喷油螺杆压缩机的流量分析外文文献翻译、中英文翻译、外文翻译
液压机外文翻译文献
液压机外文翻译文献(文档含中英文对照即英文原文和中文翻译)原文：The Analysis of Cavitation Problems in the Axial Piston Pumpshu WangEaton Corporation,14615 Lone Oak Road,Eden Prairie, MN 55344This paper discusses and analyzes the control volume of a piston bore constrained by the valve plate in axial piston pumps. The vacuum within the piston bore caused by the rise volume needs to be compensated by the flow; otherwise, the low pressure may cause the cavitations and aerations. In the research, the valve plate geometry can be optimized by some analytical limitations to prevent the piston pressure below the vapor pressure. The limitations provide the design guide of the timings andoverlap areas between valve plate ports and barrel kidneys to consider the cavitations and aerations. _DOI: 10.1115/1.4002058_ Keywords: cavitation , optimization, valve plate, pressure undershoots1 IntroductionIn hydrostatic machines, cavitations mean that cavities or bubbles form in the hydraulic liquid at the low pressure and collapse at the high pressure region, which causes noise, vibration, and less efficiency.Cavitations are undesirable in the pump since the shock waves formed by collapsed may be strong enough to damage components. The hydraulic fluid will vaporize when its pressure becomes too low or when the temperature is too high. In practice, a number of approaches are mostly used to deal with the problems: (1) raise the liquid level in the tank, (2) pressurize the tank, (3) booster the inlet pressure of the pump,(4) lower the pumping fluid temperature, and (5) design deliberately the pump itself.Many research efforts have been made on cavitation phenomena in hydraulic machine designs. The cavitation is classified into two types in piston pumps: trapping phenomenon related one (which can be prevented by the proper design of the valve plate)and the one observed on the layers after the contraction or enlargement of flow passages (caused by rotating group designs) in Ref. (1). The relationship between the cavitation and the measured cylinder pressure is addressed in this study. Edge and Darling (2) reported an experimental study of the cylinder pressure within an axial piston pump. The inclusion of fluid momentum effects and cavitations within the cylinder bore are predicted at both highspeed and high load conditions. Another study in Ref. (3) provides an overview of hydraulic fluid impacting on the inlet condition and cavitation potential. It indicates thatphysical properties (such as vapor pressure, viscosity, density, and bulk modulus) are vital to properly evaluate the effects on lubrication and cavitation. A homogeneous cavitation model based on the thermodynamic properties of the liquid and steam is used to understand the basic physical phenomena of mass flow reduction and wave motion influences in the hydraulic tools and injection systems (4). Dular et al. (5, 6) developed an expert system for monitoring and control of cavitations in hydraulic machines and investigated the possibility of cavitation erosion by using the computational fluid dynamics (CFD) tools. The erosion effects of cavitations have been measured and validated by a simple single hydrofoil configuration in a cavitation tunnel. It is assumed that the severe erosion is often due to the repeated collapse of the traveling vortex generated by a leading edge cavity in Ref. (7). Then, the cavitation erosion intensity may be scaled by a simple set of flow parameters: theupstream velocity, the Strouhal number, the cavity length, and the pressure. A new cavitation erosion device, called vortex cavitation generator, is introduced to comparatively study various erosion situations (8).More previous research has been concentrated on the valve plate designs, piston, and pump pressure dynamics that can be associated with cavitations in axial piston pumps. The control volume approach and instantaneous flows (leakage) are profoundly studied in Ref. [9]. Berta et al. [10] used the finite volume concept to develop a mathematical model in which the effects of port plate relief grooves have been modeled and the gaseous cavitation is considered in a simplified manner. An improved model is proposed in Ref.[11] and validated by experimental results. The model may analyze the cylinder pressure and flow ripples influenced by port plate and relief groove design. Manring comparedprincipal advantages of various valve plate slots (i.e., theslots with constant, linearly varying, and quadratic varyingareas) in axial piston pumps [12]. Four different numericalmodels are focused on the characteristics of hydraulic fluid,and cavitations are taken into account in different ways toassist the reduction in flow oscillations [13].The experiences of piston pump developments show thatthe optimization of the cavitations/aerations shall includethe following issues: occurring cavitation and air release,pump acoustics caused by the induced noises, maximal amplitudes of pressure fluctuations, rotational torque progression, etc. However, the aim of this study is to modifythe valve plate design to prevent cavitation erosions causedby collapsing steam or air bubbles on the walls of axial pump components. In contrast to literature studies, the researchfocuses on the development of analytical relationshipbetween the valve plate geometrics and cavitations. The optimization method is applied to analyze the pressure undershoots compared with the saturated vapor pressurewithin the piston bore.The appropriate design of instantaneous flow areas betweenthe valve plate and barrel kidney can be decided consequently.2 The Axial Piston Pump and Valve PlateThe typical schematic of the design of the axis piston pumpis shown in Fig. 1. The shaft offset e is designed in this caseto generate stroking containment moments for reducingcost purposes.The variation between the pivot center of the slipper andswash rotating center is shown as a. The swash angle αis the variable that determines the amount of fluid pumped pershaft revolution. In Fig. 1, the n th piston-slipper assembly is located at the angle of nθ. The displacement of the n thpiston-slipper assembly along the x-axis can be written asx n = R tan （α）sin （n θ）+ a sec （α） + e tan (α) (1) where R is the pitch radius of the rotating group. Then, the instantaneous velocity of the n th piston is x˙n = R 2sec ()αsin （n θ）α+ R tan （α）cos （n θ）ω+ R 2sec ()αsin （α）α + e 2sec ()αα (2) where the shaft rotating speed of the pump is ω=d n θ / dt .The valve plate is the most significant device to constraint flow in piston pumps. The geometry of intake/discharge ports on the valve plate and its instantaneous relative positions with respect to barrel kidneys are usually referred to the valve plate timing. The ports of the valve plate overlap with each barrel kidneys to construct a flow area or passage, which confines the fluid dynamics of the pump. In Fig. 2, the timing angles of the discharge and intake ports on the valve plate are listed as (,)T i d δ and (,)B i d δ. The opening angle of the barrel kidney is referred to as ϕ. In some designs, there exists asimultaneous overlap between the barrel kidney and intake/discharge slots at the locations of the top deadcenter (TDC) or bottom dead center (BDC) on the valve plate on which the overlap area appears together referred to as “cross-porting” in the pump design engineering. The cross-porting communicates the discharge and intake ports, which may usually lower the volumetric efficiency. The trapped-volume design is compared with the design of the cross-porting, and it can achieve better efficiency 14]. However, the cross-porting isFig. 1 The typical axis piston pump commonly used to benefit the noise issue and pump stability in practice.3 The Control Volume of a Piston BoreIn the piston pump, the fluid within one piston is embraced by the piston bore, cylinder barrel, slipper,valve plate, and swash plate shown in Fig. 3. There exist some types of slip flow by virtue of relative Fig. 2 Timing of the valve plate motions and clearances between thos e components. Within the control volume of each piston bore, the instantaneous mass is calculated asn M = ρn V (3) where ρ and n V are the instantaneous density and volume such that themass time rate of change can be given asFig. 3 The control volume of the piston boren n n dM dV d V dt dt dtρρ=+ (4)where d n V is the varying of the volume.Based on the conservation equation, the mass rate in the control volume isn n dM q dtρ= (5) where n q is the instantaneous flow rate in and out of onepiston. From the definition of the bulk modulus,n dP d dt dtρρβ= (6) where Pn is the instantaneous pressure within the piston bore. Substituting Eqs. (5) and (6) into Eq. (4) yields(?)n n n n n ndP q dV d V w d βθθ=- (7) where the shaft speed of the pump is n d dt θω=. The instantaneous volume of one piston bore can be calculated by using Eq. (1) asn V = 0V + P A [R tan （α）sin （n θ）+ a sec （α） + e tan (α) ](8) where P A is the piston sectional area and 0V is the volume of each piston, which has zero displacement along the x-axis (when n θ=0, π).The volume rate of change can be calculated at thecertain swash angle, i.e., α =0, such thattan cos n p n ndV A R d αθθ=（）（） (9) in which it is noted that the piston bore volume increases or decreases with respect to the rotating angle of n θ. Substituting Eqs. (8) and (9) into Eq. (7) yields0[tan()cos()] [tan sin sec tan() ]n P n n n p n q A R dP d V A R a e βαθωθαθαα-=-++（）（）（）(10)4 Optimal DesignsTo find the extrema of pressure overshoots and undershoots in the control volume of piston bores, the optimization method can be used in Eq. (10). In a nonlinear function, reaching global maxima and minima is usually the goal of optimization. If the function is continuous on a closed interval, global maxima and minima exist. Furthermore, the global maximum (or minimum) either must be a local maximum (or minimum) in the interior of the domain or must lie on the boundary of the domain. So, the method of finding a global maximum (or minimum) is to detect all the local maxima (or minima) in the interior, evaluate themaxima (or minima) points on the boundary, and select the biggest (or smallest) one. Local maximum or local minimum can be searched by using the first derivative test that the potential extrema of a function f( · ), with derivative ()f ', can solve the equation at the critical points of ()f '=0 [15]. The pressure of control volumes in the piston bore may be found as either a minimum or maximum value as dP/ dt=0. Thus, letting the left side of Eq. (10) be equal to zero yields tan()cos()0n p n q A R ωαθ-= (11) In a piston bore, the quantity of n q offsets the volume varying and then decreases the overshoots and undershoots of the piston pressure. In this study, the most interesting are undershoots of the pressure, which may fall below the vapor pressure or gas desorption pressure to cause cavitations. The term oftan()cos()p n A R ωαθ in Eq. (11) has the positive value in the range of intake ports (22ππθ-≤≤), shown in Fig. 2, which means that the piston volume arises. Therefore, the piston needs the sufficient flow in; otherwise, the pressure may drop.In the piston, the flow of n q may get through in a few scenarios shown in Fig. 3: (I) the clearance between the valve plate and cylinder barrel, (II) the clearance between the cylinder bore and piston, (III) theclearance between the piston and slipper, (IV) the clearance between the slipper and swash plate, and (V) theoverlapping area between the barrel kidney and valve plate ports. As pumps operate stably, the flows in the as laminar flows, which can be calculated as [16]312IV k k Ln i I k h q p L ωμ==∑ (12)where k h is the height of the clearance, k L is the passagelength,scenarios I –IV mostly have low Reynolds numbers and can be regardedk ω is the width of the clearance (note that in the scenario II,k ω =2π· r, in which r is the piston radius), and p is the pressure drop defined in the intake ports as p =c p -n p (13) where c p is the case pressure of the pump. The fluid films through the above clearances were extensively investigated in previous research. The effects of the main related dimensions of pump and the operating conditions on the film are numerically clarified in Refs. [17,18]. The dynamic behavior of slipper pads and the clearance between the slipper and swash plate can be referred to Refs. [19,20]. Manring et al. [21,22]investigated the flow rate and load carrying capacity of the slipper bearing in theoretical and experimental methods under different deformation conditions. A simulation tool called CASPAR is used to estimate the nonisothermal gap flow between the cylinder barrel and the valve plate by Huang and Ivantysynova [23]. The simulation program also considers the surface deformations to predict gap heights, frictions, etc., between the piston and barrel and between the swash plate and slipper. All these clearance geometrics in Eq. (12) are nonlinear and operation based, which is a complicated issue. In this study, the experimental measurements of the gap flows are preferred. If it is not possible, the worst cases of the geometrics or tolerances with empirical adjustments may be used to consider the cavitation issue, i.e., minimum gap flows.For scenario V, the flow is mostly in high velocity and can be described by using the turbulent orifice equation as((Tn d i d d q c A c A θθ= (14) where Pi and Pd are the intake and discharge pressure of the pump and ()i A θ and ()d A θ are the instantaneousoverlap area between barrel kidneys and inlet/discharge ports of the valve plate individually.The areas are nonlinear functions of the rotating angle, which is defined by the geometrics of the barrel kidney, valve plate ports, silencing grooves, decompression holes, and so forth. Combining Eqs. (11) –(14), the area can beobtained as3()K IV A θ==(15)where ()A θ is the total overlap area of ()A θ=()()i d A A θλθ+, andλis defined as=In the piston bore, the pressure variesfrom low to high while passing over the intake and discharge ports of the valve plates. It is possible that the instantaneous pressure achieves extremely low values during the intakearea( 22ππθ-≤≤ shown in Fig. 2) that may be located below the vapor pressure vp p , i.e., n vp p p ≤;then cavitations canhappen. To prevent the phenomena, the total overlap area of ()A θmight be designed to be satisfied with30()K IV A θ=≥(16)where 0()A θ is the minimum area of 0()A θ=0()()i d A A θλθ+and0λis a constant that is0λ=evaporates into a gaseous form. The vapor pressure of any substance increases nonlinearly with temperature according to the Clausius –Clapeyron relation. With the incremental increase in temperature, the vapor pressure becomes sufficient to overcome particle attraction and make the liquid form bubbles inside the substance. For pure components, the vapor pressure can be determined by the temperature using the Antoine equation as /()10A B C T --, where T is the temperature, and A, B, and C are constants [24]. As a piston traverse the intake port, the pressure varies dependent on the cosine function in Eq. (10). It is noted that there are some typical positions of the piston with respect tothe intake port, the beginning and ending of overlap, i.e., TDC and BDC (/2,/2θππ=- ) and the zero displacement position (θ =0). The two situations will be discussed as follows: (1) When /2,/2θππ=-, it is not always necessary to maintain the overlap area of 0()A θ because slip flows may provide filling up for the vacuum. From Eq. (16), letting 0()A θ=0,the timing angles at the TDC and BDC may be designed as31cos ()tan()122IV c vpk k i I P k p p h A r L ωϕδωαμ--≤+∑ (17) in which the open angle of the barrel kidney is . There is no cross-porting flow with the timing in the intake port.(2) When θ =0, the function of cos θ has the maximum value, which can provide another limitation of the overlap area to prevent the low pressure undershoots suchthat 30(0)K IV A =≥ (18)where 0(0)A is the minimum overlap area of 0(0)(0)i A A .To prevent the low piston pressure building bubbles, the vapor pressure is considered as the lower limitation for the pressure settings in Eq. (16). The overall of overlap areas then can be derived to have a design limitation. The limitation is determined by the leakage conditions, vapor pressure, rotating speed, etc. It indicates that the higher the pumping speed, the more severe cavitation may happen, and then the designs need more overlap area to let flow in the piston bore. On the other side, the low vapor pressure of the hydraulic fluid is preferred to reduce the opportunities to reach the cavitation conditions. As a result, only the vapor pressure of the pure fluid is considered in Eqs. (16)–(18). In fact, air release starts in the higher pressure than the pure cavitation process mainly in turbulent shear layers, which occur in scenario V. Therefore, the vapor pressure might be adjusted to design the overlap area by Eq. (16) if there exists substantial trapped and dissolved air in the fluid. The laminar leakages through the clearances aforementioned are a tradeoff in the design. It is demonstrated that the more leakage from the pump case to piston may relieve cavitation problems.However, the more leakage may degrade the pump efficiency in the discharge ports. In some design cases, the maximum timing angles can be determined by Eq. (17)to not have both simultaneous overlapping and highly low pressure at the TDC and BDC. While the piston rotates to have the zero displacement, the minimum overlap area can be determined by Eq. 18 , which may assist the piston not to have the large pressure undershoots during flow intake.6 Conclusions The valve plate design is a critical issue in addressing thecavitation or aeration phenomena in the piston pump. This study uses the control volume method to analyze the flow, pressure, and leakages within one piston bore related to the valve plate timings. If the overlap area developed by barrel kidneys and valve plate ports is not properly designed, no sufficient flow replenishes the rise volume by the rotating movement. Therefore, the piston pressure may drop below the saturated vapor pressure of the liquid and air ingress to form the vapor bubbles. To control the damaging cavitations, the optimization approach is used to detect the lowest pressure constricted by valve plate timings. The analytical limitation of the overlap area needs to be satisfied to remain the pressure to not have large undershoots so that the system can be largely enhanced on cavitation/aeration issues. In this study, the dynamics of the piston control volume is developed by using several assumptions such as constant discharge coefficients and laminar leakages. The discharge coefficient is practically nonlinear based on the geometrics, flow number, etc. Leakage clearances of the control volume may not keep the constant height and width as well in practice due to vibrations and dynamical ripples. All these issues are complicated and very empirical and need further consideration in the future. The results presented in this paper can be more accurate in estimating the cavitations with these extensive studies. Nomenclature0(),()A A θθ= the total overlap area between valve plate ports and barrel kidneys 2()mm Ap = piston section area 2()mm A, B, C= constants A= offset between the piston-slipper joint and surface of the swash plate 2()mmd C = orifice discharge coefficiente= offset between the swash plate pivot and the shaft centerline of the pump 2()mmk h = the height of the clearance 2()mmk L = the passage length of the clearance 2()mm M= mass of the fluid within a single piston (kg) N= number of pistons n = piston and slipper counter,p p = fluid pressure and pressure drop (bar) Pc= the case pressure of the pump (bar) Pd= pump discharge pressure (bar) Pi = pump intake pressure (bar) Pn = fluid pressure within the nth piston bore (bar) Pvp = the vapor pressure of the hydraulic fluid(bar) qn, qLn, qTn = the instantaneous flow rate of each piston (l/min) R = piston pitch radius 2()mmr = piston radius (mm )t =time (s )V = volume 3()mmwk = the width of the clearance (mm )x ,x˙= piston displacement and velocity along the shaft axis (m, m/s )x y z --=Cartesian coordinates with an origin on the shaft centerlinex y z '''--= Cartesian coordinates with an origin on swash plate pivot,αα=swash plate angle and velocity (rad, rad/s )β= fluid bulk modulus (bar ),B T δδ= timing angle of valve plates at the BDC and TDC (rad ) ϕ = the open angle of the barrel kidney (rad )ρ= fluid density (kg /m3),θω = angular position and velocity of the rotating kit (rad, rad/s )μ =absolute viscosity (Cp ),λλ=coefficients related to the pressure drop翻译：在轴向柱塞泵气蚀问题的分析本论文讨论和分析了一个柱塞孔与配流盘限制在轴向柱塞泵的控制量设计。
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。
螺杆真空泵的性能预测的研究外文文献翻译、中英文翻译
XX设计(XX)外文资料翻译系别：专业：班级：姓名：学号：外文出处：万方数据库附件： 1. 原文； 2. 译文20XX年06月Study on the performance prediction of screw vacuum pumpAbstractPumping characteristics of the screw vacuum pump were investigated. The aim of this study was to establish a method of the performance prediction and a way to design the pump that satisfies specific requirements. The performance was analysed by the balance among geometrical pumping speed, net throughput and leaks. The leaks flow through clearances between a screw rotor and a stator, and clearances between two meshing rotors. These leaks were estimated with the results based on the linearised BGK model and the flows through ideal labyrinthes. Experiments were carried out by rotors of 120 mm diameter, and pumping speed and ultimate pressure were measured. The comparison between the measurements and the predicted values shows that the present method predicts the performance of the screw pump with a sufficient accuracy for practical applications1. IntroductionIn recent years, screw vacuum pumps have become noticed, since the structure of the pump is simple and liquids or solids are hard to accumulate when sucked with gas or are condensed or solidified in the pump. An analytical model of screw vacuum pump will be useful to design a pump that satisfies specific requirements and to predict pumping characteristics under conditions which differ from the condition designed for. So, we propose an analytical model for the screw vacuum pump.2. Outline of analytical modellingFig. 1shows a pair of meshing rotors of the screw vacuum pump. The volume enclosed by a groove of screw, a crest of thread of another rotor and a stator traps gas andtransfers it from inlet side to outlet side as the rotors rotate. The model is built by the balance among geometrical pumping speed, net throughput and leaks.Display Full Size version of this image (7K)Fig. 1. Configuration of the meshing rotors of screw vacuum pump.2.1. Path of leaks inside the screw vacuum pumpThere are three kinds of clearances inside the screw pump, i.e. the clearance between rotor and stator δO, the radial clearance between rotors δI, and the axial clearance between rotors Δ. In the case of single thread, paths of leaks which come into or out of the third transfer volume (appearing in Fig. 2A) for example, are as follows.Display Full Size version of this image (11K)Fig. 2. Clearance and flowing path of leak: (A) development of a pair of screws;(B) axial clearances.The leak through the axial clearance is considered as a superposition of the major component (represented by the straight arrow in Fig. 2B) and minor component (curved arrow).2.2. Evaluating method of the leaksWe evaluate the leaks by a compound method. The method is compounded of the flow rate derived from BGK equations and diffuse reflections, and the flow rate of ideal labyrinthes.The leak through the clearance between rotor and stator, and the leak through the axial clearance between rotors are both given in the following form(1) and the leak through the radial clearance between rotors is given by(2) where M P is the mass flow rate of Poiseuille flow between parallel plates, M S mass flow rate through a slit, M C mass flow rate of Couette flow between parallel plates, M Rp mass flow rate through a gap between two cylinders induced by pressure difference, M Rr mass flow rate through a gap between two cylinders induced by rotation of the cylinders and M L mass flow rate of ideal labyrinthes.We obtained precise information on M P and M S from the studies of Hasegawa and Sone, and Sone and Itakura [3and 6], respectively. M C is determined by the fact that the dimensionless flow rate equals a half, because of the anti-symmetry of the velocity. M Rp and M Rr are determined by(3)M Rr=ρδI UQ Rr(4) where R C=1/(1/R1+1/R2), U=(R1+R2)ω. Q Rp and Q Rr are obtained by solving the MGL equation [1] by parabolic film approximation [2]. The results are shown in Fig. 3.Display Full Size version of this image (4K)Fig. 3. Nondimensional mass flux through the gap between a cylinder and a plane.The axial clearance has a non-uniform gap, as shown in Fig. 4, then we define mean or representative quantity to apply the above evaluation methods. For example,(5)defines mean square clearance, where S M is the area of lens-like domain appearing in Fig. 4. The other definitions appear in the studies of Ohbayashi et al. [4 and 5].Display Full Size version of this image (3K)Fig. 4. Contour of the gap width of the axial clearance.2.3. Pumping characteristicsAssuming that the pressure changes isothermally, the pressure p i in the i th transfer volume is represented by(6) where V is the volume of one transfer volume, M O leak rate through δO, M I leak rate throu gh δI, M Mb major component of the leak through Δ, M Mc minor component of the leak through Δ via the minimum gap and i O,i I,i Mb,i Mc represent the differences betweenthe number of upstream and downstream transfer volumes corresponding to the clearance or path indexed by subscript.Balance among geometrical pumping speed, net throughput and leaks leads to(7) where T c represents one cycle of periodic pressure change. In the case that the screws are single threaded, T c equals π/ω, because the i th transfer volume comes to the (i+1)th position after half rotation.(6) and (7) can be solved under the following periodic condition:p i t=0=p i−1t=T c, (i=1,2,3,…,n m) (8) 3. ExperimentExperiments are carried out with a screw vacuum pump whose dimensions are shown in Table 1. Fig. 5a shows the comparison between measurements and analytical predictions relating to pumping speed. Fig. 5b shows the comparison relating to ultimate pressure as a function of rotating speed. The measurements and the predictions for ultimate pressure and pumping speed in the inlet pressure over the 100 Pa range are well agreed.Table 1. Dimensions of an experimental screw vacuum pumpDisplay Full Size version of this image (10K)Fig. 5. Comparison between experimental results and analytical predictions: (a) pumping speed vs. inlet pressure; (b) ultimate pressure vs. rotation speed.4. ConclusionsThe conclusions are summarised as follows:1. the screw vacuum pump was analysed, and the analytical model of its pumping characteristics was proposed;2. the analytical model was verified through the experiments. This model has satisfying accuracy for practical applications.References1. S. Fukui, R. Kaneko, Molecular gas film lubrication, in: Handbook ofMicro/Nanotribology, CRC Press, Florida, 1995, Chapter 13, pp. 559–604.2. W.A. Gross, L.A. Matsch, V. Castelli, Fluid Film Lubrication, Wiley, New York, 1980.3. M. Hasegawa and Y. Sone. Phys. Fluids A3 3 (1991), pp. 466–477. Full Text via CrossRef4. T. Ohbayashi, T. Sawada and M. Hamaguchi. Trans. Jpn. Soc. Mech. Eng. B64 621 (1998), pp. 1419–1425.5. T. Ohbayashi, T. Sawada, H. Miyamura, Study on the screw vacuum pump with two piecewise constant lead angles, Trans. Jpn. Soc. Mech. Eng. B 65 (637) (1999) 3048–3053.6. Y. Sone and E. Itakura. J. Vac. Soc. Jpn.33 3 (1990), pp. 92–94.螺杆真空泵的性能预测的研究摘要调查螺杆真空泵的抽泵特性。
螺杆压缩机机械外文文献翻译、中英文翻译、外文翻译
英文原文Screw CompressorThe Symmetric profile has a huge blow-hole area which excludes it from any compressor applicat -ion where a high or even moderate pressure ratio is involved. However, the symmetric profile per -forms surprisingly well in low pressure compressor applications.More details about the circular p -rofile can be found in Margolis, 1978.2.4.8 SRM “A” ProfileThe SRM “A” profile is shown in Fig. 2.11. It retains all the favourable features of the symmetric profile like its simplicity while avoiding its main disadvantage,namely, the large blow-hole area. The main goal of reducing the blow hole area was achieved by allowing the tip points of the main and gate rotors to generate their counterparts, trochoids on the gate and main rotor respectively. T -he “A” profile consists mainly of circles on the gate rotor and one line which passes through the gate rotor axis.The set of primary curves consists of: D2C2, which is a circle on the gate rotor with the centre on the gate pitch circle, and C2B2, which is a circle on the gate rotor, the centre of whi ch lies outside the pitch circle region.This was a new feature which imposed some problems in the generation of its main rotor counterpart, because the mathematics used for profile generation at tha -t time was insufficient for general gearing. This eccentricity ensured that the pressure angles on th -e rotor pitches differ from zero, resulting in its ease of manufacture. Segment BA is a circle on th -e gate rotor with its centre on the pitch circle. The flat lobe sides on the main and gate rotors weregenerated as epi/hypocycloids by points G on the gate and H on the main rotor respectively. GF2 is a radial line at the gate rotor. This brought the same benefits to manufacturing as the previously mentioned circle eccentricity onFig. 2.11 SRM “A” Profile2.4 Review of Most Popular Rotor Profiles 31 the opposite lobe side. F2E2 is a circle with the cent -re on the gate pitch and finally, E2D2 is a circle with the centre on the gate axis.More details on t -he “A” profile are published by Amosov et al., 1977 and by Rinder, 1979.The “A” profile is a go od example of how a good and simple idea evolved into a complicated result. Thus the “A” pro file was continuously subjected to changes which resulted in the “C” profile. This was mainly gen erated to improve the profile manufacturability. Finally, a completely new profile, the“D” profile was generated to introduce a new development in profile gearing and to increase the gate rotor tor -que.Despite the complexity o f its final form the “A” profile emerged to be the most popular scre -w compressor profile, especially after its patent expired.2.4.9 SRM “D” ProfileThe SRM “D” profile, shown in Fig. 2.12, is generated exclusively by circles with the centres off the rotor pitch circles.Similar to the Demonstrator, C2D2 is an eccentric circle of radius r3 onthe gate rotor. B1C1 is an eccentric circle of radius r1, which, together withthe small circular arc A1J1 of radius r2, is positioned on the main rotor. G2H2is a small circular arc on the gate rotor and E2F2 is a circular arc on the gaterotor. F2G2 is a relatively large circular arc on the gate rotor which produces a corresponding curve of the smallest possible curvature on the main rotor.Both circular arc, B2C2 and F2G2 ensure a large radius of curvature in the pitch circle area. This avoids high stresses in the rotor contact region.Fig. 2.12 SRM “D” ProfileThe “G” profile was introduced by SRM in the late nineteen nineties as a replacement for the “D” rotor and is shown in Fig. 2.13. Compared with the“D” rotor, the “G” rotor has the unique feature of two additional circles in the addendum area on both lobes of the main rotor, close to the pitch circle.This feature improves the rotor contact and, additionally, generates shorter sealing lines. This can be seen in Fig. 2.13, where a rotor featuring “G” profile characteristics only on its flat side through segment H1I1 is presented.Fig. 2.13 SRM “G” Profile2.4.11 City “N” Rack Generated Rotor Profile“N” rotors are calculated by a rack generation procedure. This distinguishes them from any others. In this case, the large blow-hole area, which is a characteristic of rack generated rotors, is overcome by generating the high pressure side of the rack by means of a rotor conjugate procedure. This undercuts the single appropriate curve on the rack. Such a rack is then used for profiling both the main and the gate rotors. The method and its extensions were used by the authors to create a number of different rotor profiles, some of them used by Stosic et al., 1986, and Hanjalic and Stosic, 1994. One of the applications of the rack generation procedure is described in Stosic, 1996.The following is a brief description of a rack generated “N” rotor profile,typical of a family of rotor profiles designed for the efficient compression of air,common refrigerants and a number of process gases. The rotors are generated by the combined rack-rotor generation procedure whose features are such that it may be readily modified further to optimize performance for any specific application.2.4 Review of Most Popular Rotor Profiles 33The coordinates of all primary arcs on the rack are summarized here relative to the rack coordinate system. The lobe of the rack is divided into several arcs. The divisions between the profile arcs are denoted by capital letters and each arc is defined separately, as shown in the Figs.2.14 and 2.15 where the rack and the rotors are shown.Fig. 2.14 Rack generated “N” ProfileFig. 2.15 “N” rotor primary curves g iven on rack34 2 Screw Compressor GeometryAll curves are given as a “general arc” expressed as: axp + byq = 1. Thus straight lines, circles, parabolae, ellipses and hyperbolae are all easily described by selecting appropriate values for parameters a, b, p and q.Segment DE is a straight line on the rack, EF is a circular arc of radius r4,segment FG is a straight line for the upper involute, p = q = 1, while segment GH on the rack is a meshing curve generated by the circular arc G2H2 on the gate rotor. Segment HJ on the rack is a meshing curve generated by the circular arc H1J1 of radius r2 on the main rotor. Segment JA is a circular arc of radius r on the rack, AB is an arc which can be either a circle or a parabola, a hyperbola or an ellipse, segment BC is a straight line on the rack matching the involute on the rotor round lobe and CD is a circular arc on the rack, radius r3.More details of the “N” profile can be found in Stosic, 1994.2.4.12 Characteristics of “N” ProfileSample illustrations of the “N” profile in 2-3, 3-5, 4-5, 4-6, 5-6, 5-7 and 6-7 configurations are given in Figs. 2.16 to Fig. 2.23. It should be noted that all rotors considered were obtained automatically from a computer code by simply specifying the number of lobes in the main and gate rotors, and the lobe curves in the general form.A variety of modified profiles is possible. The “N” profile design is a compromise between full tightness, small blow-hole area, large displacement.Fig. 2.16 “N” Rotors in 2-3 configurationFig. 2.17 “N” Rotors in 3-5 configurationFig. 2.18 “N” Rotors in 4-5 configurationFig. 2.19 “N” Rotors in 4-6 configurationFig. 2.20 “N” Rotors compared with “Sigma”, SRM “D” and “Cyclon” rotorsFig. 2.21 “N” Rotors in 5-6 configurationFig. 2.22 “N” Rotors in 5-7 configurationFig. 2.23 “N” rotors in 6/7 configurationsealing lines, small confined volumes, involute rotor contact and proper gate rotor torque distribution together with high rotor mechanical rigidity.The number of lobes required varies according to the designated compressor duty. The 3/5 arrangement is most suited for dry air compression, the 4/5 and 5/6 for oil flooded compressors with a moderate pressure difference and the 6/7 for high pressure and large built-in volume ratio refrigeration applications.Although the full evaluation of a rotor profile requires more than just a geometric assessment, some of the key features of the “N” profile may be readily appreciated by comparing it with three of the most popular screw rotor profiles already described here, (a) The “Sigma” profile by Bammert,1979, (b) the SRM “D” profile by Astberg 1982, and (c) the “Cyclon” profile by Hough and Morris, 1984. All these rotors are shown in Fig. 2.20 where it can be seen that the “N” profiles have a grea ter throughput and a stiffer gate rotor for all cases when other characteristics such as the blow-hole area, confined volume and high pressure sealing line lengths are identical.Also, the low pressure sealing lines are shorter, but this is less important because the corresponding clearance can be kept small.The blow-hole area may be controlled by adjustment of the tip radii on both the main and gate rotors and also by making the gate outer diameter equal to or less than the pitch diameter. Also the sealing lines can be kept very short by constructing most of the rotor profile from circles whose centres are close to the pitch circle. But, any decrease in the blow-hole area will increasethe length of the sealing line on the flat rotor side. A compromise betweenthese trends is therefore required to obtain the best result.2.4 Review of Most Popular Rotor Profiles 39Rotor instability is often caused by the torque distribution in the gate rotor changing direction during a complete cycle. The profile generation procedure described in this paper makes itpossible to control the torque on the gate rotor and thus avoid such effects. Furthermore, full involute contact between the “N” rotors enables any additional contact load to be absorbed more easily than with any other type of rotor. Two rotor pairs are shown in Fig. 2.24 the first exhibits what is described as “negative” gate rotor torque while the second shows the more usual “positive” torque.Fig. 2.24 “N” with negative torque, left and positive torque, right2.4.13 Blower Rotor ProfileThe blower profile, shown in Fig. 2.25 is symmetrical. Therefore only one quarter of it needs to be specified in order to define the whole rotor. It consists of two segments, a very small circle on the rotor lobe tip and a straight line. The circle slides and generates cycloids, while the straight line generates involutes.Fig. 2.25 Blower profile中文译文螺杆压缩机螺杆压缩机的几何形状对称分布有一个巨大的吹孔面积不包括它任何压缩机应用在高或中等压力比参与。
压缩机专业词汇中英文对照大全，你不收藏算我输！
压缩机专业词汇中英文对照大全，你不收藏算我输！中英对照压缩机分类及配件词汇容积式压缩机 positive displacement compressor往复式压缩机（活塞式压缩机） reciprocating compressor 回转式压缩机 rotary compressor滑片式压缩机 sliding vane compressor单滑片回转式压缩机 single vane rotary compressor滚动转子式压缩机 rolling rotor compressor三角转子式压缩机 triangle rotor compressor多滑片回转式压缩机 multi-vane rotary compressor滑片 blade旋转活塞式压缩机 rolling piston compressor涡旋式压缩机 scroll compressor涡旋盘 scroll固定涡旋盘 stationary scroll, fixed scroll驱动涡旋盘 driven scroll, orbiting scroll斜盘式压缩机（摇盘式压缩机） swash plate compressor 斜盘 swash plate摇盘 wobble plate螺杆式压缩机 screw compressor单螺杆压缩机 single screw compressor阴转子 female rotor阳转子 male rotor主转子 main rotor闸转子 gate rotor无油压缩机 oil free compressor膜式压缩机 diaphragm compressor活塞式压缩机 reciprocating compressor单作用压缩机 single acting compressor双作用压缩机 double acting compressor双效压缩机 dual effect compressor双缸压缩机 twin cylinder compressor闭式曲轴箱压缩机 closed crankcase compressor开式曲轴箱压缩机 open crankcase compressor顺流式压缩机 uniflow compressor逆流式压缩机 return flow compressor干活塞式压缩机 dry piston compressor双级压缩机 compound compressor多级压缩机 multistage compressor差动活塞式压缩机stepped piston compound compressor, differential piston compressor串轴式压缩机 tandem compressor, dual compressor截止阀 line valve, stop valve排气截止阀 discharge line valve吸气截止阀 suction line valve部分负荷旁通口 partial duty port能量调节器 energy regulator容量控制滑阀 capacity control slide valve容量控制器 capacity control消声器 muffler联轴节 coupling曲轴箱 crankcase曲轴箱加热器 crankcase heater轴封 crankcase seal, shaft seal填料盒 stuffing box轴封填料 shaft packing机械密封 mechanical seal波纹管密封 bellows seal转动密封 rotary seal迷宫密封 labyrinth seal轴承 bearing滑动轴承 sleeve bearing偏心环 eccentric strap滚珠轴承 ball bearing滚柱轴承 roller bearing滚针轴承 needle bearing止推轴承 thrust bearing外轴承 pedestal bearing臼形轴承 footstep bearing轴承箱 bearing housing止推盘 thrust collar偏心销 eccentric pin曲轴平衡块crankshaft counterweight, crankshaft balance weight曲柄轴 crankshaft偏心轴 eccentric type crankshaft曲拐轴 crank throw type crankshaft连杆 connecting rod连杆大头 crank pin end连杆小头 piston pin end曲轴 crankshaft主轴颈 main journal曲柄 crank arm, crank shaft曲柄销 crank pin曲拐 crank throw曲拐机构 crank-toggle阀盘 valve disc阀杆 valve stem阀座 valve seat阀板 valve plate阀盖 valve cage阀罩 valve cover阀升程限制器valve lift guard阀升程 valve lift阀孔 valve port吸气口 suction inlet压缩机气阀 compressor valve吸气阀 suction valve排气阀 delivery valve圆盘阀 disc valve环片阀 ring plate valve簧片阀 reed valve舌状阀 cantilever valve条状阀 beam valve提升阀 poppet valve菌状阀 mushroom valve杯状阀 tulip valve缸径 cylinder bore余隙容积 clearance volume附加余隙（补充余隙） clearance pocket活塞排量 swept volume, piston displacement理论排量 theoretical displacement实际排量 actual displacement实际输气量 actual displacement, actual output of gas气缸工作容积 working volume of the cylinder活塞行程容积 piston displacement气缸 cylinder气缸体 cylinder block气缸壁 cylinder wall水冷套 water cooled jacket气缸盖（气缸头） cylinder head安全盖（假盖） safety head假盖 false head活塞环 piston ring气环 sealing ring刮油环 scraper ring油环 scrape ring活塞销 piston pin活塞 piston活塞行程 piston stroke吸气行程 suction stroke膨胀行程 expansion stroke压缩行程 compression stroke排气行程 discharge stroke升压压缩机 booster compressor立式压缩机 vertical compressor卧式压缩机 horizontal compressor角度式压缩机 angular type compressor对称平衡型压缩机 symmetrically balanced type compressor 压缩机参数词汇1. performance parameter 性能参数——表征压缩机主要性能的诸参数，如：气量、压力、温度、功率及噪声、振动等2. constructional parameter 结构参数——表征压缩机结构特点的诸参数，如：活塞力、行程、转速、列数、各级缸径、外形尺寸等3. inlet pressure/suction pressure 吸气压力（吸入压力）——在标准吸气位置气体的平均绝对全压力。
中英文文献翻译-螺杆式压缩机
英文原文Screw CompressorsThe direction normal to the helicoids, can be used to calculate the coordinates of the rotorhelicoids n x and n y from x and y to which the clearance is added as:dt dyD p x x n δ+=， dt dxD p y y n δ-=， ⎪⎭⎫ ⎝⎛+=dt dy y dt dx x D z n δ （2.19） where the denominator D is given as ：22222⎪⎭⎫ ⎝⎛++⎪⎭⎫ ⎝⎛+⎪⎭⎫ ⎝⎛=dt dy y dt dx x dt dy p dt dx p x D （2.20） n x and n y serve to calculate new rotor end plane coordinates, x 0n and y 0n ,with the clearances obtained for angles θ = n z /p and τ respectively. These on x and on y now serve to calculate the transverse clearance δ0 as the difference between them, as well as the original rotor coordinates o x and o y .If by any means, the rotors change their relative position, the clearance distribution at one end of the rotors may be reduced to zero on the flat side of the rotor lobes. In such a case, rotor contact will be prohibitively long on the flat side of the profile, where the dominant relative rotor motion is sliding, as shown in Fig. 2.29. This indicates that rotor seizure will almost certainly occur in that region if the rotors come into contact with each other.Fig. 2.29. Clearance distribution between the rotors: at suction, mid rotors, and discharge withpossible rotor contact at the dischargeFig. 2.30. Variable clearance distribution applied to the rotors It follows that the clearance distribution should be non-uniform to avoid hard rotor contact in rotor areas where sliding motion between the rotors is dominant.In Fig. 2.30, a reduced clearance of 65 μm is presented, which is now applied in rotor regions close to the rotor pitch circles, while in other regions it is kept at 85 μm, as was done by Edstroem, 1992. As can be seen in Fig. 2.31, the situation regarding rotor contact is now quite different. This is maintained along the rotor contact belt close to the rotor pitch circles and fully avoided at other locations. It follows that if contact occurred, it would be of a rolling character rather than a combination of rolling and sliding or even pure sliding. Such contact will not generate excessive heat and could therefore be maintained for a longer period without damaging the rotors until contact ceases or the compressor is stopped.2.6 Tools for Rotor ManufactureThis section describes the generation of formed tools for screw compressor hobbing, milling and grinding based on the envelope gearing procedure.2.6.1 Hobbing ToolsA screw compressor rotor and its formed hobbing tool are equivalent to a pair of meshing crossed helical gears with nonparallel and nonintersecting axes. Their general meshing condition is given in Appendix A. Apart from the gashes forming the cutter faces, the hob is simply a helical gear in which.Fig. 2.31. Clearance distribution between the rotors: at suction, mid of rotor and discharge with apossible rotor contact at the dischargeEach referred to as a thread, Colburne, 1987. Owing to their axes not being parallel, there is only point contact between them whereas there is line contact between the screw machine rotors. The need to satisfy the meshing equation given in Appendix A, leads to the rotor – hob meshing requirement for the given rotor transverse coordinate points 1o x and 1o y and their first derivative 0101dx dy .The hob transverse coordinate points 2o x and 2o y can then be calculated. These are sufficient to obtain the coordinate 2012012y x R +=The axial coordinate 2z , calculated directly, and 2R are hob axial plane coordinates which define the hob geometry.The transverse coordinates of the screw machine rotors, described in the previous section, are used as an example here to produce hob coordinates. he rotor unit leads 1P are 48.754mm for the main and −58.504mm for the ate rotor. Single lobe hobs are generated for unit leads 2P :6.291mm for the m ain rotor and −6.291mm for the gate rotor. The corresponding hob helix a ngles ψ are 85◦ and 95◦. The same rotor-to-hob centre distance C = 110mm a nd the shaft angle Σ = 50◦ are given for both rotors. Figure 2.32 contains a view to the hob.Reverse calculation of the hob – screw rotor transformation, also given in Appendix Apermits the determination of the transverse rotor profile coordinates which will be obtained as a result of the manufacturing process. These ay be compared with those originally specified to determine the effect ofFig. 2.32. Rotor manufacturing: hobbing tool left , right milling toolmanufacturing errors such as imperfect tool setting or tool and rotor deformation upon the final rotor profile.For the purpose of reverse transformation, the hob longitudinal plane coordinates 2R and 2z and 22dz dR should be given. The axial coordinate 2z is used to calculate 22P Z T =, which is then used to calculate the hob transverse coordinates:τcos 202R x =， τs i n 202R y = （2.21）These are then used as the given coordinates to produce a meshing criterionand the transverse plane coordinates of the “manufactured” rotors.A comparison between the original rotors and the manufactured rotors is given in Fig. 2.33 with the difference between them scaled 100 times. Two types of error are considered. The left gate rotor, is produced with 30um offset in the centre distance between the rotor and the tool, and the main rotor withFig. 2.33. Manufacturing imperfections0.2◦ of fset in the tool shaft angle Σ. Details of this particular meshing method are given by Stosic 1998.2.6.2 Milling and Grinding ToolsFormed milling and grinding tools may also be generated by placing 02=P in the general meshing equation, given in Appendix A, and then following the procedure of this section. The resulting meshing condition now reads as:[]0cot cot 1111111111=⎥⎦⎤⎢⎣⎡∂∂-∂∂+⎪⎭⎫ ⎝⎛∂∂+∂∂∑+-∑t x C t y p p t y y t x x p x C θ (2.22) However in this case, when one expects to obtain screw rotor coordinates from the tool coordinates, the singularity imposed does not permit the calculation of the tool transverse plane coordinates. The main meshing condition cannot therefore be applied. For this purpose another condition is derived for the reverse milling tool to rotor transformation from which the meshing angle τ is calculated:()0cot sin cot cos 12212222=-∑+∑++⎪⎪⎭⎫ ⎝⎛+C p dR dz C p dR dz z R ττ (2.23) Once obtained, τ will serve to calculate the rotor coordinates after the “manufacturing” process. The obtained rotor coordinates will contain all manufacturing imperfections, like mismatch of the rotor – tool centre distance, error in the rotor – tool shaft angle, axial shift of the tool or tool deformation during the process as they are input to the calculation process. A full account of this useful procedure is given by Stosic 1998.2.6.3 Quantification of Manufacturing ImperfectionsThe rotor – tool transformation is used here for milling tool profile generation. The reverse procedure is used to calculate the “manufactured” rotors. The rack generated 5-6 128mm rotors described by Stosic, 1997a are used as given profiles: x (t ) and y (t ). Then a tool – rotor transformation is used to quantify the influence of manufacturing imperfections upon the qualityof the produced rotor profile. Both, linear and angular offset were considered.Figure 2.33 presents the rotors, the main manufactured with the shaft angle offset 0.5◦and the gate with the centre distance offset 40 μm from that of the original rotors given by the dashed line on the left. On the right, the rotors are manufactured with imperfections, the main with a tool axial offset of 40 μm and the gate with a certain tool body deformation which resulted in 0.5◦offset of the relative motion angle θ. The original rotors are given by the dashed line.3Calculation of Screw Compressor Performance Screw compressor performance is governed by the interactive effects of thermodynamic and fluid flow processes and the machine geometry and thus can be calculated reliably only by their simultaneous consideration. This may be chieved by mathematical modelling in one or more dimensions. For most applications, a one dimensional model is sufficient and this is described in full. 3-D modelling is more complex and is presented here only in outline. A more detailed presentation of this will be made in a separate publication.3.1 One Dimensional Mathematical ModelThe algorithm used to describe the thermodynamic and fluid flow processes in a screw compressor is based on a mathematical model. This defines the instantaneous volume of the working chamber and its change with rotational angle or time, to which the conservation equations of energy and mass continuity are applied, together with a set of algebraic relationships used to define various phenomena related to the suction, compression and discharge of the working fluid. These form a set of simultaneous non-linear differential equations which cannot be solved in closed form.The solution of the equation set is performed numerically by means of the Runge-Kutta 4th order method, with appropriate initial and boundary conditions.The model accounts for a number of “real-life” effects, which may significantly influence the performance of a real compressor. These make it suitable for a wide range of applications and include the following:– The working fluid compressed can be any gas or liquid-gas mixture for which an equation of state and internal energy-enthalpy relation is known, i.e. any ideal or real gas or liquid-gas mixture of known properties.–The model accounts for heat transfer between the gas and the compressor rotors or its casing in a form, which though approximate, reproduces the overall effect to a good first order level of accuracy.– The model accounts for leakage of the working medium through the clearances between the two rotors and between the rotors and the stationary parts of the compressor.– The process equations and the subroutines for their solution are independent of those which define the compressor geometry. Hence, the model can be readily adapted to estimate the performance of any geometry or type of positive displacement machine.– The effects of liquid injection, including that of oil, water, or refrigerant can be accounted for during the suction, compression and discharge stages.– A set of subroutines to estimate the thermodynamic properties and changes of state of the working fluid during the entire compressor cycle of operations completes the equation set and thereby enables it to be solved.Certain assumptions had to be introduced to ensure efficient computation.These do notimpose any limitations on the model nor cause significant departures from the real processes and are as follows:– The fluid flow in the model is assumed to be quasi one-dimensional.–Kinetic energy changes of the working fluid within the working chamber are negligible compared to internal energy changes.–Gas or gas-liquid inflow to and outflow from the compressor ports is assumed to be isentropic.– Leakage flow of the fluid through the clearances is assumed to be adiabatic.3.1.1 Conservation EquationsFor Control Volume and Auxiliary RelationshipsThe working chamber of a screw machine is the space within it that contains the working fluid. This is a typical example of an open thermodynamic system in which the mass flow varies with time. This, as well as the suction and discharge plenums, can be defined by a control volume for which the differential equations of the conservation laws for energy and mass are written. These are derived in Appendix B, using Reynolds Transport Theorem.A feature of the model is the use of the non-steady flow energy equation to compute the thermodynamic and flow processes in a screw machine in terms of rotational angle or time and how these are affected by rotor profile modifications. Internal energy, rather than enthalpy, is then the derived variable. This is computationally more convenient than using enthalpy as the derived Variable since, even in the case of real fluids, it may be derived, without reference to pressure. Computation is then carried out through a series of iterative cycles until the solution converges. Pressure, which is the desired output variable, can then be derived directly from it, together with the remaining required thermodynamic properties.The following forms of the conservation equations have been employed in the model:中文翻译螺杆式压缩机几何的法线方向的螺旋，可以用来计算的坐标转子螺旋n x 和n y 的从x 和y 的间隙加入如：dt dyD p x x n δ+=， dt dxD p y y n δ-=， ⎪⎭⎫ ⎝⎛+=dt dy y dt dx x D z n δ (2.19) 其中分母D 被给定为：22222⎪⎭⎫ ⎝⎛++⎪⎭⎫ ⎝⎛+⎪⎭⎫ ⎝⎛=dt dy y dt dx x dt dy p dt dx p x D (2.20) n x ，n y 服务来计算新的转子端的平面的坐标，on x 和on y ，得到的间隙角θ =锌/ p 和τ 。
螺杆式压缩机的设计外文文献翻译、中英文翻译、外文翻译
英文原文1 IntroductionThe screw compressor is one of the most common types of machine used to compress gases. Its construction is simple in that it essentially comprises only a pair of meshing rotors, with helical grooves machined in them, contained in a casing, which fits closely round them. The rotors and casing are separated by very small clearances. The rotors are driven by an external motor and mesh like gears in such a manner that, as they rotate, the space formed between them and the casing is reduced progressively. Thus, any gas trapped in this case is compressed. The geometry of such machines is complex and the flow of the gas being compressed within them occurs in three stages. Firstly, gas enters between the lobes, through an inlet port at one end of the casing during the start of rotation. As rotation continues, the space between the rotors no longer lines up with the inlet port and the gas is trapped and thus compressed. Finally, after further rotation, the opposite ends of the rotors pass a second port at the other end of the casing, through which the gas is discharged. The whole process is repeated between successive pairs of lobes to create a continuous but pulsating flow of gas from low to high pressure.These machines are mainly used for the supply of compressed air in the building industry, the food, process and pharmaceutical industries and, where required, in the metallurgical industry and for pneumatic transport.They are also used extensively for compression of refrigerants in refrigeration and air conditioning systems and of hydrocarbon gases in the chemical industry. Their relatively rapid acceptance over the past thirty years is due to their relatively high rotational speeds compared to other types of positive displacement machine, which makes them compact, their ability to maintain high efficiencies over a wide range of operating pressures and flow rates and their long service life and high reliability. Consequently, they constitute a substantial percentage of all positive displacement compressors now sold and currently in operation.The main reasons for this success are the development of novel rotor profiles, which have drastically reduced internal leakage, and advanced machine tools, which can manufacture the most complex shapes to tolerances of the order of 3 micrometers at an acceptable cost. Rotor profile enhancement is still the most promising means of further improving screw compressors and rational procedures are now being developed both to replace earlier empirically derived shapes and also to vary the proportions of the selected profile to obtain the best result for the application for which the compressor is required. Despite their wide usage, due to the complexity of their internal geometry and the non-steady nature of the processes within them, up till recently, only approximate analytical methods have been available to predict their performance. Thus, although it is known that their elements are distorted both by the heavy loads imposed by pressure induced forces and through temperature changes within them, no methods were available to predict the magnitude of these distortions accurately, nor how they affect the overall performance of the machine. In addition, improved modelling of flow patterns within the machine can lead to better porting design. Also, more accurate determination of bearing loads and how they fluctuate enable better choices of bearings to be made. Finally, if rotor and casing distortion, as a result of temperature and pressure changes within the compressor, can be estimated reliably, machining procedures can be devised to minimise their adverse effects.Screw machines operate on a variety of working fluids, which may be gases, dry vapour or multi-phase mixtures with phase changes taking place within the machine. They may involve oil flooding, or other fluids injected during the compression or expansion process, or be without any form of internal lubrication. Their geometry may vary depending on the number of lobes in each rotor, the basic rotor profile and the relative proportions of each rotor lobe segment. It follows that there is no universal configuration which would be the best for all applications. Hence, detailed thermodynamic analysis of the compression process and evaluation of the influence of the various design parameters on performance is more important to obtain the best results from these machines than from other types which could be used for the same application. A set of well defined criteria governed by an optimisation procedure is therefore a prerequisite for achieving the best design for each application. Such guidelines are also essential for the further improvement of existing screw machine designs and broadening their range of uses. Fleming et al., 1998 gives a good contemporary review of screw compressor modelling, design and application.A mathematical model of the thermodynamic and fluid flow processes within positive displacement machines, which is valid for both the screw compressor and expander modes of operation, is presented in this Monograph. It includes the use of the equations of conservation of mass, momentum and energy applied to an instantaneous control volume of trapped fluid within the machine with allowance for fluid leakage, oil or other fluid injection, heat transfer and the assumption of real fluid properties. By simultaneous solution of these equations, pressure-volume diagrams may be derived of the entire admission, discharge and compression or expansion process within the machine. A screw machine is defined by the rotor profile which is here generated by use of a general gearing algorithm and the port shape and size. This algorithm demonstrates the meshing condition which, when solved explicitly,enables a variety of rotor primary arcs to be defined either analytically or by discrete point curves. Its use greatly simplifies the design since only primary arcs need to be specified and these can be located on either the main or gate rotor or even on any other rotor including a rack, which is a rotor of infinite radius. The most efficient profiles have been obtained from a combined rotor-rack generation procedure.The rotor profile generation processor, thermofluid solver and optimizer,together with pre-processing facilities for the input data and graphical post processing and CAD interface, have been incorporated into a design tool in the form of a general computer code which provides a suitable tool for analysis and optimization of the lobe profiles and other geometrical and physical parameters. The Monograph outlines the adopted rationale and method of modelling, compares the shapes of the new and conventional profiles and illustrates potential improvements achieved with the new design when applied to dry and oil-flooded air compressors as well as to refrigeration screw compressors.The first part of the Monograph gives a review of recent developments in screw compressors.The second part presents the method of mathematical definition of the general case of screw machine rotors and describes the details of lobe shape specification. It focuses on a new lobe profile of a slender shape with thinner lobes in the main rotor, which yields a larger cross-sectional area and shorter sealing lines resulting in higher delivery rates for the same tip speed.The third part describes a model of the thermodynamics of the compression-expansion processes, discusses some modelling issues and compares the shapes of new and conventional profiles. It illustrates the potentialimprovements achievable with the new design applied to dry and oil-flooded air compressors as well as to refrigeration screw compressors. The selection of the best gate rotor tip radius is given as an example of how mathematical modelling may be used to optimise the design and the machine’s operating conditions.The fourth part describes the design of a high efficiency screw compressor with new rotor profiles. A well proven mathematical model of the compression process within positive displacement machines was used to determine the optimum rotor size and speed, the volume ratio and the oil injection position and jet diameter. In addition, modern design concepts such as an open suction port and early exposure of the discharge port were included, together with improved bearing and seal specification, to maximise the compressor efficiency. The prototypes were tested and compared with the best compressors currently on the market. The measured specific power input appeared to be lower than any published values for other equivalent compressors currently manufactured. Both the predicted advantages of the new rotor profile and the superiority of the design procedure were thereby confirmed.1.1 Basic ConceptsThermodynamic machines for the compression and expansion of gases and vapours are the key components of the vast majority of power generation and refrigeration systems and essential for the production of compressed air and gases needed by industry. Such machines can be broadly classified by their mode of operation as either turbomachines or those of the positive displacement type.Turbomachines effect pressure changes mainly by dynamic effects, related to the change of momentum imparted to the fluids passing through them. These are associated with the steady flow of fluids at high velocities and hence these machines are compact and best suited for relatively large mass flow rates. Thus compressors and turbines of this type are mainly used in the power generation industry, where, as a result of huge investment in research and development programmes, they are designed and built to attain thermodynamic efficiencies of more than 90% in large scale power production plant. However, the production rate of machines of this type is relatively small and worldwide, is only of the order of some tens of thousands of units per annum.Positive displacement machines effect pressure changes by admitting a fixed mass of fluid into a working chamber where it is confined and then compressed or expanded and, from which it is finally discharged. Such machines must operate more or less intermittently. Such intermittent operation is relatively slow and hence these machines are comparatively large. They are therefore better suited for smaller mass flow rates and power inputs and outputs. A number of types of machine operate on this principle such as reciprocating, vane, scroll and rotary piston machines.In general, positive displacement machines have a wide range of application, particularly in the fields of refrigeration and compressed air production and their total world production rate is in excess of 200 million units per annum. Paradoxically, but possibly because these machines are produced by comparatively small companies with limited resources, relatively little is spent on research and development programmes on them and there are very few academic institutions in the world which are actively promoting their improvement.One of the most successful positive displacement machines currently in use is the screw or twin screw compressor. Its principle of operation, as indicated in Fig. 1.1, is based on volumetric changes in three dimensions rather than two. As shown, it consists, essentially, of a pair of meshing helical lobed rotors, contained in a casing.The spaces formed between the lobes on each rotor form a series of working chambers in which gas or vapour is contained. Beginning at the top and in front of the rotors, shown in the light shaded portion of Fig. 1.1a, there is a starting point for each chamber where the trapped volume is initially zero. As rotation proceeds in the direction of the arrows, the volume of that chamber then increases as the line of contact between the rotor with convex lobes, known as the main rotor, and the adjacent lobe of the gate rotorFig. 1.1. Screw Compressor Rotorsadvances along the axis of the rotors towards the rear. On completion of one revolution i.e. 360◦by the main rotor, the volume of the chamber is then a maximum and extends in helical form along virtually the entire length of the rotor. Further rotation then leads to reengagement of the main lobe with the succeeding gate lobe by a line of contact starting at the bottom and front of the rotors and advancing to the rear, as shown in the dark shaded portions in Fig. 1.1b. Thus, the trapped volume starts to decrease. On completion of a further 360◦of rotation by the main rotor, the trapped volume returns to zero.The dark shaded portions in Fig. 1.1 show the enclosed region where therotors are surrounded by the casing, which fits closely round them, while the light shaded areas show the regions of the rotors, which are exposed to external pressure. Thus the large light shaded area in Fig. 1.1a corresponds to the low pressure port while the small light shaded region between shaft ends B and D in Fig. 1.1b corresponds to the high pressure port.Exposure of the space between the rotor lobes to the suction port, as their front ends pass across it, allows the gas to fill the passages formed between them and the casing until the trapped volume is a maximum. Further rotation then leads to cut off of the chamber from the port and progressive reduction in the trapped volume. This leads to axial and bending forces on the rotors and also to contact forces between the rotor lobes. The compression process continues until the required pressure is reached when the rear ends of the passages are exposed to the discharge port through which the gas flows out at approximately constant pressure. It can be appreciated from examination of Fig. 1.1, is that if the direction of rotation of the rotors is reversed, then gas will flow into the machine through the high pressure port and out through the low pressure port and it will act as an expander. The machine will also work as an expander when rotating in the same direction as a compressor provided that the suction and discharge ports are positioned on the opposite sides of the casing to those shown since this iseffectively the same as reversing the direction of rotation relative to the ports. When operating as a compressor, mechanical power must be supplied to shaft A to rotate the machine. When acting as an expander, it will rotate automatically and power generated within it will be supplied externally through shaft A.The meshing action of the lobes, as they rotate, is the same as that of helical gears but, in addition, their shape must be such that at any contact position, a sealing line is formed between the rotors and between the rotors and the casing in order to prevent internal leakage between successive trapped passages. A further requirement is that the passages between the lobes should be as large as possible, in order to maximise the fluid displacement per revolution. Also, the contact forces between the rotors should be low in order to minimise internal friction losses.A typical screw rotor profile is shown in Fig. 1.2, where a configuration of 5–6 lobes on the main and gate rotors is presented. The meshing rotors are shown with their sealing lines, for the axial plane on the left and for the cross-sectional plane in the centre. Also, the clearance distribution between the two rotor racks in the transverse plane, scaled 50 times (6) is given above.Fig. 1.2. Screw rotor profile: (1) main, (2) gate, (3) rotor external and (4) pitch circles, (5) sealing line, (6) clearance distribution and (7) rotor flow area between the rotors and housingOil injected Oil FreeFig. 1.3. Oil Injected and Oil Free CompressorsScrew machines have a number of advantages over other positive displacement types. Firstly, unlike reciprocating machines, the moving parts all rotate and hence can run at much higher speeds. Secondly, unlike vane machines, the contact forces within them are low, which makes them very reliable. Thirdly, and far less well appreciated, unlike the reciprocating, scroll and vane machines, all the sealing lines of contact which define the boundaries of each cell chamber, decrease in length as the size of the working chamber decreases and the pressure within it rises. This minimises the escape of gas from the chamber due to leakage during the compression or expansion process.1.2 Types of Screw CompressorsScrew compressors may be broadly classified into two types. These are shown in Fig. 1.3 where machines with the same size rotors are compared:1.2.1 The Oil Injected MachineThis relies on relatively large masses of oil injected with the compressed gas in order to lubricate the rotor motion, seal the gaps and reduce the temperature rise during compression. It requires no internal seals, is simple in mechanical design, cheap to manufacture and highly efficient. Consequently it is widely used as a compressor in both the compressed air and refrigeration industries.1.2.2 The Oil Free MachineHere, there is no mixing of the working fluid with oil and contact between the rotors is prevented by timing gears which mesh outside the working chamber and are lubricated externally. In addition, to prevent lubricant entering the working chamber, internal seals are required on each shaft between the working chamber and the bearings. In the case of process gas compressors, double mechanical seals are used. Even with elaborate and costly systems such as these, successful internal sealing is still regarded as a problem by established process gas compressor manufacturers. It follows that such machines are considerably more expensive to manufacture than those that are oil injected.Both types require an external heat exchanger to cool the lubricating oil before it is readmitted to the compressor. The oil free machine requires an oil tank, filters and a pump to return the oil to the bearings and timing gear.The oil injected machine requires a separator to remove the oil from the high pressure discharged gas but relies on the pressure difference between suction and discharge to return the separated oil to the compressor. Theseadditional components increase the total cost of both types of machine but the add on cost is greater for the oil free compressor.1.3 Screw Machine DesignSerious efforts to develop screw machines began in the nineteen thirties, when turbomachines were relatively inefficient. At that time, Alf Lysholm, a talented Swedish engineer, required a high speed compressor, which could be coupled directly to a turbine to form a compact prime mover, in which the motion of all moving parts was purely rotational. The screw compressor appeared to him to be the most promising device for this purpose and all modern developments in these machines stem from his pioneering work. Typical screw compressor designs are presented in Figs. 1.4 and 1.5. From then until the mid nineteen sixties, the main drawback to their widespread use was the inability to manufacture rotors accurately at an acceptable cost. Two developments then accelerated their adoption. The first was the development of milling machines for thread cutting. Their use for rotor manufacture enabled these components to be made far more accurately at an acceptable cost. The second occurred in nineteen seventy three, when SRM, in Sweden, introduced the “A” profile, which reduced the internal leakage path area, known as the blow hole, by 90%. Screw compressors could then be built with efficiencies approximately equal to those of reciprocating machines and, in their oil flooded form, could operate efficiently with stage pressure ratios of up to 8:1. This was unattainable with reciprocating machines. The use of screw compressors, especially of the oil flooded type, then proliferated.Fig. 1.4. Screw compressor mechanical partsFig. 1.5. Cross section of a screw compressor with gear boxTo perform effectively, screw compressor rotors must meet the meshing requirements of gears while maintaining a seal along their length to minimise leakage at any position on the band of rotor contact. It follows that the compressor efficiency depends on both the rotor profile and the clearances between the rotors and between the rotors and the compressor housing.Screw compressor rotors are usually manufactured on pecialized machines by the use of formed milling or grinding tools. Machining accuracy achievable today is high and tolerances in rotor manufacture are of the order of 5 μm around the rotor lobes. Holmes, 1999 reported that even higher accuracy was achieved on the new Holroyd vitrifying thread-grinding machine, thus keeping the manufacturing tolerances within 3 μm even in large batch production. This means that, as far as rotor production alone is concerned, clearances betweenthe rotors canbe as small as 12 μm.中文译文1 引言螺杆式压缩机是一种最常见的用来压缩气体的机器。
螺杆压缩机毕业设计(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。
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。
压缩 机 中英词汇
Foundations 基础
Guide vane 导叶
Impeller 叶轮
Intercooling 中间冷却器
Coupling 联轴器
Diaphragm 隔板
Diffuser 扩压器
Discharge nozzle 排气接管
Discharge volute 排气蜗壳
Elastohydrodynamic 弹性流体动压
Elliptical 椭圆形
Film thickness 膜厚
Flow 流量
Fluid film 流体膜
Gas 气体
Hydrostatic 流体静压
Journal 轴颈
Liner 衬套
Materials 材料
Stress distribution 应力分布
Tie rods 拉杆
Damped systems 阻尼系统
Dewhirl vanes 破涡片
Inlet nozzle 进气（接）管
Inlet volute 进气蜗壳
Multistage 多级
Off design operation 非设计工况操作
Oil system 密封油系统
Absolute pressure 绝对压力
Absolute temperature 绝对温度
Adiabatic compression 绝热压缩过程
Air padding (压缩)空气（填充）输送
Aluminalkyle 烷基铝
Performance 性能
Slope 倾斜
Splitter vane 分流叶片
制冷压缩机中英文对照外文翻译文献
中英文对照外文翻译Small COMPRESSORCompressor refrigeration system is the core and heart of its decision to the refrigeration system capabilities and features. This paper not only energy efficient, noise and vibration and refrigeration agent analyzed small refrigeration compressor technical performance, Analysis also have appeared in recent years, the new, special small compressor main feature for us small refrigeration compressor future development trend of laying a technological foundation.As we all know, the compressor refrigeration system is the core and heart. Compressor and decided that the cooling system capacity and features. In a sense, the cooling system design and matching of the compressor is the ability demonstrated. Therefore, countries in the world are all in the refrigeration industry refrigeration compressor research invested a tremendous amount of energy, new research direction, and research results continue to emerge. Compressor technology and performance level with each passing day.1.A compressor Efficiency StudyCompressor refrigeration system is the core energy components, improve the efficiency of refrigeration systems of the most direct and effective means is to increase the efficiency of the compressor, It will bring the energy consumption decreased significantly. Moreover, can only avoid the system take measures (such as simply increasing heat exchanger area, etc.) caused by the consumption of materials increased. In recent years, as world energy shortage situation worsens day by day, more and more attention to various energy-saving work the energy efficiency ofproducts made by the ever-increasing demands. Due to losses such as friction, leakage, harmful heat, the electrical loss, flow resistance, noise vibration of existence, Compressor work far below the actual efficiency of theoretical efficiency. Therefore, from a theoretical point of view, any reduction in a loss of arbitrary measures to improve the efficiency of the compressor. The objective facts have led to the energy saving compressor scope, direction, width, research topics and results varied.On the current international energy-efficient compressors research concentrated mainly in a few areas : research lubrication properties Compressor parts of the friction bearings to reduce friction characteristics of power, improve the efficiency of the compressor; reduce leakage losses to improve the efficiency of the compressor; using frequency modulation technology or refrigeration system through the effort with the user load to match the best energy saving In this regard the particular frequency technology has been relatively mature well known and not repeat them here. Valve Research is an old topic but it is also an eternal topic, Improvement of the valve designed to improve the efficiency of the compressor also Nagamochi endless harvest. Research in this area many times, from the valve material, sports law, optimizing the structure of the applicable theory, exhaustive testing methods. In short, energy-saving compressors on the research in recent years has become one of the refrigeration industry first hot issues.In recent years, domestic refrigeration compressor industry to studyenergy-saving products are also giving great concern. Progress larger products mainly refrigerator compressor industry. In China efficient refrigerators GEF projects to promote and support, both the enterprises for energy-efficient products is the understanding of the performance of refrigerator compressors have a qualitative leap. At present, domestic enterprises refrigerator compressor products of the highest energy efficiency has reached 1.95%. Refrigerator compressor domestic enterprises to take a lot of technical measures such as high efficiency motors or synchronous motor, concave valves, Plane thrust bearing, low viscosity lubricants, the new Getter muffler,reducing friction losses, and achieved great results. The main problem is the lack of domestic enterprises currently free technology, the technology has to imitate the line mainly, Most of the enterprises to build their own technology infrastructure also unconscious, nor the interest, and this restricts the development of technological capacity.Relative to the refrigerator compressor industry, domestic energy-efficientair-conditioning compressor study it was not perturbed, Over the years the efficiency of the compressor is no substantive change, greater market demand makes most of the air-conditioning compressor enterprises will concentrate on expanding production on. With the nation on the air conditioner energy efficiency standards set for the further improvement of China's air conditioner exports various perils of showing, domestic air-conditioning compressor of this short-sighted enterprises will be unable to adapt to the energy-saving development of the situation. Enterprise also on the follow-up is weak.2. Compressor noise and vibration studyCurrently, the noise has been regarded as one of the serious pollution. Household refrigeration equipment as the source of power and heart, refrigeration compressor noise, to be a measure of its performance as an important indicator. In fact, to a compressor speaking, Most of the noise is due to shell by some noise from the source excitation (such as springs, refrigerant pressure pulsation, exhaust pipe, lubricants etc. excited). But compressor noise sources and pathways complex and diverse, which gives the compressor noise silencer brought great difficulties.On the compressor noise, vibration and foreign scholars have conducted a large number of long-term research. Here in this regard to the main research results are summarized below :The main refrigeration compressor noise Exaggerative inlet, exhaust radiation aerodynamic noise, mechanical moving parts of machinery noise and noise-driven motor three components :2.1 Aerodynamic noiseCompressor inlet airflow noise is due to the intake manifold pressure pulsation in the elections. Inlet-frequency noise and the intake manifold gas Lane same frequency pulsating with the speed of the compressor. Compressor exhaust noise is due to air in the exhaust pipe caused by fluctuating pressures. Exhaust noise than the inlet noise weak, so the compressor aerodynamic noise generally Inlet mainly noise2.2 Mechanical NoiseCompressor mechanical noise, including members of the general impact and friction, the piston vibration, noise impact of the valve, These noise with randomness, was puted.2.3Electromagnetic noiseCompressor electromagnetic noise is generated by the motor. Motor noise and aerodynamic noise and mechanical noise is weaker compared. Noise source compressor inlet, exhaust, aerodynamic noise, the strongest, followed by mechanical noise and electromagnetic noise. Through in-depth studies, we can further that the main compressor noise from the vibration (from the Department of spring, Refrigeration medium pressure pulsation and smoke exhaust pipe and lubricants have incentive) to the ambient medium spread formation noise. On the effort to reduce compressor noise, much of the literature (abbreviated) proposed a series of measures and the Noise and Vibration Reduction program :① increase rigid shell structure to improve the overall resonance frequency reduces vibration amplitude;② curvature of the shell to avoid mutation, the surface, and the natural frequency is inversely proportional to the radius of curvature. shell shape it should be the smallest curvature radius;③ spring bearing flags will be moved to higher rigid position;④ shell should be used as little as possible of the plane; bending stress and the stress coupling membrane (just on the surface) will shell itself is fairly rigid. Therefore compressor shell to be used as little as possible planar structure;⑤ avoid the exhaust pipe and condenser incentive, optimizing exhaust flow pulsation, Exhaust pipe used in the introduction of additional volume to the elimination of pressure fluctuation spectrum of high-order harmonics;⑥ non-symmetric shell shape; Symmetrical three-dimensional structure means that the axis, along the main axis biggest stress of least resistance. Therefore it is asymmetrical shell structure means that the compressor can be greatly reduced along the axis direction of a force while the probability;⑦ set inlet, exhaust muffler, the closed Compressor Muffler generally muffler. It uses Cross Section, resonant cavity caused acoustic impedance changes in reflectivity or sound energy consumption. or use acoustic-acoustic send phase difference of 180 degrees to offset the muffler of noise. Shell compressor in the lateral closed Unicom a Helmholtz resonator, namely : Helmholtz resonator from the chamber through the neck hole and shell compressor connected into the internal cavity, to reduce compressor cavity stimulated acoustic modal amplitude. The results showed :resonator resonance frequency modulation of the actual compressor cavity stimulated the greatest vibration modes, will be substantially reduced resonance peak response spectrum and lead to significant change. However, it will affect the appearance and the compressor refrigerator settings, the research results are not yet applied to products.Lubricants and residual volume-coil motor windings will lead to the same types of bulk compressor levels between different (from levels average). By changing the shell external support to increase torsional stiffness and reduce vibration surface; Noise study the complex requirements of researchers has strong theory, the enterprise has good skills base and the need for greater investment and a longer timeframe. This is domestic enterprises compressor one of the weak links, which is now basically in the qualitative phase of experimental research, Along with a great chance and randomness.3. new refrigerants ApplicationBased on the new environmental requirements of refrigerant compressor refrigeration industry is a hot issue. As for the refrigerator product R22 refrigerant substitutes the end of the work, new refrigerant compressor in the past few years mainly concentrated in the air conditioning industry. Apart from the now relatively mature R410A, R407C the study, The largest is the hot issue of CO2 compressor. This is the only issue for a briefing.CO2 currently on the research and application of concentrated mainly in three aspects : one is the most urgent need of alternative refrigerants applications, such as automotive air conditioning, as refrigerant emissions, environmental harm, must be adopted as soon as possible without endangering the environment refrigerants; the other is to consider the characteristics of CO2 cycle, the most favorable to the use of this cycle of occasions, If heat pump water heater is to supercritical CO2 in hot conditions decentralization there is a significant temperature slip will help heat Water heated to a higher temperature characteristics of the focus of public attention; anotherone is CO2 into account the nature of heat transfer properties and characteristics of using CO2 as a refrigerant, taking into account CO2 good cold flow properties and heat transfer characteristics, use it as a cascade refrigeration cycle cryogenic stage refrigerants.Compressor transcritical carbon dioxide as an air conditioning system efficiency and reliability of the most affected parts, It should be fully integrated supercritical carbon dioxide cycle specific characteristics of a new design. Like ammonia and CO2, the adiabatic exponent K value higher, up 1.30, it may result in the compressor discharge temperature high, However, as the needs of CO2 compressor pressure ratio small, there is no need for cooling the compressor itself. Adiabatic index is high pressure over the small, I can reduce the gap compressor further expansion of the volume losses to the higher volume efficiency compressors. After experimental and theoretical research, Jurgen Horst SUB and found Kruse, reciprocating compressor is a good film sliding seal as the preferred CO2 system. 8:3 its carbon dioxide compressor exhaust valve for improved Exhaust improved compressor efficiency of carbon dioxide increased by 7%.As the carbon dioxide pressure is far greater than the traditional critical circulatory pressure, compressor shaft seal design requirements than the original compressor is much higher, compressor shaft seal leakage over a period of time is still hampered Actually, the main reason.Danfoss, Denso, ZEXEL such as carbon dioxide compressor has entered the stage of small batch production.The IEA in March 1999, the Joint Japan, Norway, Sweden, Britain and the United States to activate the "Selected Issue on CO2 as working fluid Compression Systems in the "three-year project.Beginning in 1994, BMW, DAIMLERBENZ VOL O, Germany's Volkswagen and Danfoss. Péchiney and other European companies launched the famous "RACE" to the joint project, the Joint European well-known universities, automotive air conditioning manufacturers and other developed CO2 automotive air-conditioning system. Subregion Motor Company has production equipment CO2 carair-conditioning systems of cars, Germany KONVECTA production to the quality of CO2 in the air-conditioned Buses run from 1996 to date. DANFOSS, the Obrist Austria, the United Kingdom have developed a carbon dioxide compressor motor. Japan DENSO, ZEXEL CO2 compressor has entered the stage of mass production.With major manufacturers inputs, the type of CO2 compressor with ordinary motor compressor trend line major swing to determine the displacement swashplate, scroll and the main variable displacement.4. New principle of refrigeration compressorsIn recent years, the new structure and working principle of refrigeration compressor made a more progress, mainly linear compressor, Elliptic compressors, compressor rotor Swing, spiral vane compressor, in the past, the author has been described in the article, here will not repeat it.Linear compressor which is the domestic refrigerator compressor industry the focus of attention. In 2004 the International Compressor Engineering Conference has five linear compressor on the article. LG and researchers still Sunpower two main companies. The past two years, several domestic enterprises in the refrigerator compressor to the development of the linear compressor, However, enterprises have the technical foundation for the domestic financial strength and the limitations of scientific research institutions, believe in a short period of time can not enter the stage of industrialization.5 the classification of the refrigeration compressor5.1 reciprocating compressorReciprocating compressor is a kind of traditional refrigeration compressor, the biggest characteristic is to achieve the capacity and pressure than any of the design. Although it is widely applied, but the market share is gradually reduced.So far, the fridge (including small freezing and cold storage device) host compound compressor is ever to give priority to. Through the optimal design of valve structure, friction pair, reciprocating refrigerator compressor refrigeration coefficient of power refrigerating capacity (units) by 1.0 (w/w) of the early ninety s to today's 1.8 or so; In addition to the energy saving technology progress, and environmental protection is closely related to the refrigerant alternative technology has also made gratifying progress, refrigerator system in our country has a large number of using R600 hydrocarbons, such as small refrigeration device is also used the new working substance such as everything. To further improve the efficiency of the reciprocating compressor refrigerator, to reduce the system noise is still the development direction of it.5.2 linear compressorStill make reciprocating linear compressor, due to the linear motion of the motor can be directly drives the piston reciprocating motion, so as to avoid the complexity of the crank connecting rod mechanism and the resulting mechanical power consumption. Linear compressor assembly as the refrigerator system has appeared, the refrigeration coefficient of linear refrigerator compressor has more than 2.0 (w/w), market prospects look good. The main problem is the design of the compressor oil system and the effective control of linear motor displacement limit point and the corresponding anti-collision cylinder technology.5.3 the swash plate compressorSwash plate compressor is also a kind of variant structure of reciprocating compressor, is mainly used in automotive air conditioning system at present. Afterdecades of development, the swash plate compressor has become a very mature model, in possession of more than 70 of the automotive air conditioning compressor market. In spite of this, because it still belongs to the series of reciprocating structure, so in the car air conditioning system can effect comparing (refrigeration coefficient) and only around 1.5, weight and volume is big, big. Because of the mature of swashplate automobile air conditioning compressor technology, combined with technology, further improvement in the foreseeable future, will continue to maintain a certain market share, but in a certain displacement range by substituting is inevitable.5.4 rotor compressorRotor compressor in the 1970 s by the attention of domestic, it represents the structure including the rolling piston type, sliding-vane, etc. On the rolling piston type is widely used in household air conditioner at present, there are also some applications on the refrigerator. This kind of compressor don't need air suction valve, make it suitable for variable speed operation, which can improve system performance by frequency conversion control. In order to ensure high power (3 p) of the motor output power in the performance of the rolling piston compressor, the domestic research and development and the end of last century, double rotor rolling piston compressor, is now on the market. Double rotor on the rolling piston compressor structure has two advantages: (1) force of the rotating system be improved, the machine vibration and noise is reduced; (2) increase the standalone swept volume and improve the output power of the motor.Below 3 p air conditioning unit, temporarily can not replace a good model of the rolling piston compressor. So improve the efficiency of the compression process, reduce noise and motor speed control and the R410A and other related technical issues after new refrigerating agent, etc., is a research direction of the rolling piston compressor.Sliding vane compressor is a kind of rotor compressor, mainly used to provide compressed air, displacement is in commonly 0.3-3 m3 / min, the market share is low.Rotary vane compressor sliding vane compressor is a kind of transition structure, because of its better starting performance, the compression process torque change is not big, at present is mainly used for miniature cars and some smaller displacement plumbing vehicle air conditioning system. The dynamic characteristics under high speed is the main technology of this compressor research direction.5.5 screw compressorScrew compressor with small size, light weight, easy to maintenance etc., is a model of the fast development in refrigeration compressor. On the one hand, the screw type line, structural design has made considerable progress, on the other hand, the introduction of special screw rotor milling especially grinding, improve the machining precision and machining efficiency of key parts, makes the performance of screw compressor has been effectively improved, industrialization production of the necessary hardware also has the safeguard. At present, the screw compressor is given priority to with compressed air, in medium ReBengShi air conditioning has successful application in the system. Due to increasing the reliability of the screw compressor work within the scope of the medium refrigerating capacity has gradually replace of reciprocating compressor and occupied most of the centrifugal compressor market. 5.6 scroll compressorScroll compressor has been rapid development in the past ten years, the structure of the basic theory, research and development to achieve large-scale industrial production, industrial prototype constitutes the compressor technology development new luminescent spot. The development of numerical control processing technology to realize the mass production, the vortex compressor incomparable performance advantage is the precondition of its vast in the market. A few short years, has been in the field of cabinet air conditioning holds an absolute advantage. In cabinet air conditioning system, scroll compressor refrigeration coefficient has amounted to 3.4 (w/w); In the field of automotive air conditioning, the refrigeration coefficient of scroll compressor has amounted to 2.0 (w/w), and shows strong competition potential. The development of the vortex compressor is to enlarge itsrange of refrigerating capacity, further improve the efficiency, using alternative working medium and lower the manufacturing cost, etc.Since there is no valve, compression force and torque and small changes in the structure make it more suitable for the advantages of frequency control of motor speed operation, it also become the main direction of scroll compressor technology development. Development of scroll compressor of variable displacement mechanism is the key point of the development of the technology. At present, the use of axial sealing technology, "flexible" theory can realize cooling/heating capacity of 10% to 100% within the scope of the regulation.Due to the vortex compressor suction exhaust characteristic of almost continuous, low starting torque and liquid impact resistance, created the condition for parallel use of vortex compressor. In parallel with the vortex compressor can greatly increase the cooling capacity of the unit, can increase from the current single 25 horsepower to single unit 100 horsepower (4 sets of single parallel), and makes the cold quantity adjustment is more reasonable, give full play to the single machine with the highest efficiency. But single in parallel, one of the biggest problems is the oil return is not the average of the unit when using single machine burning phenomenon.3.1.5 centrifugal compressorAt present in large quantity of cold (greater than 1500 kw) remain within the scope of advantage, this is mainly benefited from the cold quantity range, it has incomparable system overall efficiency. The movement of the centrifugal compressor parts little and simple, and its manufacturing precision is much lower than the screw compressor, these are the characteristics of the manufacturing cost is relatively low, and reliable. Relatively speaking, the development of centrifugal compressor is slow, due to the challenges of the screw compressor and absorption chiller. Centrifuge market capacity is around 700 ~ between 1200, because under the premise that the current technology, the machine is mainly used for air conditioning of large buildings, demand is limited. In recent years because of the large infrastructure projects are built, the centrifugal refrigeration and air conditioning compressor is becoming a hot spot ofattention again. Solve surge phenomenon, improve the volume adjustment and the adaptability to change with working condition, miniaturization technology is the main development direction of the centrifugal compressor technology.3.1.6 other structure formsSingle tooth of the compressor, some structures, such as cross slider compressor unique positive displacement compressor also has a certain degree of development, but has not been formed in the domestic production capacity.5. Special refrigeration compressorsAlthough domestic enterprises household refrigeration compressors long accustomed to large-scale production mode, we are accustomed to the number of effectiveness. However, the fierce price competition situation, as products become increasingly lower profit margins, When the production of millions of compressors can only make a few million dollars of profit, some on special refrigeration compressors can be regarded as a way out. Special refrigeration compressor exhaustive, it is impossible in this enumeration. But their common feature is their production scale is small, a single high-profit products faster transition, In most cases the need for the user's requirements designed. These products lead to more and more domestic enterprises to the compressor. If the number of domestic enterprises are developing or already have production capacity of the refrigerator compressor truck翻译小型制冷压缩机研究压缩机是制冷系统的核心和心脏，它决定了制冷系统的能力和特征。
单螺纹杆机毕业设计外文翻译--对压缩机单螺杆专用加工机床的介绍
Dedicated to the single screw compressor machine updated the IntroductionAbstract: This paper describes four areas from the existing single-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 actualapplication in the design specifications of the machine into two, a φ95 ~ 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 use of 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 cylindrical spindle at t he 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 of Tool 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 the role 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 w orkpiece. 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 that transmission 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 applic ations 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 existing single-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。
中间补气螺杆压缩机及其喷油参数的理论分析与研究---优秀毕业论文参考文..
分类号 __ 学号 M********* 学校代码 10487 密级硕士学位论文中间补气螺杆压缩机及其喷油参数的理论分析与研究学位申请人：孙超学科专业：制冷及低温工程指导教师：王盛卫教授陈焕新教授答辩日期：2012年1月5日A Dissertation Submitted in Partial of Fulfillment of the Requirementsfor the Degree of Master EngineeringAnalysis and Research on the Screw Compressor Through Compensating Vapor in the Course ofCompression and Its Oil-injection ParametersCandidate : Sun ChaoMajor : Refrigeration & CryogenicsSupervisor: Prof. Wang ShengweiProf. Chen HuanxinHuazhong University of Science & TechnologyWuhan 430074, P. R. ChinaJanuary, 2012独创性声明本人声明所呈交的学位论文是我个人在导师指导下进行的研究工作及取得的研究成果。
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中文译文4.3 在喷油螺杆压缩机的流量4.3.1 网格生成的油润滑压缩机阳极和阴极的转子有40个数值细胞沿各叶片间的圆周方向，6细胞在径向和轴向方向上的112。
这些形式为转子和壳体444830细胞总数。
为了避免需要增加网格点的数量，如果一个更精确的计算是必需的，一个适应的方法已应用于边界的定义。
时间变化的数量为25，在这种情况下，一个内部循环。
的对阳极的转子转一圈所需的时间步骤的总数是那么125。
在转子中的细胞数为每个时间步长保持相同。
以实现这一目标，一个特殊的网格移动程序开发中的时间通过压缩机转速的确定步骤，正如4章解释。
对于初始时间步长的数值网格图4-15提出。
图4数值网格喷油螺杆压缩机444830细胞4.3.2数学模型的油润滑压缩机数学模型的动量，能量，质量和空间方程问题，如第2.2节所描述的，但一个额外的方程的标量属性油的浓度的增加使石油对整个压缩机性能的影响进行计算。
本构关系是一样的前面的例子。
石油是一种被动的物种在模型处理，这不混合液体-空气的背景。
对空气的影响占通过物质和能量的来源是加上或减去的主要流模型相应的方程。
在这种情况下，动量方程通过拖曳力的影响如前所述。
建立工作条件和从吸气开始全方位1巴压力获得6，7压力的增加，8和9条近450000细胞放电，数值网格对于每一种情况下只有25时间步骤来获得所需的工作条件，其次是进一步的25的时间的步骤来完成一个完整的压缩机循环。
每个时间步所需的约30分钟的运行时间在一个800 MHz的AMD 速龙处理器计算机内存需要约450 MB。
4.3.3对油的数值模拟和实验结果的比较—淹没式压缩机在压缩机中的腔室，在压缩机内的循环的实验得到的压力历史和测得的空气流量和压缩机功率的情况下，测量的速度场担任了宝贵的基础，以验证CFD计算的结果。
要获得这些值，5/6喷油压缩机中，已经描述的，测试安装在压缩机实验室在城市大学伦敦，如图4-16上的钻机。
4-16喷油螺杆空气压缩机5 / 6-128mm（= 90mm）在测试床4.3流的喷油螺杆压缩机该试验台满足螺杆压缩机的接受所有pneurop /程序的要求试验。
压缩机是根据ISO 1706和交付流程测试测定了BS 5600。
高质量的压力传感器测量的压力，与在入口带到压缩机的读数，从压缩机排出和在分离器。
温度是通过热电偶测量FeCo入口和放电从压缩机、油分离器后。
测量透射电子显微镜—温度也被两个，油和冷却水的入口端油冷却器。
从冷却器和压缩机的油流量的计算能量和质量平衡。
通过实验室型转矩仪传感器测量扭矩的IML 色氨酸—500连接发动机和压缩机驱动轴之间。
压缩机是由一个100千瓦的柴油发动机的最大输出驱动，这可能在可变速度操作。
测得的是压缩机的转速频率计、信号转换为电流后，转移到一个数据记录器。
图4-17电脑屏幕上的压气机试验台的测量程序压缩机流量测量到BS 5600与所述的孔板通过压力换能器的PDCR 120/35WL超过压差测量经营范围为0〜200千帕所有相关的脉动量的测量值被用于获得的热力学循环的细节。
这些，在截留容积的压力应用是最重要的，因为它需要绘制机器的PV图。
因此，从开发建设的整个光伏图仅需4离散点在机器外壳的压力变化的记录。
ENDEVCO压阻式传感器，E8180B被用于测量瞬时同时压缩机中的绝对压力值。
每个传感器重新有线的压力在一个叶片空间。
从开始的吸入端， 4反式生产者被定位在所述压缩机壳体的变化记录在每个连续叶片空间。
当绘制顺序，他们给了压力 - 时间整个压缩机工作循环的图。
在两个压缩机的横截面图4-18速度矢量图4 18速度矢量在两个压缩机横截面前截面由不得通过吸入口,底部——截面B-B所有测量值被自动记录和转移到个人电脑通过一个高速InstruNet数据记录器。
数据采集系统启用高速测量的频率以超过2千赫。
收购和测量程序的电脑是写给这在VisualBasic,允许在线测量和计算,压缩机工作参数。
一个电脑屏幕上记录的测量程序给出了图4 17。
在图4-18中，在两个横截面的速度矢量。
其中一个这些是通过进气口和油喷射管，另一个是靠近排出。
图4-19示出了在通过压缩机的垂直截面中的速度。
高的速度值的差距，两者之间的转子和他们的住房和两个转子之间，所产生的尖锐的压力梯度通过的间隙。
这些有清楚区别的速度在叶片间区域其中的流体流动相对缓慢。
引起的流体流有仅由运动的数值网格，这是产生的方式，以跟随的运动在时间上的转子。
最上方的图显示了通过的吸入口和油喷射开口的横截面。
再循环吸入口是巨大的，因为油的位置，似乎是高喷射孔。
如果油注入已进一步向下游的位置，再循环已经减少。
底部的图，它示出了横靠近排放口部分，表明更多的再循环环存在于叶片与较低压力下，如在该图的顶部可见。
在高压区域进行平滑处理的速度相对较低的值，类似的壁的速度在一定程度上。
在轴向截面C-C速度场，它穿过转子沿转子内尖，在图4-19所示图4-19速度矢量在压缩机轴向截面CC平滑的速度是在高压力区域中可见的右端的图像。
在压缩机的上部，其中，低压力和低气压梯度时，流态多弯曲，从而表明流漩涡。
也有在吸入口的远端再循环的同时，在同时，流经端口的轴向的一部分是更密集在截面A-A的油分布和压力场被显示在顶部和底部图分别如图4-20所示。
如前所述，一些流体再循环从工作腔的吸入口通过压缩机间隙。
图4-20表示，与空气一起，油从逸出加压工作腔室的吸入口，通过转子到转子漏路径。
在吸入口的油的存在下也肉眼观察期间这种压缩机的测试。
然而，没有测量，用其制成的。
图4-20截面通过入口和喷油口A-A油顶–质量浓度，底压力分布一些有限的结果，在油分布的实验研究兴等人（2001 ）公布的螺杆式压缩机。
在这种情况下，油流观察到通过使由透明材料制成的压缩机壳体。
虽然作者没有完整地记录了他们的结果，它似乎从什么他们出版的3-D计算所得到的油流模式在他们的实验中获得的那些类似。
在吸入口的热油的存在下，虽然有益的转子的润滑，增加了气体的工作腔室的温度，然后再关闭。
这减少了被困的质量因此压缩机的容量，是另一个的影响不由螺杆压缩机的过程的一维模型，建模。
图4-21显示了在压缩机内的压力分布与阳极转子转速为5000rpm 。
这个数字表示内的压力的每个工作腔几乎是均匀的，并且其可以被视为例如几乎所有的计算和比较。
由于这个原因，所得到的结果的3-D计算可以与从测量得到的那些相比。
图4-21两个转子之间的轴向部分 - 压力分布在工作腔的内压力的变化，如图4-22所示，作为一个阳转子轴角度的功能。
这里的压力轴角图与从压缩机测试结果相比。
结果显示放电的压力是 6，7， 8和9巴绝对压力在轴速度为5000rpm 。
在所有情况下，进气压力为1巴。
预测和之间的协议测量值是合理的，尤其是在压缩过程中。
一些差异被记录在吸入和排出区。
那些在抽吸区域是可能的后果，在图中可见的流量波动4-19 ，这表明，在抽吸过程中的流动和在最开始的压缩还没有这样衰减。
另一方面，压阻式传感器用于测量压力进行在较低的压力更高的错误确保接近零在这些领域的差异，这是。
记录的差异在高压端，在放电过程中，可能产生的被导致的无法捕捉真正geometryaccurately的。
计算出的放电端口简化了从真实的。
它也映射到具有相对低的细胞数。
的计算精度上的网目尺寸的影响是分析在第4.3.5节中更详细地说明。
英文原文The male and female rotors have 40 numerical cells along each interlobe in the circumferential direction, 6 cells in the radial direction and 112 in the axial direction. These form a total number of 444,830 cells for both rotors and the housing.To avoid the need to increase the number of grid points, if a more precise calculation is required, an adaptation method has been applied to the boundary definition.The number of time changes was 25 for one interlobe cycle in this case. The total number of time steps needed for one full rotation of the male rotor is then 125. The number of cells in the rotors was kept the same for each time step. To achieve this, a special grid moving procedure was developed in which the time step was determined by the compressor speed, as explained in Chapter 4. The numerical grid for the initial time step is presented in Figure 4-15.Figure 4-15 Numerical grid for oil injected screw compressor with 444,830 cells4.3.2 Mathematical Model for an Oil-Flooded CompressorThe mathematical model consists of the momentum, energy, mass and space equations, described in section 2.2, but an additional equation for the scalar property of oil concentration was added to enable the influence of oil on the entire com-pressor performance to be calculated.The constitutive relations are the same as in the previous example. The oil is treated in the model as a ‘passive’apry species, which does not mix with the background fluid - air. Its influence on the air is accounted arefor through the energy and mass sources which are added to or subtracted from the appropriate equation of the main flow model. In this case, the momentum equation is affected by drag forces as described earlier.To establish the full range of working conditions and starting from a suction pressure of 1 bar to obtain an increase in pressure of 6, 7, 8 and 9 bars at dis-charge, a numerical mesh of nearlyd450,000 cells was used. For each case only 25 time steps were required to obtain the requiredworking conditions, followed by a further 25 time steps to complete a 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 450 MB.4.3.3 Comparison of the Numerical and Experimental results for an Oil- Flooded CompressoIn the absence of velocity field measurements in the compressor chamber, an experimentally obtained pressure history within the compressor cycle and the measured air flow and compressor power served as a valuable basis to validate the results of the CFD calculation. To obtain these values, the 5/6 oil flooded compressor, already described, was tested on a rig installed in the compressor labo-ratory at City University London, Figure 4-16.Figure 4-16 Oil-Injected air screw compressor 5/6-128mm (a=90mm) in the test bedThe test rig meets all Pneurop/Cagi requirements for screw compressor acceptance tests. The compressor was tested according to ISO 1706 and its delivery flow wasmeasured following BS 5600.The pressures were measured with high quality pressure transducers, with readings taken at the inlet to the compressor, discharge from the compressor andin the separator.The temperatures were measured by FeCo thermocouples at the inlet to and discharge from the compressor and after the oil separator. Measurements of temperature were also taken of both, the oil and the cooling water at the inlet end of the oil cooler. The oil flow rate was calculated from the cooler and compressor energy and mass balances.Torque was measured by a laboratory type torque meter transducer IML TRP500 connected between the engine and the compressor driving shaft. The compressor was driven by a diesel engine prime mover of 100 kW maximum output,which could operate at variable speed. The compressor speed was measured by a frequency meter and the signal was transferred to a data logger after converting to current.Figure 4-17 Computer screen of compressor test rig measuring programThe compressor flow was measured by an orifice plate according to BS 5600 with the differential pressure measured by a pressure transducer PDCR 120/35WL over an operating range of 0-200 kPa.The measured values of all relevant pulsating quantities were used to obtain details of the thermodynamic cycle. Of these, the pressure in the trapped volume was the most significant since it was required to plot the machine p-V diagram. Accordingly, a method was developed to construct an entire p-V diagram from the recording of pressure changes at only 4 discrete points in the machine casing.Endevco piezoresistive transducers E8180B were used to measure the instan-taneous values of the absolute pressure in the compressor. Each transducer re-corded the pressure in one interlobe space. Starting from the suction end, 4 transducers were positioned in the compressor casing to record the changes in each consecutive interlobe space. When plotted in sequence they gave a pressure-time diagram for the whole compressor working cycle.Figure 4-18 Velocity vectors in the two compressor cross sectionsTop – cross section A-A through the suction port, Bottom – cross section B-BAll measured values were automatically logged and transferred to a PC through a high-speed InstruNet data logger. The data acquisition system enabled high speed measurements to be made at frequencies of more then 2 kHz. An acquisition and measuring program for the PC was written for this in Visual Basic that permitted online measurement and calculation of the compressor working parameters. A computer screen record of this measuring program is given in Figure 4-17.In Figure 4-18 the velocity vectors in two cross sections are presented. One of these is through the inlet port and oil injection pipe and the other is close to dis-charge. Figure 4-19 shows the both locities in the vertical section through the com-pressor. High velocity values in the gaps, both between the rotors and their hous-ing and between the two rotors, are generated by the sharp pressure gradients through the clearances. These are clearly distinguished from the velocities in the interlobe regions where the fluid flows relatively slowly. The fluid flow is caused there only by movement of the numerical mesh, which is generated in a manner to follow the movement of the rotors in time.The top diagram shows the cross section through both the suction port and oil injection openings. Recirculation in the suction port is substantial and seems to be high because of the position of the oil injection hole. If the oil injection had been positioned further downstream, the recirculation would have been reduced. The bottom diagram, which shows a cross section close to the discharge port, indicates that more recirculation is present in the lobes with lower pressures, as is visible in the top of the diagram. The velocities in the high pressure regions are smoothed torelatively low values, to some ex-tent similar to the wall velocities.The velocity field in the axial section C-C, which crosses both rotors along therotor bore cusp, is shown in Figure 4-19.Figure 4-19 Velocity vectors in the compressor axial section C-CSmoothing of the velocities is visible in the high pressure regions at the right end of the figure. In the upper portions of the compressor, where both, low pressures and low pressure gradients occur, flow patterns are more curved, thus indicating flow swirls. There is also recirculation in the far end of the suction port while, at the same time, the flow through the axial part of the port is more intensive.The oil distribution and pressure field in the cross section A-A are shown on the top and bottom diagrams of Figure 4-20 respectively. As noted earlier, some fluid recirculates from the working chamber to the suction port through the compressor clearances. Figure 4-20 indicates that together with air, the oil escapes from the pressurised working chamber to the suction port through the rotor-to-rotor leakage paths. The presence of oil in the suction port was also observed visually during tests on this compressor. However, no measurements were made of it.Figure 4-20 Cross section through the inlet port and oil injection port A-ATop –mass concentration of oil, Bottom - Pressure distributionSome limited results of an experimental investigation on oil distribution within a screw compressor are published by Xing et al (2001). In that case, the oil flow was observed by making the compressor casing from a transparent material. Although the authors do not have a complete record of their results, it appears from what they published that the oil flow patterns obtained from the 3-D calculations are similar to those obtained in their experiments. The presence of hot oil in the suction port, although beneficial for the lubrication of the rotors, increases the gas temperature before the working chamber is closed. This reduces the trapped mass and hence the compressor capacity and is another of the effects which are not modelled by one-dimensional models of screw compressor processes.Figure 4-21 shows the pressure distribution within the compressor with a male rotor speed of 5000 rpm. This figure indicates that the pressure within the each working chamber is almost uniform and that it can be regarded as such for almost all calculations and comparisons. Due to that, the results obtained from the 3-D calculations may be compared with those obtained from measurements.Figure 4-21 Axial section between two rotors - Pressure distributionThe change in pressure within the working chamber is shown in Figure 4-22 as a function of the male rotor shaft angle. Here the pressure-shaft angle diagrams are compared with results from the compressor tests. The results shown are for discharge pressures of 6, 7, 8 and 9 bar absolute at a shaft speed of 5000 rpm. In all cases, the inlet pressure was 1 bar. The agreement between the predicted and measured values is reasonable, especially during the compression process. Some differences are recorded in the suction and discharge regions. Those in the suctionegion are probably the consequence of the flow fluctuations visible in Figure 4-19, which shows that the flow during suction and at the very beginning of the compression is not so damped. On the other hand, the piezoresistive transducers used for the measurement of pressure are subjected to a higher error at lower pressure differences, which are close to zero in these areas. The differences recorded at the high pressure end, during the discharge process, are probably generated because of the inability to capture real geometryaccurately. The calculated discharge port was simplified from the real one. It was also mapped with a relatively low number of cells. The influence of the mesh size on the calculation accuracy is ana-lysed in more detail in section 4.3.5。