喷油螺杆压缩机的流量分析外文文献翻译/中英文翻译/外文翻译

中文译文  4.3 在喷油螺杆压缩机的流量   4.3.1 网格生成的油润滑压缩机      阳极 和 阴极 的转子有 40个数值细胞沿各叶片间的圆周方向， 6细胞在径向和轴向方向 上的 112。 这些形式为转子和壳体 444830细胞总数。为了避免需要增加网格点的数量，如果一个更精确的计算是必需的，一个适应的方法已应用于边界的定义。      时间变化的数量为 25，在这种情况下，一个内部循环。的对 阳极 的转子转一圈所需的时间步骤的总数是那么 125。在转子中的细胞数为每个时间步长保持相同。以实现这一目标，一个特殊的网格移动程序开发中的 时间通过压缩机转速的确定步骤，正如 4章解释。对于初始时间步长的数值网格图 4-15提出 。                                  图 4数值网格喷油螺杆压缩机 444830细胞  4.3.2数学模型的油润滑压缩机      数学模型的动量，能量，质量和空间方程问题，如第 2.2节所描述的，但一个额外的方程的标量属性油的浓度的增加使石油对整个压缩机性能的影响进行计算 。 本构关系是一样的前面的例子。石油是一种被动的物种在模型处理，这不混合液体 -空气的背景。对空气的影响占通过物质和能量的来源是加上或减 去的主要流模型相应的方程。在这种情况下，动量方程通过拖曳力的影响如前所述。       建立工作条件和从吸气开始全方位 1巴压力获得 6， 7压力的增加， 8和 9条近 450000细胞放电，数值网格对于每一种情况下只有 25时间步骤来获得所需的工作条件，其次是进一步的25的时间的步骤来完成一个完整的压缩机循环。每个时间步所需的约 30分钟的运行时间在一个 800 MHz 的 AMD 速龙 处理器计算机内存需要约 450 MB。  4.3.3对油的数值模拟和实验结果的比较 淹没式压缩机      在压缩机中的腔室，在压缩机内的循环的实验得 到的压力历史和测得的空气流量和压缩机功率的情况下，测量的速度场担任了宝贵的基础，以验证 CFD 计算的结果。要获得这些值，  5/6喷油压缩机中，已经描述的，测试安装在压缩机实验室在城市大学伦敦，如图 4-16上的钻机。   4-16 喷油螺杆空气压缩机 5 / 6-128mm（ = 90mm）在测试床   4.3流的喷油螺杆压缩机      该试验台满足螺杆压缩机的接受所有 pneurop /程序的要求试验。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 Visual Basic,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 cells 4.3.2 Mathematical Model for an Oil-Flooded Compressor  The 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 passiveapry 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   nearly  d450,000 cells was used. For each case only 25 time steps were required to obtain the required working 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 Compresso In 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 bed The 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 program The 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 sections Top cross section A-A through the suction port, Bottom cross section B-B  All 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 to relatively 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 the rotor bore cusp, is shown in Figure 4-19. Figure 4-19 Velocity vectors in the compressor axial section C-C Smoothing 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-A Top  mass concentration of oil, Bottom - Pressure distribution Some 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