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附录A译文3.1 螺杆压缩机性能的计算内部能量守恒 (3.1)其中是角度的旋转的主旋翼h =h（ ）的比焓，m =m （ ）是质量流率p = （ ） ，工作腔的控制体积中的流体压力， Q = Q（ ）的流体之间的热传递和压缩机周围， V = V （ ） ，压缩机工作腔中的本地卷。在上述方程中，输入和输出的下标表示的流体流入及流出。 流体的总焓流入由以下组件： (3.2)其中，下标L，G表示泄漏增益SUC ，抽吸条件，和油为石油。 流体总流出焓包括： (3.3)指数升， l表示泄漏损耗和dis表示放电条件与m显示表示放电注入的油或其它液体污染的气体的质量流率 右手法侧的能量方程由模型的下列术语流体和压缩机的螺杆转子和壳体，并通过它们的周边，由于气体的温度的差异，上述壳体和转子的表面之间的热交换的传热系数求值表达式= 0.023， RE0占.8 。通过主转子的外径和内径之间的差异为特征长度的雷诺数和努塞尔数。这可能不是用于此目的的最合适的尺寸，但出现的特征长度在0.2的指数部分的传热系数的表达式，因此，只要它表征压缩机的体积，它仍然在同一个数量级，作为其他特征尺寸的影响不大的机器。特征速度为Re数的计算从本机的质量流量和横截面面积。这里的表面，在其上进行热交换，以及壁温，依靠的主旋翼的旋转角度 。上述所表示的商品的大量摄入量和其平均焓由于工作体积的气体流入的能量增益决定。因此，能量的流入的旋转角变化。在吸入期间，等于气体进入工作容积带来的平均气体焓。3.1.1螺杆压缩机性能的计算然而，在吸入口关闭时，一定量的压缩气体通过间隙泄漏到压缩机工作腔 。该气体的质量，以及其焓在气体泄漏方程的基础上确定。工作体积充满了气体，由于泄漏，只有当工作体积周围的空间中的气体压力较高，否则无泄漏，或它是在相反的方向，即从对其他压力通风系统的工作腔。总流入焓进一步校正的焓的量带入工作腔注入的油。由于从工作体积的气体流出的能量损失是指由商品质量的流出和平均气体焓。在工作过程中，这是进入排放气室，被压缩的气体的同时，在扩展的情况下，由于不适当的排出压力，这是通过在较低压力下工作体积到邻近的空间的间隙泄漏的气体。如果工作腔中的压力低于在排出室，排放口是打开的，该流程将在相反的方向，即从排出气室进入工作腔。质量的变化，有一个负号 其假定的焓等于压力腔中的平均气体焓。供给的工作气体在压缩过程中的热力学表示由术语PdV d 。这个术语是从本地的压力和体积变化率进行评估。后者被定义产生瞬时工作体积和其旋转角度的变化的螺杆运动学的关系得到的。事实上，术语的dV /差d可确定瞬时interlobe区，捕获和重叠区域校正。如果油或其它流体注入上述压缩机的工作腔，油质量的流入和其焓应包括在流入条款 而事实，尽管在混合物中的油的质量分数显着的体积流率时，其效果是不明显的，因为油的体积分数通常是非常小的。总流出的流体的质量，还包括注入的油，其中的较大部分仍然与工作流体混合。气体之间的热传递和油滴描述由一个一阶微分方程确定。质量连续性方程 (3.4)质量连续性方程 (3.5)质量的流出率包括： (3.6)质量流率的每一个方程满足连续性方程 (3.7)其中W m/s表示流体速度， - 流体密度和A - 流体截面区域。得到的瞬时密度 = （）被困在控制量与相应的瞬时体积V的大小从瞬时的质量为m ，密度为 =m/ V 。3.1.2吸气和排气口从压缩机的几何形状的横截面面积A得到的旋转角度，它可以被认为是周期函数。吸气口区域被定义为： (3.8)SUC装置上面的吸气口开口，并且ASUC的时刻开始的值，0表示为在吸入口的横截面面积的最大值。 如果未指定不同的旋转角度的基准值，吸入口关闭时，假设在吸管末端 = 0。排放口区同样被定义为： （3.9）其中下标e表示放电结束， c表示排出口的横截面面积的最大值压缩和ADIS ， 0表示结束。吸入和排出端口流体速度 （3.10）其中，为吸入/排放孔的流量系数，而下标1和2表示所考虑的端口的上游和下游 ，在计算机代码中提供计算，如果H2 H1反向流动。3.1.3气体泄漏 泄漏量的主要部分是总流速的螺纹机，因此发挥了重要作用，因为它们影响的过程都影响了压缩机的质量流率或压缩机送货，即容积效率和压缩工作的热力学效率。对于实际计算时压缩机的过程中泄漏的影响，这是方便区分的两种类型泄漏，根据他们的方向方面的工作室：增益和损失的泄漏。增益来自排放气室，并从相邻的工作腔室获得，其中有一个较高的压力泄漏。亏损泄漏离开吸气室和邻近腔室向具有较低的压力的腔室流动。泄漏速度的计算如下考虑的流体流过的间隙。该过程本质上是绝热的Fanno流。为了简化计算，该流程是有时被假设为在恒定的温度条件下，而不是在等焓。此处出发从当时的绝热条件下进行分析以差的形式，小的旋转角的变化来表示，即在本模型中，泄漏只有很轻微的影响。本模型只考虑气体泄漏，没有尝试考虑到泄漏的气 - 液混合物中，可掺入适当减少间隙的间隙油膜的影响效果。一个理想化的间隙被假定为具有矩形形状，并漏出的液体的质量流量的连续性方程所表达： (3.11) = （Re,Ma） ，其中r和w是泄漏气体的密度和速度，Ag= lgg表示间隙的横截面面积， lg代表泄漏间隙的长度，封口线，g表示泄漏间隙的宽度或间隙，泄漏流排放系数。在螺杆式压缩机：领先的尖端密封线之间形成的主栅极的转子指向尖端和套管，落后的顶端密封线主栅极反向尖端和套管之间形成四个不同的密封线来区分，前部之间的密封线排出转子正面壳体和转子之间的密封线inter lobe 。所有密封线有间隙差距形成泄漏区域。此外，叶顶间隙泄漏区域伴随着通过吹孔区。据的类型和位置的泄漏间隙， 5个不同的泄漏可以被识别，即：通过后前端密封和通过领先和前密封的密封和收益损失。第五，的“ through leakage ”不直接影响在工作腔内的过程，而是通过从排放气室向吸入口。泄漏的气体速度是来自动量方程，粘液壁的摩擦 （3.12）其中f （Re，Ma）的摩擦系数，这是依赖于雷诺数和马赫数， Dg是间隙的有效直径， Dg 2g和dx的长度增量。从连续性方程，并假设T常量来消除压力的气体密度，该方程可以被集成在压力从位置2处的高压侧到低压侧的间隙，得到1位： (3.13)其中， = fLg / Dg+ 泄漏流电阻的特点， Lg表示间隙泄漏流方向，f表示摩擦系数和局部阻力系数，代表间隙长度。 可以评价为每个间隙为一个函数，它的尺寸和形状和流动特性。 a是声音的速度。全部程序需要的模式，包括摩擦和阻力系数的雷诺数和马赫数条件为每个不同的间隙。同样地，在工作流体的摩擦损耗也可以被定义为方程中的局部摩擦因子和流体的速度相关的叶尖速度，密度，和小摩擦面积。本模型采用为每个特定的压缩机的类型和用途，用一个简单的函数的值作为输入参数来确定。这些方程引入到所述压缩机的型号的泄漏流率计算的每个间隙在当地的旋转角度为 。3.1.4油或液体注射注射油或其他液体起到润滑，冷却，密封的目的，修改螺杆式压缩机的热力学过程中 增大油喷射的影响，下面的段落概述了设计计算程序。相同的程序，可以应用到任何其它液体注射治疗。特殊效果，如气体或其缩合物混合溶解在注入液体或反之亦然应单独核算，如果他们预期会影响的过程。一个程序纳入到模型中的这些现象将概述如下。用一个方便的参数来定义所注入的油质量 - 气质量比，/ ，通过打开的油口流入。这是假设为均匀分布，从该油流入可以评价为 (3.14)作为输入参数的油 - 气质量比预先规定。注入到压缩机中的油除了润滑的主要目的是用于冷却气体。为了提高冷却效率使用将油雾化成细小液滴的喷雾装置，其中的气体和油之间的接触表面增加。通过使用特殊设计的喷嘴，或通过简单的高压喷射的雾化，液滴尺寸的分布可以被定义在对于一个给定的油喷射系统的油 - 气质量流量和速度比。此外，每个不同的目标油滴的大小，直到它击中后面液滴，可以在拉格朗日框架内解决，会计重力惯性，阻力，和其他油滴的动力学方程，每个液滴的转动半径或套管内壁半径的大小。该溶液的液滴平行动量方程能量方程应产生的量与其周围的气体的热交换。在本模型中，一个简单的程序中，通过气体的热交换而与周围的气体和油液滴的之间瞬时的热传递由微分方程确定。假设液滴保留一个球形的形式，与预定的假设平均液滴直径dS的液滴和气体之间的热交换可以表示在一个简单的冷却条件（），其中敖是液滴表面， ，是从索特平均直径的液滴和浩是液滴的表面上的表达式确定的传热系数。热交换必须采取平衡的热量的变化率或每单位时间的液滴送入量，Q=，其中是油的比热和下标o表示油滴。油滴温度的变化率可以表示为 (3.15)传热系数可获得： Nu=2+0.6Re (3.16)方程在两个时间/角度步骤整合产生新的油滴温度在每一个新的时间/角度步骤： (3.17)其中， p是在以前的时间步长的油滴温度， k是液滴的无量纲的时间常数，K = /T = / ， = / 的实时恒定的墨滴。对于给定的油滴平均直径， dS的时间常数的无量纲的形式 (3.18)进一步假设派生的液滴温度来表示油的平均温度，即，这是进一步用于计算的气 - 油混合物的焓。上述方法是基于这样的假设，油液滴的时间常数小于通过气体的行进时间之前，击中的转子或壳体的壁，或到达压缩机排出口的液滴。这意味着，热交换是通过气体在压缩过程中的微滴行进时间内完成。这个先决条件获得雾化油注入。这将产生足够小的液滴尺寸，给出了通过选择一个适当的喷嘴角度的小液滴的时间常数，并且，在一定程度上，决定初始油喷雾速度。在独立计算的溶液液滴动量方程的基础上，针对不同的液滴的轨迹平均直径和初始速度进行计算。目前在使用中，大部分螺杆式压缩机都具有典型的尖之间的20和50米/秒的速度，很好的满足这个条件是直径小于50微米的油滴。欲了解更多详细信息，请参阅Stosic等， 1992年。因为液滴动态包含一个完整的模型，将包含复杂的计算机代码，其结果将总是依赖于油注入喷嘴的设计和角度，本计算代码使用上述的简单的方法。这被认为是完全令人满意的范围内的不同的压缩机。输入参数是只索特平均直径的油滴， Ds和油性能 - 密度，粘度和比热。3.1.5计算流体属性在一个理想的气体，内部热的气 - 油混合物中的能量由下式给出： (3.19)式中，R是气体常数， 为绝热指数因此，在压缩机工作腔中的流体的压力或温度，可以显示计算由输入油的油温T的方程为： (3.20)如果k趋于0，即高传热系数或小油滴的条件下，油温快速接近的气体温度在一个真正的气体的情况下，情况比较复杂，因为不能明确计算的温度和压力。然而，由于内部的能量可以表示为温度的函数和特定的体积。只有被简化的计算过程可以通过采用内部能量作为一个因变量，而不是焓，通常的做法是：状态方程P = F1 （T， V）和特定的内部能量U = F2 （ T，V ）方程通常脱钩。因此，温度可以从已知的特定的内能和得到的微分方程的解的特定体积计算。附录B外文文献3.1 Calculation of Screw Compressor Performance The Conservation of Internal Energy （3.1） where is angle of rotation of the main rotor, h = h() is specific enthalpy, m = m () is mass flow rate p = p(), fluid pressure in the working chamber control volume, Q = Q(), heat transfer between the fluid and the compressor surrounding, V = V () 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: (3.2) where subscripts l, g denote leakage gain suc, suction conditions, and oil denotes oil. The fluid total outflow enthalpy consists of: (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 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, 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 of this 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 signand 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 . 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 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 Equation （3.4） The mass inflow rate consists of: （3.5） The mass outflow rate consists of: （3.6） Each of the mass flow rate satisfies the continuity equation (3.7)where wm/s denotes fluid velocity, fluid density and A the flow crosssectionarea. The instantaneous density = () is obtained from the instantaneous mass m trapped in the control volume and the size of the corresponding instantaneous volume V , as = m/V .3.1.2 Suction and Discharge Ports The cross-section area A is obtained from the compressor geometry and it may be considered as a periodic function of the angle of rotation . The suction port area is defined by: (3.8)where suc means the starting value of at the moment of the suction port opening, and Asuc, 0 denotes the maximum value of the suction port crosssection area. The reference value of the rotation angle is assumed at the suction port closing so that suction ends at = 0, if not specified differently. The discharge port area is likewise defined by: (3.9)where subscript e denotes the end of discharge, c denotes the end of compression and Adis, 0 stands for the maximum value of the discharge port crosssectional area. Suction and Discharge Port Fluid Velocities (3.10)where 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 h2 h1.3.1.3 Gas Leakages Leakages 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: (3.11)where r and w are density and velocity of the leaking gas, Ag = lgg the clearance gap cross-sectional area, lg leakage clearance length, sealing line, g leakage clearance width or gap, = (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