力，而其除了处理有极高粘性的材料时都是无效的，如硬质PVC 材料和有超高分子量的聚乙烯。

溶解物流在输出部分是受内摩擦系数（粘度）影响，尤其是当模具提供了一个高阻力的熔体流时。

常见且更多使用的单螺杆型使用的是传统设计，即机筒和螺杆保持基本一致的直径，包括具有例如减小螺杆通道体积，有连续可变速度、压力控制，和通风（挥发）系统的挤出机。

一些特殊的设计使用了圆锥或抛物线外形的螺杆，用以达到特殊的混合和捏合效果。

它们可以包含偏心的核心，根据不同坡度变化的动程，揉捏转子，适应性的核心环，和间歇的轴向运动。

桶内可能有螺纹，可伸缩的螺杆形状以及进料设备。

一个成功的挤出操作需要密切注意很多细节，如（1）进给材料的质量和在适当温度下的物质流，（2）足以融化、但不会分解聚合物的温度曲线，以及（3）不会分解塑料的启动和关机。

应采取措施，防止促进塑料表面上水分的胶合和湿气的吸收凝结，如颜料浓缩物中的色素。

表面凝结，可通过储存于密封的塑料容器（吸湿性塑料）中在使用前约24 小时与工作区域保持同温来避免。

螺杆压缩机性能的计算吸入室中占主导地位外文文献翻译、中英文翻译、外文翻译

英文原文3.1 One Dimensional Mathematical Model 51The Conservation of Internal Energyθωω.d dv p Q h m h m dQ du out out in in += （3.1） where θ is angle of rotation of the main rotor, h = h (θ) is specific enthalpy, m ˙ = m ˙ (θ) is mass flow rate p = p (θ), fluid pressure in the working chamber control volume, ˙Q = ˙Q (θ), heat transfer between the fluid and the compressor surrounding, ˙V = ˙V (θ) local volume of the compressor working chamber.In the above equation the subscripts in and out denote the fluid inflow and outflow. The fluid total enthalpy inflow consists of the following components:oil oil g l g l suc suv in in h m h m h m h .,,...m ++= (3.2) where subscripts l , g denote leakage gain suc, suction conditions, and oil denotes oil. The fluid total outflow enthalpy consists of:l l l l dis dis ou out h m h m h m ,,..+= (3.3) where indices l , l denote leakage loss and dis denotes the discharge conditions with m ˙ dis denoting the discharge mass flow rate of the gas contaminated with the oil or other liquid injected.The right hand side of the energy equation consists of the following terms which are model The heat exchange between the fluid and the compressor screw rotors and casing and through them to the surrounding, due to the difference in temperatures of gas and the casing and rotor surfaces is accounted for by the heat transfer coefficient evaluated from the expression Nu = 0.023 Re0.8. For the characteristic length in the Reynolds and Nusselt number the difference between the outer and inner diameters of the main rotor was adopted. This may not be the most appropriate dimension for this purpose, but the characteristic length appears in the expression for the heat transfer coefficient with the exponent of 0.2 and therefore has little influence as long as it remains within the same order of magnitude as other characteristic dimensions of the machine and as long as it characterizes the compressor size. The characteristic velocity for the Re number is computed from the local mass flow and the cross-sectional area. Here the surface over which the heat is exchanged, as well as the wall temperature, depend on the rotation angle θ of the main rotor. The energy gain due to the gas inflow into the working volume is represented by the product of the mass intake and its averaged enthalpy. As such, the energy inflow varies with the rotational angle. During the suction period, gas enters the working volume bringing the averaged gas enthalpy,52 3 Calculation of Screw Compressor Performance which dominates in the suction chamber. However, during the time when the suction port is closed, a certain amount of the compressed gas leaks into the compressor working chamber through the clearances. The mass ofthis gas, as well as its enthalpy are determined on the basis of the gas leakage equations. The working volume is filled with gas due to leakage only when the gas pressure in the space around the working volume is higher, otherwise there is no leakage, or it is in the opposite direction, i.e. from the working chamber towards other plenums.The total inflow enthalpy is further corrected by the amount of enthalpy brought into the working chamber by the injected oil.The energy loss due to the gas outflow from the working volume is defined by the product of the mass outflow and its averaged gas enthalpy. During delivery, this is the compressed gas entering the discharge plenum, while, in the case of expansion due to inappropriate discharge pressure, this is the gas which leaks through the clearances from the working volume into the neighbouring space at a lower pressure. If the pressure in the working chamber is lower than that in the discharge chamber and if the discharge port is open, the flow will be in the reverse direction, i.e. from the discharge plenum into the working chamber. The change of mass has a negative sign and its assumed enthalpy is equal to the averaged gas enthalpy in the pressure chamber.The thermodynamic work supplied to the gas during the compression process is represented by the term pdV d θ . This term is evaluated from the local pressure and local volume change rate. The latter is obtained from the relationships defining the screw kinematics which yield the instantaneous working volume and its change with rotation angle. In fact the term dV/d ϕ can be identified with the instantaneous interlobe area, corrected for the captured and overlapping areas. If oil or other fluid is injected into the working chamber of the compressor, the oil mass inflow and its enthalpy should be included in the inflow terms. In spite of the fact that the oil mass fraction in the mixture is significant, its effect upon the volume flow rate is only marginal because the oil volume fraction is usually very small. The total fluid mass outflow also includes the injected oil, the greater part of which remains mixed with the working fluid. Heat transfer between the gas and oil droplets is described by a first order differential equation. The Mass Continuity Equationout out in in h m h m d m d ...θω= （3.4） The mass inflow rate consists of:oil g l suc in in m m m h m .,..++= （3.5）3.1 One Dimensional Mathematical Model 53The mass outflow rate consists of: .,..l l dis out m m m += (3.6)Each of the mass flow rate satisfies the continuity equationA m ρω=. (3.7) where w [m/s] denotes fluid velocity, ρ – fluid density and A – the flow crosssectionarea. The instantaneous density ρ = ρ(θ) is obtained from the instantaneous mass m trapped in the control volume and the size of the corresponding instantaneous volume V , as ρ = m/V .3.1.2 Suction and Discharge PortsThe cross-section area A is obtained from the compressor geometry and it may be considered as a periodic function of the angle of rotation θ. The suction port area is defined by:⎪⎪⎭⎫ ⎝⎛=suc o suc A A θθπsin ,suc (3.8) where suc means the starting value of θ at the moment of the suction port opening, and A suc , 0 denotes the maximum value of the suction port crosssection area. The reference value of the rotation angle θ is assumed at the suction port closing so that suction ends at θ = 0, if not specified differently.The discharge port area is likewise defined by:⎪⎪⎭⎫ ⎝⎛--=s e c o dis A A θθθθπsin ,dis (3.9)where subscript e denotes the end of discharge, c denotes the end of compression and A dis , 0 stands for the maximum value of the discharge port crosssectional area.Suction and Discharge Port Fluid Velocities)(212h h -=μω (3.10)where μ is the suction/discharge orifice flow coefficient, while subscripts 1 and 2 denote the conditions downstream and upstream of the considered port. The provision supplied in the computer code will calculate for a reverse flow if h 2 < h 1.54 3 Calculation of Screw Compressor Performance3.1.3 Gas LeakagesLeakages in a screw machine amount to a substantial part of the total flow rate and therefore play an important role because they influence the process both by affecting the compressor mass flow rate or compressor delivery, i.e. volumetric efficiency and the thermodynamic efficiency of the compression work. For practical computation of the effects of leakage upon the compressor process, it is convenient to distinguish two types of leakages, according to their direction with regard to the working chamber: gain and loss leakages. The gain leakages come from the discharge plenum and from the neighbouring working chamber which has a higher pressure. The loss leakages leave the chamber towards the suction plenum and to the neighbouring chamber with a lower pressure.Computation of the leakage velocity follows from consideration of the fluid flow through the clearance. The process is essentially adiabatic Fanno-flow. In order to simplify the computation, the flow is is sometimes assumed to be at constant temperature rather than at constant enthalpy. This departure from the prevailing adiabatic conditions has only a marginal influence if the analysis is carried out in differential form, i.e. for the small changes of the rotational angle, as followed in the present model. The present model treats only gas leakage. No attempt is made to account for leakage of a gas-liquid mixture, while the effect of the oil film can be incorporated by an appropriate reduction of the clearance gaps.An idealized clearance gap is assumed to have a rectangular shape and the mass flow of leaking fluid is expressed by the continuity equation:g l l l A m ωρμ=. (3.11)where r and w are density and velocity of the leaking gas, Ag = lg δg the clearance gap cross-sectional area, lg leakage clearance length, sealing line, δg leakage clearance width or gap, μ = μ(Re, Ma) the leakage flow discharge coefficient.Four different sealing lines are distinguished in a screw compressor: the leading tip sealing line formed between the main and gate rotor forward tip and casing, the trailing tip sealing line formed between the main and gate reverse tip and casing, the front sealing line between the discharge rotor front and the housing and the interlobe sealing line between the rotors.All sealing lines have clearance gaps which form leakage areas. Additionally, the tip leakage areas are accompanied by blow-hole areas.According to the type and position of leakage clearances, five different leakages can be identified, namely: losses through the trailing tip sealing and front sealing and gains through the leading and front sealing. The fifth, “throughleakage ” does not directly affect the process in the working chamber, but it passes through it from the discharge plenum towards the suction port. The leaking gas velocity is derived from the momentum equation, which accounts for the fluid-wall friction:3.1 One Dimensional Mathematical Model 5502211=++Dg dxf dpd l ωρωω (3.12)where f (Re, Ma) is the friction coefficient which is dependent on the Reynolds and Mach numbers, Dg is the effective diameter of the clearance gap, Dg ≈ 2δg and dx is the length increment. From the continuity equation and assuming that T ≈ const to eliminate gas density in terms of pressure, the equation can be integrated in terms of pressure from the high pressure side at position 2 to the low pressure side at position 1 of the gap to yield:⎪⎪⎭⎫ ⎝⎛+-==1222122.ln 2m p p a A g l l ςρρωρ (3.13)where ζ = fLg/Dg + Σξ characterizes the leakage flow resistance, with Lg clearance length in the leaking flow direction, f friction factor and ξ local resistance coefficient. ζ can be evaluated for each clearance gap as a function of its dimensions and shape and flow characteristics. a is the speed of sound.The full procedure requires the model to include the friction and drag coefficients in terms of Reynolds and Mach numbers for each type of clearance.Likewise, the working fluid friction losses can also be defined in terms of the local friction factor and fluid velocity related to the tip speed, density, and elementary friction area. At present the model employs the value of ζ in terms of a simple function for each particular compressor type and use. It is determined as an input parameter.These equations are incorporated into the model of the compressor and employed to compute the leakage flow rate for each clearance gap at the local rotation angle θ.3.1.4 Oil or Liquid InjectionInjection of oil or other liquids for lubrication, cooling or sealing purposes, modifies the thermodynamic process in a screw compressor substantially. The following paragraph outlines a procedure for accounting for the effects of oil injection. The same procedure can be applied to treat the injection of any other liquid. Special effects, such as gas or its condensate mixing and dissolving in the injected fluid or vice versa should be accounted for separately if they are expected to affect the process. A procedure for incorporating these phenomena into the model will be outlined later.A convenient parameter to define the injected oil mass flow is the oil-to-gas mass ratio, m oil /m gas, from which the oil inflow through the open oil port, which is assumed to be uniformly distributed, can be evaluated asπ21....z m m m m gas oiloil = (3.14) where the oil-to-gas mass ratio is specified in advance as an input parameter56 3 Calculation of Screw Compressor PerformanceIn addition to lubrication, the major purpose for injecting oil into a compressor is to cool the gas. To enhance the cooling efficiency the oil is atomized into a spray of fine droplets by means of which the contact surface between the gas and the oil is increased. The atomization is performed by using specially designed nozzles or by simple high-pressure injection. The distribution of droplet sizes can be defined in terms of oil-gas mass flow and velocity ratio for a given oil-injection system. Further, the destination of each distinct size of oil droplets can be followed until it hits the rotor or casing wall by solving the dynamic equation for each droplet size in a Lagrangian frame, accounting for inertia gravity, drag, and other forces. The solution of the droplet energy equation in parallel with the momentum equation should yield the amount of heat exchange with the surrounding gas.In the present model, a simpler procedure is adopted in which the heat exchange with the gas is determined from the differential equation for the instantaneous heat transfer between the surrounding gas and an oil droplet. Assuming that the droplets retain a spherical form, with a prescribed Sauter mean droplet diameter dS , the heat exchange between the droplet and the gas can be expressed in terms of a simple cooling law Qo = hoAo (T gas − T oil), where Ao is the droplet surface, Ao = d 2 S π, dS is the Sauter mean diameter of the droplet and ho is the heat transfer coefficient on the droplet surface, determined from an empirical expression. The exchanged heat must balance the rate of change of heat taken or given away by the droplet per unit time, Qo = moc oil dTo/dt = moc oil ωdTo/d θ, where c oil is the oil specific heat and the subscript o denotes oil droplet. The rate of change of oil droplet temperature can now be expressed as:()oilo o gas o c m T T A h d dT ωθ-=00 (3.15) The heat transfer coefficient ho is obtained from:33.06.0Pr Re 6.02u +=N (3.16)Integration of the equation in two time/angle steps yields the new oil droplet temperature at each new time/angle step:k kT T T po gas o +-=1, (3.17)where To,p is the oil droplet temperature at the previous time step and k is the non-dimensional time constant of the droplet, k = τ/Δt = ωτ/Δθ, with τ = moc oil /hoAo being the real time constant of the droplet. For the given Sauter mean diameter, dS , the non-dimensional time constant takes the formθωθω∆=∆=o oil S O o oil o h c d A h c m k 6 (3.18) The derived droplet temperature is further assumed to represent the average temperature of the oil, i.e. T oil ≈ To , which is further used to compute the enthalpy of the gas-oil mixture.3.1 One Dimensional Mathematical Model 57The above approach is based on the assumption that the oil-droplet time constant τ is smaller than the droplet travelling time through the gas before it hits the rotor or casing wall, or reaches the compressor discharge port. This means that heat exchange is completed within the droplet travelling time through the gas during compression. This prerequisite is fulfilled by atomization of the injected oil. This produces sufficiently small droplet sizes to gives a small droplet time constant by choosing an adequate nozzle angle, and, to some extent, the initial oil spray velocity. The droplet trajectory computed independently on the basis of the solution of droplet momentum equation for different droplet mean diameters and initial velocities. Indications are that for most screw compressors currently in use, except, perhaps for the smallest ones, with typical tip speeds of between 20 and 50m/s, this condition is well satisfied for oil droplets with diameters below 50 μm. For more details refer to Stosic et al., 1992.Because the inclusion of a complete model of droplet dynamics would complicate the computer code and the outcome would always be dependant on the design and angle of the oil injection nozzle, the present computation code uses the above described simplified approach. This was found to be fully satisfactory for a range of different compressors. The input parameter is only the mean Sauter diameter of the oil droplets, dS and the oil properties – density, viscosity and specific heat.3.1.5 Computation of Fluid PropertiesIn an ideal gas, the internal thermal energy of the gas-oil mixture is given by:()()()()oil oil gasoil gas mcT pV mcT mRT mu mu T +-=+-=+=11γγ (3.19)where R is the gas constant and γ is adiabatic exponentHence, the pressure or temperature of the fluid in the compressor working chamber can be explicitly calculated by input of the equation for the oil temperature T oil:()()()()()oilOIL mc mR k mcT U k T ++-+-=111γ (3.20) If k tends 0, i.e. for high heat transfer coefficients or small oil droplet size, the oil temperature fast approaches the gas temperature.In the case of a real gas the situation is more complex, because the temperature and pressure can not be calculated explicitly. However, since the internal energy can be expressed as a function of the temperature and specific volume only, the calculation procedure can be simplified by employing the internal energy as a dependent variable instead of enthalpy, as often is the practice. The equation of state p = f 1(T,V ) and the equation for specific internal energy u = f 2(T,V ) are usually decoupled. Hence, the temperature can be calculated from the known specific internal energy and the specific volume obtained from the solution of differential equations, whereas the pressure中文译文33.1一维数学模型 51内部能量守恒θωω.d dv p Q h m h m dQ du out out in in += (3.1) 其中θ是角度的旋转的主旋翼h =h （ θ ）的比焓，m ˙ =m ˙ （ θ ）是质量流率p = （ θ ），工作腔的控制体积中的流体压力， ˙ Q = Q˙（ θ ）的流体之间的热传递和压缩机周围， ˙ V = ˙V （ θ ），压缩机工作腔中的本地卷。

外文翻译----设计加工螺杆式压缩机的内摆线

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。

包络法的资产负债-螺杆压缩机转子外文文献翻译、中英文翻译、外文翻译

英文原文AEnvelope Method of GearingFollowing Stosic 1998, screw compressor rotors are treated here as helical gears with nonparallel and nonintersecting, or crossed axes as presented at Fig. A.1. x01, y01 and x02, y02are the point coordinates at the end rotor section in the coordinate systems fixed to the main and gate rotors, as is presented in Fig. 1.3. Σ is the rotation angle around the X axes. Rotation of the rotor shaft is the natural rotor movement in its bearings. While the main rotor rotates t hrough angle θ, the gate rotor rotates through angle τ = r1w/r2wθ = z2/z1θ, where r w and z are the pitch circle radii and number of rotor lobes respectively. In addition we define external and internal rotor radii: r1e= r1w+ r1 and r1i= r1w− r0. The dista nce between the rotor axes is C = r1w+ r2w. p is the rotor lead given for unit rotor rotation angle. Indices 1 and 2 relate to the main and gate rotor respectively.Fig. A.1. Coordinate system of helical gears with nonparallel and nonintersecting AxesTh e procedure starts with a given, or generating surface r1(t, θ) for which a meshing, or generated surface is to be determined. A family of such gener-ated surfaces is given in parametric form by: r2(t, θ, τ ), where t is a proﬁle parameter while θ and τ ar e motion parameters.r 1 =r 1(t, θ)=[ x 1,y 1,z 1]=x 01cosθ-y 01 sinθ, x 01 sinθ+ y 01 cosθ,p 1θ] (A,.1) ⎥⎦⎤⎢⎣⎡∂∂∂∂=∂∂0,,111t y t x t r=⎥⎦⎤⎢⎣⎡∂∂+∂∂∂∂-∂∂0,cos sin ,sin cos 0101011θθθθt y t x t y t x (A.2) []0,,0,,01010111x y y x r -=⎥⎦⎤⎢⎣⎡∂∂∂∂=∂∂θθθ (A.3) [][]∑+∑∑-∑-===c o s s i n ,s i n c o s ,,,),,(1111122222z y z y C x z y x t r r τθ []202020202,sin sin ,sin cos p y x y x ττττ+-= (A.4) [][]2020202022222,sin cos ,sin sin ,,p y x y x p x y r τττττ-+=-=∂∂ []∑--∑∑-+∑∑-∑=s i n )(c o s ,c o s )(s i n ,c o s s i n 121211C x p C x p y p θ (A.5) The envelope equation, which determines meshing between the surfaces r1 and r2:0222=∂∂∙⎪⎭⎫ ⎝⎛∂∂⨯∂∂τθr r t r (A.6) together with equations for these surfaces, completes a system of equations. If a generating surface 1 is deﬁned by the parameter t, the envelope may be used to calculate another parameter θ, now a function of t, as a meshing condition to define a generated surface 2, now the function of both t and θ. The cross product in the envelope equation represents a surface normal and ∂r2 ∂τ is the relative, sliding velocity of two single points on the surfaces 1 and 2 which together form the common tangential point of contact of these two surfaces. Since the equality to zero of a scalar triple product is an invariant property under the applied coordinate system and since the relative velocity may be concurrently represented in both coordinate systems, a convenient form of the meshing condition is deﬁned as:0211111=∂∂∙⎪⎭⎫ ⎝⎛∂∂⨯∂∂-=∂∂∙⎪⎭⎫ ⎝⎛∂∂⨯∂∂τθθθr r t r r r t r （A.7） Insertion of previous expressions into the envelope condition gives:[]⎪⎭⎫ ⎝⎛∂∂+∂∂∑-+-t y y t x x p p x C 1111211cot )( 0)cot (12111=⎥⎦⎤⎢⎣⎡∂∂∑-+∂∂+t x C p t y p p θ (A.8) This is applied here to derive the condition of meshing action for crossed helical gears of uniform lead with nonparallel and nonintersecting axes. The method constitutes a gear generation procedure which is generally applicable. It can be used for synthesis purposes of screw compressor rotors, which are electively helical gears with parallel axes. Formed tools for rotor manufacturing are crossed helical gears on non parallel and non intersecting axes with a uniform lead, as in the case of hobbing, or with no lead as in formed milling and grinding. Templates for rotor inspection are the same as planar rotor hobs. In all these cases the tool axes do not intersectthe rotor axes.Accordingly the notes present the application of the envelope method to produce a meshing condition for crossed helical gears. The screw rotor gearing is then given as an elementary example of its use while a procedure for forming a hobbing tool is given as a complex case.The shaft angle Σ, centre dist ance C, and unit leads of two crossed helical gears, p1 and p2 are not interdependent. The meshing of crossed helical gears is still preserved: both gear racks have the same normal cross section proﬁle, and the rack helix angles are related to the shaft angle as Σ = ψr1+ ψr2. This is achieved by the implicit shift of the gear racks in the x direction forcing them to adjust accordingly to the appropriate rack helix angles. This certainly includes special cases, like that of gears which may be orientated so that the shaft angle is equal to the sum of the gear helix angles: Σ = ψ1+ ψ2. Furthermore a centre distance may be equal to the sum of the gear pitch radii :C = r1+ r2.Pairs of crossed helical gears may be with either both helix angles of the same sign or each of opposite sign, left or right handed, depending on the combination of their lead and shaft angle Σ. The meshing condition can be solved only by numerical methods. For the given parameter t, the coordinates x01 and y01 and their derivatives ∂x01∂t and ∂y01∂t are known. A guessed value of parameter θ is then used to calculate x1, y1, ∂x1 ∂t and ∂y1∂t. A revised value of θ is then derived and the procedure repeated until the difference between two consecutive values becomes sufficiently small.For give n transverse coordinates and derivatives of gear 1 proﬁle, θ can be used to calculate the x1, y1, and z1 coordinates of its helicoid surfaces. The gear 2 helicoid surfaces may then be calculated. Coordinate z2 can then be used to calculate τ and ﬁnally, its transverse proﬁle point coordinates x2, y2 can be obtained.A number of cases can be identiﬁed from this analysis.(i) When Σ = 0, the equation meets the meshing condition of screw machine rotors and also helical gears with parallel axes. For such a case, the gear helix angles have the same value, but opposite sign and the gear ratio i = p2/p1 is negative. The same equation may also be applied for the gen-eration of a rack formed from gears. Additionally it describes the formed planar hob, front milling tool and the template control instrument.122 A Envelope Method of Gearing(ii) If a disc formed milling or grinding tool is considered, it is suffcient to place p2= 0. This is a singular case when tool free rotation does not affect the meshing process. Therefore, a reverse transformation cannot be obtained directly.(iii) The full scope of the meshing condition is required for the generation of the proﬁle of a formed hobbing tool. This is therefore the most compli-cated type of gear which can be generated from it.BReynolds Transport TheoremFollowing Hanjalic, 1983, Reynolds Transport Theorem deﬁnes a change of variable φ in a control volume V limited by area A of which vector the local normal is dA and which travels at local speed v. This control volume may, but need not necessarily coincide with an engineering or physical material system. The rate of change of variable φ in time within the volume is:⎰∂∂=⎪⎭⎫ ⎝⎛∂∂vV dV t t ρφφ (B.1) Therefore, it may be concluded that the change of variable φ in the volume V is caused by: – change of the speciﬁc variable m /φϕ=in time within the volume because of sources (and sinks) in the volume, ⎪⎭⎫ ⎝⎛∂∂t ϕdV which is called a local change and – movement of the control volume which takes a new space with variable ϕ in it and leaves its old space, causing a change in time of ϕfor ρϕv.dA and which is called convective changeThe ﬁrst contribution may be represented by a volume integral:.()dV t V⎰∂∂ρϕ (B.2) while the second contribution may be represented by a surface integral:⎰⋅AdA V ρϕ (B.3) Therefore:()⎰⎰⎰⋅+∂∂==⎪⎭⎫ ⎝⎛∂∂AV V V dA V dV t dV dt d t ρϕρϕρϕφ ( B.4) which is a mathematical representation of Reynolds Transport Theorem.Applied to a material system contained within the control volume V m which has surface A m and velocity v which is identical to the fluid velocity w, Reynolds Transport Theorem reads:()dA W dV t d dt d t AmVm Vm Vm ⋅+∂∂==⎪⎭⎫ ⎝⎛∂∂⎰⎰⎰ρϕρϕρϕφV (B.5) If that control volume is chosen at one instant to coincide with the control volume V , the volume integrals are identical for V and Vm and the surface integrals are identical for A and Am , however, the time derivatives of these integrals are different, because the control volumes will not coincide in the next time interval. However, there is a term which is identical for the both timesintervals:()()dV t dV t VmV ⎰⎰∂∂=∂∂ρϕρϕ (B.6) therefore,⎰⎰⋅-⎪⎭⎫ ⎝⎛∂∂=⋅-⎪⎭⎫ ⎝⎛∂∂AV Am Vm dA v t dA w t ρϕφρϕφ (B.7) or:()dA v w t t AV Vm ⋅-+⎪⎭⎫ ⎝⎛∂∂=⎪⎭⎫ ⎝⎛∂∂⎰ρϕφφ (B.8) If the control volume is ﬁxed in the coordinat e system, i.e. if it does not move, v = 0 and consequently:()dV tt V V ⎰∂∂=⎪⎭⎫ ⎝⎛∂∂ρϕφ (B.9) therefore:()⎰⎰⋅+∂∂=⎪⎭⎫ ⎝⎛∂∂AV Vm dA w dV t t ρϕρϕφ (B.10) Finally application of Gauss theorem leads to the common form:()()dV w dV t t VV Vm ⎰⎰⋅∇+∂∂=⎪⎭⎫ ⎝⎛∂∂ρϕρϕφ (B.11) As stated before, a change of variable φ is caused by the sources q within the volume V and influences outside the volume. These effects may be proportional to the system mass or volume or they may act at the system surface.The ﬁrst effect is given by a volume integral and the second effect is given by a surface integral. ()⎰⎰⎰⎰=⋅∇+=⋅+=⎪⎭⎫ ⎝⎛∂∂VV A Vm Am A Vm qdV dV q qv dA q qvdV t φ (B.12) q can be scalar, vector or tensor.The combination of the two last equations gives:()⎰⎰⎰=⋅+∂∂A VV qdV dA w dV t ρϕρϕ Or:()()0=⎥⎦⎤⎢⎣⎡-⋅∇+∂∂⎰dV q w t V ρϕρϕ (B.13) Omitting integral signs gives:()()0=-⋅∇+∂∂q w tρϕρϕ (B.14)This is the well known conservation law form of variable ρϕφ=. Since for ϕ = 1, this becomes the continuity equation: ()0=⋅∇+∂∂w tρρ ﬁnally it is: ()()0=-∇⋅+⎥⎦⎤⎢⎣⎡⋅∇+∂∂q w w t ϕρρρϕ Or: ()q w tdt D =∇⋅+∂∂=ϕρϕρϕ (B.15) dt D /ϕ is the material or substantial derivative of variable ϕ. This equation is very convenient for the derivation of particular conservation laws. As previously mentioned ϕ = 1 leads to the continuity equation, ϕ = u to the momentum equation, ϕ = e, where e is speciﬁc internal energy, leads to the energy equation, ϕ = s, to the entropy equation and so on.If the surfaces, where the fluid carrying var iable Φ enters or leaves the control volume, can be identiﬁed, a convective change may conveniently be written:∙∙∙-∙-∙Φ-Φ=-==⋅⎰⎰out in out in A m m m d dA w )()(ϕϕϕϕρ (B.16) where the over scores indicate the variable average at entry/exit surface sections. This leads to the macroscopic form of the conservation law:()Q m m Q dt d dt d out in out in VV +-=+Φ-Φ=⎥⎦⎤⎢⎣⎡=⎪⎭⎫ ⎝⎛Φ∙-∙-∙∙)()(ϕϕρϕ (B.17) which states in words: (rate of change of Φ) = (inflow Φ) − (outflow Φ) +(source of Φ)中文译文A包络法的资产负债螺杆压缩机转子Stosic 1998年之后，被视为非平行不相交的螺旋齿轮，或在图的交叉轴。

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。

Atlas Copco阿特拉斯科普柯固定式螺杆空气压缩机中英文说明书-推荐下载

Atlas Copco 阿特拉斯科普柯固定式螺杆空气压缩机电脑控制器操作方法控制面板控制面板功能：1， I ：开机按钮。

按该按钮让空压机开机，自动运行指示灯点亮时表示电脑控制器有效运行（自动运行状态）。

2，O ：停机按钮。

按该按钮让空压机停机，自动运行指示灯熄灭。

3，液晶显示器：显示关于空压机运行状态，保养要求或者故障信息。

4，“↑”，“↓”：上下滚动键：该键用来翻阅显示屏。

5，“ 〓 ”：表格键：该键用来选择跟着水平箭头的参数。

6，“F1”，“F2”，“F3”：功能键：该键用于控制和编制空压机，参见后面的内容。

7，“●自动”：自动运行指示灯：表示电脑控制器正自动控制空压机运行。

8，“●报警”：总报警指示灯：通常熄灭，当有不正常情况出现时，该灯点亮或闪烁，参见后面的内容。

9，“●电源”：电源指示灯：指示和上电源。

10：自动运行图标。

11，：报警图标。

12，13，另外，空压机机体上有一只紧急停机按钮：有紧急情况出现时，按紧急急停按钮让空压机立刻停机，故障排除后，拔出按钮解除锁定，并按重新设置功能键F3复位。

1显示屏-键1.1 典型显示图空压机出气压力/Compressor outlet7.0 bar自动加载/Automatically loaded↓菜单/Menu卸载/unloadF1F2F3显示器有四行：1)前三行:——显示屏显示传感器的名称和实际的读数。

——测量值的单位和传感器的实际读数。

——关于空压机运行的信息（空压机停机，等等），保养要求（比如油过滤器和空气过滤器）或者是故障信息（比如故障停机）。

2）在四行位于三个功能键的上方（F1/F2/F3），显示这些键（F1/F2/F3）对应的实际功能。

1．2 滚动键（↑↓）这些滚动键标有垂直的箭头，允许滚动显示屏。

只要在显示屏的最右边的位置上有一个指向下面的箭头，其对应的滚动键就可以用来查阅下面的内容。

只要在显示屏的最右边的位置上有一个指向上面的箭头，其对应的滚动键就可以用来查阅前面的内容。

压力机简介外文文献翻译、中英文翻译

外文资料Introduction to pressPress is a compact structure versatility press. With a versatile, high production efficiency, the press can be widely used in cutting, punching, blanking, bending, riveting and forming and other processes. By applying strong pressure on the metal blank metal to plastic deformation and fracture processed into parts. Large pulley driven by a motor (usually doubles as a flywheel) by mechanical presses work triangle belt driven through a clutch gear pair and slider-crank mechanism, the slider and punch straight down.1 Basic Introduction.1.1 Screw PressScrew press with a screw and nut as a transmission mechanism, and by the flywheel screw drive forward and reverse rotary motion into reciprocating slide down forging machinery. When working, the electric motor to rotate the flywheel to accelerate energy savings, through the screw, nut push the slider downward movement. When the slider contacts the workpiece, is forced to decelerate to a complete stop of theflywheel, into rotational kinetic energy of the impact energy stored by the slider against the workpiece to deform. After the fight, the flywheel motor reversing, driving the slider to rise back to its original position. Screw press with a nominal specification work force to represent.1.2 crank pressCrank press is one of the most common cold stamping equipment, used for cold stamping die working platform. Its structure is simple, easy to use. By the bed in the form of the body structure is different, crank press can be divided into open or closed crank press Crank presses; driven by the number of links can be divided into single-point or multi-point press press; press slide number is one or two can be divided into single-action or double-action press press.1.3 Friction PressFriction press machine is a strong universal pressure processing machine is widely used, can be used in a variety of pressure processing industries. In the machinery manufacturing industry, friction press applications more widely available to complete forging, upsetting, bending, correction, coining work, and some have no flash forging presses to accomplish this. Because of the press in the use of a larger universal, and therestructure, installation, handling and auxiliary equipment is simple and low-cost advantages, so in machinery manufacturing, automobile, tractor and aviation industries in the press shop, forge shop and forging workshops are widely used, but also for blanking and refractory suppression work.Over 1.4 pressesMulti-station press is an advanced press equipment, is an integrated multi-press, usually by head unit, feed mechanism, consisting of the press and the tail section. Fastest beat up to 40 beats / min or more, to meet the high-speed automated production. Unstacking head unit can be divided into units, magnetic belts and cleaning, oiling equipment; feed mechanism feeding arms generally composed; presses are generally divided into more than one slider and slider, choose according to the different needs,Tail section is generally constituted by a conveyor belt.2 mechanical structurePress consists of four parts: a four-column hydraulic press machine; combined control cabinet; electric heating systems and insulation devices; mold conveyor gantry. The above composition with integrated design, make nice shape,appearance, compact structure is simple, reliable, easy to maintain.2.1 on a four-column hydraulic press machinePress should have a reliable structural rigidity resistance to deformation, hydraulic station location, press row set mold hoisting connected devices, hydraulic station and press row has a removable dust cover. Can the direction of the workpiece length (3000) direction. Technical parameters are as follows: Nominal 190T; effective table size 3000 ×750 mm2; pressure exhaust velocity of 75 ~ 100 mm / s; maximum open distance of 550 mm (excluding heating plate); dwell time 8h (workpiece 130 ℃) ; press the base height of 0.5 ~ 0.55m; press row and base flatness of 0.2 mm; press row lamination base average gap ≤0.25 mm (not less than 10 measuring points).2.2 Heating SystemHeating the heating system is made of steel plate, heating, insulation layer. Two upper and lower heating plates are connected with the press and the press row fixed base and easy removal, heating plate made of high quality 45 # steel tune tribute treatment. Electric heating plate placed in the heating tube and a thermocouple. Thermocouple placement fasteningreliable, easy to disassemble. Heating plate Technical parameters: the plane size 3050 ×650 mm2; flatness of 0.2 mm; measuring point arrangement: on board 3:00; under the board 6:00. Heating pipes made of stainless steel outer casing, operating voltage 220V. Unilateral qualify, the number is a multiple of 3. Electric heating pipe replacement convenient and reliable. Wiring of the protective devices. Insulation layer is disposed between the heating plate and the base (or press row), the heating plate is surrounded by the insulation plate provided good insulation performance should be selected materials, the heating plate is quite planar dimensions, the thickness of not less than 20 mm.2.3 Control CabinetRack presses the action and temperature cabinet. Instrument panel and switch on the cabinet; power switch, indicator, voltmeter, ammeter; press the action button and indicator light; pressure indicating instrument; heating switch and indicator light; digital temperature regulator (accuracy ±1 ℃); insulation timing control, over-temperature alarm; maximum temperature 180 ℃; heating rate of 1.7 ℃~ 2.5 ℃/ min; temperature inhomogeneities ±3 ℃, the heating power of each phase power balance; temperature regulator temperatureregion corresponding real-time control, When all reaches the set value ±3 ℃, timer expires after a set time, the temperature control system can automatically process heat or power by the setting.2.4 Mold transport aircraftA carriage with an upper input roller structure presses one end connected to the base (removable), one end placed on the ground (height-adjustable). Carrying capacity 3T; lose height on the surface of the roll heating plate flush; plane size (2800 ~ 2900) ×750mm2 (available part); compartment temperature pressure plate on press row and around the base to reduce mold heating process The heat loss with aluminum (or stainless steel) and the insulation material. To facilitate the placement may be divided into several pieces.3 mechanical principlesPresses are usually driven by a motor through the friction disc flywheel rim leaving the flywheel rotates, so this press known as friction press, China's largest cattle friction press is 25 megabytes. Larger size presses with hydraulic system drives the flywheel, called hydraulic screw press, the maximum size of 125 megabytes cattle. Later appeared with an electric motordirectly drives the flywheel presses, It's compact, less transmission links, due to the frequent change of control of the higher electrical requirements, and requires special motors.Rotary presses no fixed lower dead point of the larger die forgings, can repeatedly blow molding can be singles, and even play jog. Strike force and deformation of the workpiece on large deformation hitting power, small deformation (such as cold hit) when combating force. In these respects, it is similar to the hammer. But it's a slider low speed (about 0.5 m / sec, 1/10 only hammer), the striking force through a closed rack, so smooth, vibration is much smaller than the hammer, do not need a great foundation. Presses equipped with skid insurance agencies, the maximum impact force limited to two times the nominal pressure to protect the safety equipment.The lower forging presses are equipped with top of the device. Both screw press forging hammer, forging machinery and other mechanical presses role, universal and strong, can be used for forging, stamping, deep drawing processes. Furthermore, screw press, friction press particularly simple structure, easy to manufacture, it is widely used. Screw press drawback is lower productivity and mechanical efficiency.4 Mechanical UseImproper operation or stamping presses, die set type presses the leading cause damage and downtime. Appropriate training and stamping die press operator set can ensure that they follow the correct procedures. It will be able to quickly reduce downtime.Operations to the former class points brake actuator shaft filling oil, stick ball classes before the first class of the day, No. 20-30 grease gun filling machine oil amount, clutch pressure injection site every day before class time with lubricating oil gun . Class before stopping the machine for cleaning.Check the fasteners, padded exterior missing pieces. Check the clutch and springs, belts. Check machine lubrication devices. Check the electrical wiring is damaged, aging, motors, solenoids are normal. Check the accuracy of the crankshaft and rail wear. Check the brakes, clutch, slider, closed block, closed loop. Check the electrical control. Coupling bolts fuselage bench testing and adjustments.According to press a different machine types and processing requirements, develop targeted, practical safety procedures and make the necessary job training and safety education. Using the unit and the operator must strictly complywith the provisions and procedures of design and manufacturing units to provide safety instructions and proper use, maintenance. Press generally safe operation requirements are as follows: Before moving equipment, to check manipulation of the press section, clutch and brakes are in an active state, the security guards are complete handy slider-crank mechanism ministries without exception. Abnormal should immediately take the necessary measures not sick operation, disassembly and damaged safety devices is strictly prohibited. Prior to the formal job subject to idle test, confirmed only after all part of the normal work. Should clean up before starting work bench all unnecessary items, to prevent car vibration or shock switch off the wounded man suddenly start causing sliders. Operation must use tools, prohibited by hand directly into his die extract, hand tools shall not be placed on the mold.When adjusting the position of the workpiece in the die area or strip stuck in the mold of the workpiece, the foot must leave the foot pedals. People operating the same press should have a unified command, the signal is clear, until the other party to make a clear answer, and then confirm the action away from the danger zone. Sudden power failure or operator should be finished off the power, and the manipulator to return to neutralclutch, brake in braking. Presses case for overhaul, adjustments and the installation, adjustment, disassembly die, disconnect energy (eg electricity, gas, liquid) in the machine, the machine stops operation carried out, and put the pads in the slider plus reliable Support. Listed at the start switch machine warning notice.5 Mechanical Maintenance5.1 attention to safety during maintenanceEnsure correct locking / opening process to ensure the safety of maintenance personnel to ensure the brakes prior to carry out maintenance work in the bottom of the sleeve into the trip. If this is done, it is no longer a lock sleeve.5.2 Maintenance ProcessSo the press operator presses added to the process to maintain them in case of problems can always remind maintenance or management personnel. They have carried out the operation of the press every day, so it can more easily hear or see signs of the problem. As a daily driver with a car, an exception can be noticed for the first time in the sound, which is also the same for the operator.5.3 Environmental Control workshopA clean press allows the operator or maintenance personnel can quickly discover if a problem occurs. Such as oil spills, leakage, breakage, etc., if the press is clean, then it's easy to find out its location.5.4 ensure that the press is in the equilibrium positionPrecision balanced press can get a better job, so best to check once a year. Charged with the operation of pneumatic systems and balance of pneumatic brake systems, you need to check whether there is a gas leak, because the Ministry of proper air pressure can affect performance brakes and balance system, and they control the press to stop time, if there are problems will make the operation and equipment in danger. In addition, all systems have pneumatic regulators, lubricators and reservoir. Accumulation in the gas pipeline water a day should be excluded.5.5 for some time to replace the oil and filterImproper lubrication system for the maintenance of the press is one of the main reasons the press downtime occurs, for some reason, a lot of press operator at the time of the oil circulation system with screen operation with occasional replacement of the filter. Should ensure that the replacement of oil while replacing the filter, usually also need regularreplacement.5.6 Maintenance of mechanical presses keyFor a mechanical press routine maintenance projects, which aspects are most important? The answer is the operator. All maintenance and issue press Find to start on the operator, the operator can often find problems early in the press, for the press to prevent the major components generated sustained damage. The operator can press attention by strange noise emitted; abnormal temperature rise, smoke, debris or some parts of the metal particles occurs; pipeline leaks, and found a problem. The operator should ask ourselves every day some questions answered by checking records to maintain long-term maintenance records.6 InstallationUsers will press shipped to the installation site, open the box, and then according to the instructions for installation work, will press on cement foundations have been beaten with a level tested around the table at the front and rear of the fuselage, horizontal degrees required to allow every length 1000 mm 0.20 mm, adjust if uneven material iron, until compliance with the requirements. The foundation is shown in Figure ten. After thecement solidification should be tightened evenly feet Lo Lo playing Manager (M24), with a spirit level to check the levelness of the fuselage table, presses must be completely solidified to begin work on the basis. After the installation is complete press should carefully kerosene oil will rust on the machined surface wash pressesKerosene not fall on the paint surface to avoid paint damage. Should be carefully cleaned and depression shoulder clean, dry after cleaning with cotton yarn, coated with a thin layer of oil. And check the lubrication hole is smooth, not be clogging.After the well, the press should be installed on a reliable grounding wire.中文翻译压力机简介压力机是一种结构精巧的通用性压力机。

机械设计制造及自动化中英文对照外文翻译文献

中英文对照外文翻译文献(文档含英文原文和中文翻译)使用CBN砂轮对螺杆转子进行精密磨削的方法摘要：针对高精度加工螺杆转子，这篇论文介绍了利用立方氮化硼（CBN砂轮）对螺杆转子进行精密磨削的加工方法。

首先，使用小型电镀CBN砂轮磨削螺杆转子。

精确的CBN砂轮轴向轮廓的模型是在齿轮啮合理论的基础上建立开发的。

考虑到螺杆转子和涂层厚度之间的间隙，主动砂轮的修整引入了CBN的砂轮的设计方法。

主动砂轮的形状采用低速电火花线切割技术（低速走丝线切割机）进行加工线CBN主动砂轮的成形车刀采用低速走丝机切割机进行加工。

CBN螺杆转子砂轮采用本文提出的原理进行有效性和正确性的验证。

电镀CBN砂轮对螺杆转子进行加工，同时进行机械加工实验。

在实验中获得的数据达到GB10095-88五级认证。

关键词： CBN砂轮精密磨削螺杆转子砂轮外形修整专业术语目录：P 螺杆转子的参数H 螺杆转子的直径Σ砂轮和转子的安装角度Au 砂轮和转子的中心距8 螺旋转子接触点的旋转角x1, y1, z1:转子在σ系统中的位置x, y, z: 砂轮端面的位置x u ,y u ,z u: x, x y z轴的法向量n x ,ny,nz:X Y Z轴的端面法向量n u , nu, nu:砂轮的角速度的矢量:砂轮模块的角速度wu:螺旋转子的角速度w1螺旋转子模块的角速度转子接触点的角速度转子表面接触点的初始速度砂轮表面接触点的角速度砂轮表面接触点的初始速度l砂轮的理论半径砂轮轴的理想位置砂轮表面的修改半径砂轮轴的修改位置砂轮表面的法向量1.引言螺旋转子是螺杆压缩机、螺钉、碎纸机以及螺杆泵的关键部分。

转子的加工精度决定了机械性能。

一般来说，铣刀用于加工螺旋转子。

许多研究者，如肖等人[ 1 ]和姚等人[ 2 ]，对用铣刀加工螺旋转子做了大量的工作。

该方法可以提高加工效率。

然而，加工精度低和表面粗糙度不高是其主要缺点。

随着高新技术的发展，一些新的加工技术被用来制造螺旋转子。

压缩机专业英语中英对照文本

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 plate。

单螺杆机床论文中英文对照资料外文翻译文献

中英文对照资料外文翻译文献Dedicated to the single screw compressor machine updated the IntroductionAbstract: This paper describes four areas from the existingsingle-screw machine layout and structure, and put out the advantages and disadvantages of the list, because of the compressor plant single-screw machine tools and machine tool external Security information, the above introduction there is inevitably one-sided and wrong, and are therefore single-screw compressor for the production of reference works.First, introduce the layout of machine toolsDecide the size of the compressor displacement of the stars round, screw diameter, mesh size and the size of the center distance, so different in diameter screw, machine tool spindle and the rotary center are also different. To meet the processing of different diameter screw, single screw Currently the layout of machine tools in general there are several options.The first is: machine tool rotary tool spindle center and the center distance for the fixedMachine tool rotary tool spindle center and the center distance for the fixed, can not adjust the center distance. Processing of several of the screw diameter on the center distance required several different specifications of the machine.Advantages: simple structure of the machine.Disadvantage: each machine can only process a specification of the screw, when the market on a certain specification requirements when the screw compressor, resulting in a machine, other machine idle.The second: the machine tool spindle box for rotaryProcessing screw machine according to the size of the diameter at the processing before a point of rotating spindle box. Spindle box that the machine can turn on a machine at the above-mentioned article on the use of the improvements, with the first structure of a machine tool is basically the same.Advantages: the structure of machine tool easy to adapt to a variety of specifications of the processing screw.One disadvantage: after the rotating spindle box and the tool spindle turning center line distance between the center line of accurate measurement difficult.2 disadvantage: after the rotating spindle spindle box and the front surface of the rotary cutter centerline distance between the reduction of the larger diameter of the screw processing is limited. The third: the machine tool spindle box for horizontal mobileBox at the bottom of the spindle and the base there is arranged between the rectangular sliding rail, spindle box perpendicular to the direction of movement of spindle centerline and perpendicular to the centerline of the tool rotation. Through the power of the spindle box spline shaft to the base of the tool feed mechanism. Screw diameter, according to the size of the processing in the processing of the previous round by hand to the body put into the screw spindle box moved to the appropriate location, and then screw the spindle box on a fixed base. Spindle box available from the mobile Grating detection, position error ± 0.005mm. Horizontal spindle box can be used as a mobile machine can process diameter φ95 ~ φ385mm any kind between the screw specifications.Φ95 ~ φ385mm processing because of the diameter of the screw, causing the front surface and the tool spindle rotation the distance between the center line of the margin is too large, the actual application in the design specifications of the machine into two, aφ95 ~ φ205mm machine screw diameter Another φ180 ~ φ385mm machine screw diameter.Advantages: a variety of tools to adapt to the specifications of the processing screw, each screw specifications need not be provided with the appropriate machine tools.Disadvantage: the structure of machine tools and machine tool assembly of the two kinds of more complex machine tools, machine tools than the cost of two kinds of machine tools before the high. Second, introduce the structure of machine tool spindleThe level of machine tool spindle box on the main axis and the base of the vertical axis determines the degree of precision was the precision screw machining, at the same time screw compressor at a speed of thousands of high-speed rotary switch, the accuracy of the screw will be less so that the compressor have a fever, vibration, low efficiency, such as wear and tear situation quickly.Currently available single-screw machine spindle structure of the program has the following two.The first is: bearing radial clearance is not adjustable spindle structureBefore spindle bearing out the use of one pairs of cylindrical roller bearings and thrust ball bearing combination of both, the main useof double row cylindrical roller bearings under radial cutting force, the use of two ball bearings to bear axial thrust cutting force.After the general adoption of the spindle bearings out one pairs of cylindrical roller bearings or a ball bearing to the heart.Main advantages of this structure: the main axis of the processing and assembly of simple, low cost.One disadvantage: because the main axis of the radial bearing clearance can not be adjusted so poor precision spindle. Although the use of bearings and shaft diameter fit to eliminate the radial bearing clearance, but each bearing diameter and radial clearance is not a fixed value, so it is difficult to design and processing to the quasi-axial-radial and bearings with bore tolerances.2 disadvantage: it is very difficult to buy in the market of domestically produced or imported, C, D or P4, P5 class thrust ball bearings, machine tool manufacturing plant commonly used alternative to the use of ordinary class bearings, which also affected the accuracy of the enhance spindle.Bearing radial clearance adjustable spindle structure do not apply to the general accuracy of the general machine tools, does not apply to require a higher accuracy of the spindle of machine tools. The second: the radial bearing clearance adjustable spindle structureBefore the adoption of a spindle bearing P4 class of double row tapered hole cylindrical roller bearings and a P4-class double row ball bearing thrust to the combination of heart. The use of the spindle hole of the double row tapered cylindrical roller bearings under radial cutting force, the use of double row ball bearing thrust to the heart to bear part of the axial and radial cutting force cutting force.Spindle bearings generally used after a P5 class of double row tapered hole cylindrical roller bearings.Double row tapered hole cylindrical roller bearings with inner ring and shaft are tapered 1:12, bearing lock nut with a round led a bearing in the axial displacement of the inner ring bearings and expansion, to reduce or eliminate Bearing radial clearance purposes.Main structure of such advantages: high precision spindle. At the front spindle diameter φ230mm noodle on the end measuring spindle Beat value of 0.010mm. Φ230mm cylindric al spindle at the front end on the radial axis measurement value of Beat 0.005mm. The second structure of the spindle of a precision spindle accuracy than the first about 50% improve.Main disadvantage of this structure:The principal axis of the more complicated process, the spindle assembly also has the experience necessary to make the workers to operate the spindle achieve the desired numerical accuracy. Third, the depth of the tool feed controlRequired different processing screw diameter spiral groove depth is also different from the depth of the spiral groove mm from dozens to more than 100 millimeters range around the tool into the institutions required to feed the thousands of ring rotation in order to achieve a screw machining .Feed because of the tool in the tool rotating at the same time achieve motion feed, so on a number of general machine tools used in mechanical, electrical control method of depth of cut does not apply to single-screw machine.Single screw machine tools give agencies into the following different methods can be feed to control the depth of purpose.The first is: friction clutch and electrical switches to control the depth of the tool feedIts principle is to control depth of cut increases the tool cutter feed mechanism increases the load torque so that the tool feeding mechanism of the friction transmission chain slipping clutch, a mechanical linkage concurrent silent trigger electrical switches,optical signal prompted operator, when manual operator to disconnect the tool into the power sector.The advantages of this control method are: the control method is simple and spare parts processing and operational power from the impact of a sudden.Disadvantage are: processing of different diameter screw to adjust the clutch friction discs pressed the preload spring.Material because of the density of each screw, and the hardness of the existence of subtle differences in the degree of cutting tools sharp differences exist, thus the accuracy of this control method was not too accurate, may lead to screw spiral groove depth tolerance is too large.The second: use of an electromagnetic clutch, encoder control tool into the mix to the depth ofTool feed system, equipped with electromagnetic clutch and a tool for detecting the number of rotating ring gear and a gun encoder. It is a tool of control principle hand screw surface encoder to start counting switch, then start counting counting device, when the rotary tool to pre-set number of laps when the cutting depth is reached, the electromagnetic clutch automatic off open to the power tool into the concurrent silent, optical signal parts prompted the operator has finished processing.The detection device through the digital display shows the number of feed circles or feed. Torn off and the electromagnetic clutch, the tool does not only into the rotation with the vertical shaft to the sport.The advantages of this control method are: the depth of the spiral groove screw tolerance control more accurate, because of several significant table shows the depth of processing, or want a few laps and the depth of processing or circle the number of operations is also very intuitive and user-friendly.Disadvantage are: electrical control of machine tools at the same time more complex parts of this control method at the processing plant, if a sudden power failure, the prior data set will be lost.If you add in the electrical control of the battery to power at the early-dimensional detection devices to maintain the job, the problem can be resolved.Four, the control gear drive spaceSingle screw machine screw in the processing, due to the spiral groove in the rotary tool and the workpiece rotation to complete the synthesis process. Just cut into the workpiece when the tool in the tangential direction of rotation has been going on a greater resistance knife, cutting tool at the workpiece to be cut when therole of the spiral groove, the tool in the tangential direction of rotation has been going up against a smaller knife and even by the spiral groove thrust workpiece.Because there is a box-hole processing machine tool, gear and other processing error, the tool axis of rotation of the drive space is too large, large amount of so-called open.Detect drive way too much space is a fixed power input shaft and output shaft rotation shaking, in the case of the transmission structure of conventional design and manufacture of machine tools, the transmission output shaft angle space at more than ten degrees to the dozens of degrees. Transmission gap caused by too large spiral screw groove surface then there is obvious marks, thus affecting the machining accuracy of the screw.Upon completion of the assembly machine tool axis of rotation of the drive space is too large, in fact are subject to various errors gear, creating a backlash of the gear is too large.Machine tools in the mechanical transmission gear are used regardless of the accuracy of a few of the class, the designers take into account the gear manufacturing error, processing error box center distance, temperature, lubricating oil film thickness, the assembly error and other factors, machine design must ensure thattransmission gear A certain amount of backlash, backlash decide the size of the gear tooth thickness tolerance size.Single-screw machine has the Main Drive from other machine tool structure specificity. In order to reduce transmission or reasonable gap single-screw machine tools currently used by the following two ways.The first is: the installation at the output shaft brakeTool at the output shaft rotating the location of cylindrical symmetry with radial brake, brake stand up to the tool front-end of the cylindrical rotary output shaft, brake for spring preload.The working principle of the brake is generated by the friction brake to increase the output shaft damping, reducing the sensitivity of the rotation axis.Are: brake and easy does not change the structure of the original machine tool structure, the method of indirect reduction to achieve the purpose of drive space, in practical applications there is a certain effect.One disadvantage: the pre-spring brake tool because of the cylindrical output shaft to exert a greater radial force, in fact increases the load machine torque, resulting in increased motor power at the same time gears, bearings to accelerate wear and tear.Disadvantage 2: pre-spring brake because of the output shaft of the cylindrical tool to exert a greater radial force on the possible geometry of the tool output shaft a negative impact on accuracy.Conclusion: This article describes four areas from existingsingle-screw machine layout and structure, and put out the advantages and disadvantages of the list, because of the compressor plant single-screw machine tools and machine tool external Security information, the above introduction there is inevitably one-sided and wrong, and are therefore single-screw compressor for the production of reference works.对压缩机单螺杆专用加工机床的介绍更新时间摘要：本文从四个方面介绍了国内现有单螺杆加工机床的布局和结构，并把优缺点一一列举出来，由于压缩机生产厂的单螺杆加工机床和机床资料对外保密，以上介绍难免有片面、不妥之处，因此仅供单螺杆压缩机生产厂参考。

外文翻译---螺杆式空气压缩机

中英文翻译The screw air compressorScrew air compressor is the injection of single stage double screw compressor, divided into single screw air compressor and screw air compressor, efficient use of belt drive, drive the host rotation of compressed air, through the oil for cooling the compressed air in the host, the host of air and oil discharge of the mixed gas after coarse, fine two channel separation, the compressed oil separated out in the air, finally get the clean compressed air. Cooler for cooling the compressed air and oil. With reliable performance, small vibration, low noise, convenient operation, few wearing parts, high efficiency is the biggest advantage of. Screw compressor is a new type compressor, air is compressed alveolar ridge volume change installed on the casing are mutually parallel rotor meshing and achieve. The rotor side in the case and it with precision internal rotation so that gas rotor slot between continuously cyclical changes in volume and along the rotor axis, by the suction side to the discharge side suction, compression, exhaust, complete three working process.Twin screw compressor is a double volume type rotary compressor, which is the main (Yang) side (Yin) with two rotor, composed of meshing pair, vice principal of rotor profile of external components of the element volume closed with the inner wall of the casing, shortage of twin screw is 40 cubic meters and above machine type to add gear, increase power consumption and prone to head lock; while the worm compressor is a uniaxial volume type rotary compressor, the meshing pair is composed of a worm wheel and two symmetrical layout, wall by the worm screw groove and the star wheel tooth surface and the casing form element volume closed, but the problem is the Star blade material has yet to be improved.The main performance parameters of screw type air compressor power, volume flow rate, pressure, exit temperature, discharge pressure and speed etc.. And in the design of screw air compressor, rotor of a pair of mutually meshing is a very important parameter. Because the performance of the compressor is closely related with the. The Yin and Yang of screw compressor rotor can be regarded as a pair of mutually meshed helical gear, therefore, the Yin and Yang of screw compressor rotor profile, but also to meet the meshing law. Rotor tooth surface and the vertical axis of the rotor section called rotor profile. The rotor profile of screw compressor is divided into symmetric and asymmetric line profile, and unilateral and bilateral type line profile, tooth top center line on both sides of the line are identical, called the symmetric line. On the contrary, the tooth top center line on both sides of the line is not at the same time, called asymmetric line. Only one side has a line in the internal or external rotor pitch circle, called single line. Pitch circle, are stylish lines, called bilateral profile. Generally speaking, the development and design of new line directly affect the performance of screw air compressor, design of screw air compressor performance also depends on the type of line.Screw compressor is constantly engaged on the rotor output compressed gas,therefore the spindle speed variation, volume flow rate, the exhaust pressure of the compressor will be affected, so the spindle speed is a major factor affecting the performance of screw compressor. When the exhaust pressure increases, compressor power consumption also increases, power increases, the decrease of economic benefit, so it has significant effect on energy consumption and exhaust pressure of the compressor. At the same time, some test results show that the external environment temperature will influence the performance of screw compressor. China in different seasons and different regions of the temperature difference, air temperature and ambient temperature of compressor is different, this parameter will directly affect the performance of screw compressor. Therefore, the factors that influence the performance of screw compressor were analyzed, will have the very big help to use screw compressor.Screw compressor body is divided into two kinds, one kind is the belt transmission type, another kind is direct drive. Compressor belt transmission type which is suitable for 22KW power, is composed of 2 manufacturing according to the speed ratio of the belt pulley power via the belt transmission; direct drive type is the 1 coupling electric source and host together, worm compressor is directly driven by the worm rotation, while the double screw compressor is required add a gear in order to improve the speed of rotor.The working process of Twin Screw Compressors: motor through a coupling, gear or belt driven by two main rotor, rotor meshing with each other, the main rotor is directly driven by the auxiliary rotor to rotate together, inhaled air in relatively under the action of negative pressure, the peak, and the tooth groove under the agreement, the gas is transported compression, when the rotor meshing surface to and the casing exhaust port, the compressed gas discharge.The working process of screw compressor: motor with coupling or belt to power the worm shaft, driven by the worm gear worm star relative move in grooves, sealing element volume change, gas, transportation compression, when the design pressure, exhaust from the host shell triangle on the left and right symmetrical outlet to the oil and gas separator in the.The host for the housing of screw compressor is provided with a spray hole, depending on the pressure difference, in the process of oil spray to the compression chamber to cool the gas, compressed, the sealing parts gap, and to vibration, noise and lubricating effectInhalation of dust in the air was blocking filter, in order to avoid premature wear and oil separator of compressor to be blocked, usually run 1000 hours or after a year, to replace the filter, dust, replace the time interval to shorten. Filter repair must stop, reduce stop time, recommended for the last new or cleaned standby filter.Clean filter steps are as follows:The 1 to the two end surface of a plane turns tap filter, with most heavy and dry sand.2 with less than 0.28MPa of dry air and air blowing along the opposite direction, the nozzle and the folded paper less distance of 25 mm, and along the length direction, bottom blowing.Grease 3 element, should be dissolved with no foam detergent in warm water wash, the warm water immersion for 15 minutes at least element, and use clean water hose with wash, do not use heating method to accelerate drying, a filter can be washed 5 times, and then discarded reusable.The 4 filter is put in a check, such as thinning, pinhole or damaged should be abandoned.CoolerThe cooler tube, outer surface should pay special attention to the decision to keep clean, otherwise it will reduce the cooling effect, should be working conditions, regular cleaning.Gas tank / oil and gas separatorGas tank / oil separator according to the pressure vessel manufacturing standards and acceptance, not arbitrarily modify.Safety valveInstalled in the tank / oil and gas separator safety valve inspection at least once a year, adjust the safety valve should have professional and responsible, at least every three months to marathon a lever again, so that the valve opening and closing time, a safety valve to work properly.Test steps are as follows:1 close the gas supply valve;2 connected to water;The 3 starting unit;4 observe the working pressure, slowly clockwise pressure regulating bolt, when the pressure reaches the set value, the safety valve does not open or up to the specified value before the open, it must be adjusted.The adjustment procedure is as follows:1 remove the cap and seal;The 2 valve opening soon, loosen locknut and a positioning bolt half ring, valve open too late to loosenThe nut about a circle and release the positioning bolts half circle.3 repeat detection step, the safety valve set pressure value, still can not open, again adjustment.Motor overload relayThe relay normally open contact, must be closed, when the current exceeds the rated value, cut off the motor power supply.The maintenance of screw compressor relative to the piston compressor is to lower the probability of a failure of many, but if use improper maintenance, the advantages of screw compressor is difficult to play. Many users began to use piston compressor screw air compressor, due to the maintenance of screw compressor does not understand, leading to frequent failure, causing the conflict between users and enterprises. So before using screw compressor to his use and maintenance instructions carefully read.Installation place, on the machine's air ventilation, pressure station power supply cable and air switch specifications, water pressure and flow, exhaust pipe sizes haveguidance. The instructions also principle and structure of the machine are described in detail, there are certain basic knowledge of mechanical and electrical personnel through carefully reading, can make correct judgement and treatment for common problems. There is a consumable part of screw compressor to the regular replacement of main lubricating oil, oil filter, air filter, oil filter, notable is replaced in the environment of different parts of the frequency also has the difference, so the screw compressor to daily use and inspection record.Transmission gear and belt driveIn the transmission system of the air compressor, generally can be divided into direct transmission and belt drive, long-term since, two kinds of transmission way merits has been one of the focus of the debate. Direct drive screw air compressor refers to the spindle motor through a coupling and a gear to drive the rotor, direct drive which in fact is not the real meaning of. Direct drive on real significance is directly connected to motor and rotor (coaxial) and the speed. This is obviously a little. So that the direct drive no energy consumption point of view is wrong.Another way of drive for belt drive, the drive through the belt wheels of different diameters is allowed to change the rotating speed of the rotor.Belt drive system discussed below is a representative shall meet the following conditions of the automation system of the latest technology:Belt 1, each operating state of tension reached the optimal value2, by avoiding the tension, greatly prolong the working life of the belt, while reducing the motor and rotor bearing load;3, always make sure that the correct belt wheel is connected;4, replacement of the belt is quite easy and fast, and do not need to adjust the original setting;5, the belt drive system safe and trouble-free operation.The comparison of transmission1 efficiency.Gear transmission efficiency can reach 98%~99% excellent, belt drive design excellent under normal working conditions and efficiency reached 99%. The difference does not depend on the choice of drive mode, and depending on the manufacturer's design and manufacturing level.2 no-load power consumptionFor direct gear driving mode, no-load pressure generally maintained at more than 2.5bar, some even as high as 4bar, to ensure that the gear lubrication.The belt drive mode, theory of no-load pressure can be zero, because the rotor into the oil to lubricate the rotor and bearing. General for the sake of safety, the pressure maintained at about 0.5bar.Taking gear compressor a 160kW for example, 8000 hours per year, of which 15% (1200 hours) time for no-load, the machine every year than the same power air compressor belt drive more consumption of electricity (28800kwh no-load assuming two machine pressure is 2bar, the difference of the energy consumption of about 15%.), over the long term, this will be no small cost.3 oil lossThe actual users experienced all know, oil loss under the condition of the first victims will be the gear box. Belt drive system is not the existence of such security problems.4 according to user requirements design working pressureThe working pressure and manufacturers usually user requirements of the standard model of the pressure is not completely consistent. For example, the user's requirements in accordance with the pressure of 10bar, after processing equipment, pipe length and different requirements of sealing, the pressure of air compressor may be 11 or 11.5bar. In this case, the general will install a rated pressure of 13bar e air compressor and in the field to the outlet pressure set for the required pressure. The exhaust volume will basically remain unchanged, because the ultimate working pressure is reduced, but did not increase the speed of rotor.The belt transmission design manufacturers on behalf of the modern technology simply change the belt pulley diameter and the pressure of work designed to be completely consistent with the requirements of users, so that users with a motor with a power which can get more air volume. For the wheel transmission, is not so easy. The installed 5 air compressor pressure changeSometimes because of technological conditions for the production of the user changes, the design pressure of air compressor of the original purchase may be too high or too low, will change, but for the compressor gear transmission, this work is very difficult and expensive, and for the belt drive type air compressor is be an easy job to do, only changing the belt pulley you can.6 installing the new bearingWhen the rotor bearing needs to be replaced, the compressor gear transmission, gear box and gear box bearing at the same time the cost of overhaul, let users to accept. The belt drive air compressor, simply does not exist in this topic.7 replacement sealAny screw compressors are used in an annular seal, to a certain life need to be replaced. For gear drive type air compressor, must first separate motor, coupling, access to the shaft seal, making this work is time-consuming, thus increasing maintenance costs. The belt drive compressors, only need to remove the belt wheel, easier.8 motor rotor bearing damage orFor gear drive air compressor, when the motor or the rotor bearing damage, often hurt is important parts of the direct and indirect double damage. The belt drive air compressor does not exist.9 the structure of noiseFor gear drive air compressor, the motor is connected with the rigidity of the rotor, rotor vibration compression chamber can be transmitted to the gearbox and motor bearing, which not only increases the wear of motor bearing, while increasing the noise of the machines.螺杆式空气压缩机螺杆式空气压缩机是喷油单级双螺杆压缩机，分为单螺杆式空气压缩机及双螺杆式空气压缩机，采用高效带轮传动，带动主机转动进行空气压缩，通过喷油对主机内的压缩空气进行冷却，主机排出的空气和油混合气体经过粗、精两道分离，将压缩空气中的油分离出来，最后得到洁净的压缩空气。

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。

【机械类文献翻译】对螺杆压缩机的三维分析开发

英文原文Applications4.1 IntroductionThis chapter demonstrates the scope of the method developed for the three-dimensional analysis of a screw compressor. The CFD package used in this case was COMET developed by ICCM GmbH Hamburg, today a part of CD-Adapco. The analysis of the flow and performance characteristics of a number of types of screw machines is performed to demonstrate a variety of parameters used for grid generation and calculation.The first example is concerned with a dry air screw compressor. A common compressor casing is used with two alternative pairs of rotors. The rotors have identical overall geometric properties but different lobe profiles. The application of the adaptation technique enables convenient grid generation for geometrically different rotors. The results obtained by three dimensional modelling are compared with those derived from a one-dimensional model, previously verified by comparison with experimental data..The relative advantages of each rotor profile are demonstrated.The second example shows the application of three dimensional flow analysis to the simulation of an oil injected air compressor. The results, thus obtained, are compared with test results obtained by the authors from a compressor and test rig, designed and built at City University. They are presented in the form of both integral parameters and a p-∂indicator diagram. Calculations based on the assumptions of the laminar flow are compared to those of turbulent flow. The effect of grid size on the results is also considered and shown here.The third example gives the analysis of an oil injected compressor in an ammonia refrigeration plant.This utilises the real fluid property subroutines in the process calculations and demonstrates the blow hole area and the leakage flow through the compressor clearances.The fourth example presents two cases, one of a dry screw compressor to show the influence of thermal expansion of the rotor on screw compressor performance and one of a high pressure oil-flooded screw compressor to show the influence of high pressure loads upon the compressor performance.4.2 Flow in a Dry Screw CompressorDry screw compressors are commonly used to produce pressurised air, free of any oil. A typical example of such a machine, similar in configuration to the compressor modelled, is shown in Figure 4-1. This is a single stage machine with 4 male and 6 female rotor lobes. The male and female rotor outer diameters are 142.380 mm and 135.820 mm respectively, while their centre lines are 108.4 mm apart. The rotor length to main diameter ratio l/d=1.77. Thus, the rotor length∂=248.40 is driven at a speed of 6000 rpm by an is 252.0 mm. The male rotor with wrap angleωelectric motor through a gearbox. The male and female rotors are synchronised through timinggears with the same ratio as that of the compressor rotor lobes i.e. 1.5. The female rotor speed is therefore 4000 rpm. The male rotor tip speed is then 44.7m/s, which is a relatively low value for a dry air compressor. The working chamber is sealed from its bearings by a combination of lip and labyrinth seals.Each rotor is supported by one radial and one axial bearing, on the discharge end, and one radial bearing on the suction end of the compressor. The bearings are loaded by a high frequency force, which varies due to the pressure change within the working chamber. Both radial and axial forces, as well as the torque change with a frequency of 4 times the rotational speed. This corresponds to 400Hz and coincides with the number of working cycles that occur within the compressor per unit time.Figure 4-1 Cross section of a dry screw compressorThe compressor takes in air from the atmosphere and discharges it to a receiver at a constant output pressure of 3 bar. Although the pressure rise is moderate, leakage through radial gaps of 150 m is substantial. In many studies and modelling ,procedures, volumetric losses are assumed to be a linear function of the cross sectional area and the square root of pressure difference, assuming that the interlobe clearance is kept more or less constant by the synchronising gears. The leakage through the clearances is then proportional to the clearance gap and the length of the leakage line. However, a large clearance gap is needed to prevent contact with the housing caused by rotor deformation due to the pressure and temperature changes within the working chamber. Hence, the only way to reduce leakage is to minimise the length of the sealing line. This can be achieved by careful design of the screw rotor profile. Although minimising,leakage is an important means of improving a screw compressor efficiency, it is not the only one. Another is to increase the flow area between the lobes and thereby increase the compressor flow capacity, thereby reducing the relative effect of leakage. Modern profile generation methods take these various effects into account by means of optimisation procedures which lead to enlargement of the male rotor interlobes and reduction in the female rotor lobes. Thefemale rotor lobes are thereby strengthened and their deformation thus reduced.To demonstrate the improvements possible from rotor profile optimisation, a three dimensional flow analysis has been carried out for two different rotor profiles within the same compressor casing, as shown in Figure 4-2. Both rotors are of the “N” type and rack generated.Figure 4-2…N‟ Rotors, Case-1 upper, Case-2 lowerCase 1 is an o lder design, similar in shape to SRM “D” rotors. Its features imply that there is a large torque on the female rotor, the sealing line is relatively long and the female lobes are relatively weak.Case 2, shown on the bottom of Figure 4-2, has rotors optimised for operating on dry air. The female rotor is stronger and the male rotor is weaker. This results in higher delivery, a relatively shorter sealing line and less torque on the female rotor. All these features help to improve screw compressor performance.The results of these two analyses are presented in the form of velocity distributions in the planesdefined by cross-sections A-A and B-B, shown in Figure 4-1.In the case of this study, the effect of rotor profile changes on compressor integral performance parameters can be predicted fairly accurately with one-dimensional models, even if some of the detailed assumptions made in such analytical models are inaccurate. Hence the integral results obtained from the three-dimensional analysis are compared with those from a one-dimensional model.4.2.1 Grid Generation for a Dry Screw CompressorIn Case-1, the rotors are mapped with 52 numerical cells along the interlobe on the male rotor and 36 cells along each interlobe on the female rotor in the circumferential direction. This gives 208 and 216 numerical cells respectively in the circumferential direction for the male and female rotors. A total of 6 cells in the radial direction and 97 cells in the axial direction is specified for both rotors. This arrangement results in a numerical mesh with 327090 cells for the entire machine. The cross section for the Case-1 rotors is shown in Figure 4-3. The female rotor is relatively thin and has a large radius on the lobe tip. Therefore, it is more easily mapped than in Case-2 where the tip radius is smaller, as shown in Figure 4-4.Figure 4-3 Cross section through the numerical mesh for Case-1 rotorsThe rotors in Case 2 are mapped with 60 cells along the male rotor lobe and 40 cells along the female lobe, which gives 240 cells along both rotors in the circumferential direction. In the radial direction, the rotors are mapped with 6 cells while 111 cells are selected for mapping along the rotor axis. Thus, the entire working chamber for this compressor has 406570 cells. In this case, different mesh sizes are applied and different criteria are chosen for the boundary adaptation of these rotors. The main adaptation criterion selected for the rotors is the local radius curvature with a grid point ratio of 0.3 to obtain the desired quality of distribution along the rotor boundaries. By this means, the more curved rotors are mapped with only a slight increase in the grid size to obtain a reasonable value of the grid aspect ratio. To obtain a similar grid aspect ratiowithout adaptation, 85 cells would have been required instead of 60 along one interlobe on the female rotor. This would give 510 cells in the circumferential direction on each of rotors. If the number of cells in the radial direction is also increased to be 8 instead of 6 but the number of cells along axis is kept constant, the entire grid would contain more then a million cells which would, in turn, result in a significantly longer calculation time and an increased requirement for computer memory.Figure 4-4 Cross section through the numerical mesh for Case-2 rotors4.2.2 Mathematical Model for a Dry Screw CompressorThe mathematical model used is based on the momentum, energy and mass conservation equations as given in Chapter 2. The equation for space law conservation is calculated in the model in order to obtain cell face velocities caused by the mesh movement. The system of equations is closed by Stoke‟s, Fourier‟s and Fick‟s laws and the equation of state for an ideal gas. This defines all the properties needed for the solution of the governing equations.4.2.3 Comparison of the Two Different Rotor ProfilesThe results obtained for both Case 1 and Case 2 compressors are presented here. To establish the full range of working conditions and to obtain an increase of pressure from 1 to 3 bars between the compressor suction and discharge, 15 time steps were required. A further 25 time steps were then needed to complete the full compressor cycle. Each time step needed about 30 minutes running time on an 800 MHz AMD Athlon processor. The computer memory required was about 400 MB. In Figure 4-5 the velocity vectors in the cross and axial sections are compared. The top diagram is given for Case-1 rotors and the bottom one for Case-2. As may be seen, the Case 2 rotors realised a smoother velocity distribution than the Case 1 rotors. This may have some advantage and could have increased the compressor adiabatic efficiency by reduction in flow drag losses. In both cases, recirculation within the entrapped working chamber occurs as consequence of the drag forces in the air as shown in the figure. On the other hand, different fluid flow patterns can be observed inthe suction port. The velocities within the working chambers and the suction and discharge ports are kept relatively low while the flow through the clearance gaps changes rapidly and easily reaches sonic velocity.Figure 4-5 Velocity field in the compressor cross section for Case1 and Case2 rotorsFigure 4-6 Velocity field in the compressor axial section for Case1 and Case2 rotors These differences are confirmed in the view of the vertical compressor section through the female rotor axis, shown in Figure 4-6. In Case 2, lower velocities are achieved not only in the working chamber but also in the suction and discharge ports. In the suction port, this is significant because of the fluid recirculation which appears at the end of the port. This recirculation causes losses which cannot be recovered later in the compression process. Therefore, many compressors are designed with only an axial port instead of both, radial and axial ports. Such a situation reduces suction dynamic losses caused by recirculation but, on the other hand, increases thevelocity in the suction chamber which in turn decreases efficiency. Some of these problems can beavoided only by the design of screw compressor rotors with larger lobes and a bigger swept volume and a shape which allows the suction process to be completed more easily. However, rotor profile design based on existing one-dimensional procedures neglects flow variations in the ports and hence is inferior for this purpose. In such cases, only a full three dimensional approach such as this, will be effective.中文译文应用4.1简介本章介绍了对螺杆压缩机的三维分析开发的方法的范围。

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。

压缩机专业英语中英对照文本

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 filter 29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 filter 34灭菌滤芯Sterilization filter35露点Dew point36压力露点Pressure dew point37自动排水器Automatic drain valve 38电子排水器Electronic drain valve 39智能排水器Intelligent drain valve 40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 separator 46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 separator 60安全芯safe filter61联轴器flex62卸荷阀unloading valve63最小压力阀Min pressure valve 64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 plate71Motor speed 马达转速zr37vsd(zr45vsd)Delivery air 输送空气压力zr250zr250。

螺杆压缩机毕业设计(含外文翻译)

螺杆压缩机毕业设计（含外文翻译）华东交通大学毕业设计（论文）任务书毕业设计开题报告书螺杆压缩机的介绍双螺杆压缩机属于回转式压缩机，是一种工作容积作旋转运动的容积式气体压缩机械。

气体的压缩是通过容积的变化来实现，而容积的变化又是借压缩机的一个或几个转子在气缸里作旋转运动来达到。

回转式压缩机的工作容积不同于往复式压缩机，它除了周期性地扩大和缩小外，其空间位置也在变更。

回转式压缩机靠容积的变化来实现气体的压缩，这一点与往复式压缩机相同，它们都属于容积式压缩机；回转式压缩机的主要机件（转子）在气缸内作旋转运动，这一点又与速度式压缩机相同。

所以，回转式压缩机同时兼有上述两类机器的特点。

回转式压缩机没有往复运动机构，一般没有气阀，零部件（特别是易损件）少，结构简单、紧凑，因而制造方便，成本低廉；同时，操作简便，维修周期长，易于实现自动化。

回转式压缩机的排气量与排气压力几乎无关，与往复式压缩机一样，具有强制输气的特征。

回转式压缩机运动机件的动力平衡性良好，故压缩机的转数高、基础小。

这一优点，在移动式机器中尤为明显。

回转式压缩机转数高，它可以和高速原动机（如电动机、内燃机、蒸汽轮机等）直接相联。

高转数带来了机组尺寸小、重量轻的优点。

同时，在转子每转一周之内，通常有多次排气过程，所以它输气均匀、压力脉动小，不需设置大容量的储气罐。

回转式压缩机的适应性强，在较大的工况范围内保持高效率。

排气量小时，不像速度式压缩机那样会产生喘振现象。

在某些类型的回转式压缩机（如罗茨鼓风机、螺杆式压缩机）中，运动机件相互之间，以及运动机件与固定机件之间，并不直接接触，在工作容积的周壁上无需润滑，可以保证气体的洁净，做到绝对无油的压送气体（这类机器成为无油回转压缩机）。

同时，由于相对运动的机件之间存在间隙以及没有气阀，故它能压送污浊和带液滴、含粉尘的气体。

但是，回转式压缩机也有它的缺点，这些缺点是：由于转数较高，加之工作容积与吸排气孔口周期性地相通、切断，产生较为强烈的空气动力噪声，其中螺杆式压缩机、罗茨鼓风机尤为突出，若不采取消音措施，即不能被用户所利用。

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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。