喷油螺杆压缩机的流量分析外文文献翻译

无锡 太湖学院 毕业设计（论文） 外文资料翻译 信机 系 机械工程及自动化 专业 院 （系） ： 信 机 系 专 业 ： 机械工程及自动化 班 级 ： 机械 97 姓 名 ： 陈 浩 学 号 ： 0923807 外文出处 ： 机械专业英语教程 附 件 ： 1.译文； 2.原文； 3.评分表 2013年 5月 20日 英文原文 4.3 Flow in an Oil Injected Screw Compressor Figure 4-27 Comparison of pressure change for turbulent and laminar flow calculations The difference in the compressor flow obtained from laminar and turbulent calcu-lations is presented in Figure 4-28. The mass flows at suction and discharge are given as functions of the shaft angle. On average, 4% higher low is calculated with the turbulent model. The difference was greater at the discharge end of the compressor, both in the mean value and in the amplitude. This agrees with the re-sults obtained from the approximate calculations where turbulent transport through clearances is significant. The difference in flow obtained at the suction end is, on average, less than 3%. This shows that a compressor with a large suc-tion opening has no significant dynamical losses, although turbulence exists in the compressor low pressure domains. It is expected that the difference between the laminar and turbulent flow calculations will be smaller for higher discharge pres-sures and lower compressor speeds. Figure 4-28 Comparison of fluid flow at inlet and exit of screw compressor The integral parameters obtained from both the laminar and turbulent numerical models are presented in Table 4-2. According to these results, it can be concluded that turbulence has some influence on the screw compressor. Its effect is greater at lower pressure ratios and low compressor speeds. More detailed insights into the results obtained from the k-model of turbulence can be found in the following four figures； Figure 4-29 shows the kinetic energy of turbulence. The dissipation rate is presented in Figure 4-30, the turbulent vis-cosity in Figure 4-31 and the dimensionless distance from wall y+ is given in Figure 4-32. Figure 4-29 Kinetic energy of turbulence within the screw compressor 4.3 Flow in an Oil Injected Screw Compressor Figure 4-30 Dissipation rate within the screw compressor Figure 4-31 Turbulent viscosity within the screw compressor Figure 4-32 Dimensionless distances from the wall within the compressor The results in all these diagrams are presented in horizontal sections through the blow hole areas on the suction and discharge side of the compressor, in vertical sections through the rotor axes and in cross sections at suction and discharge. The kinetic energy of turbulence, dissipation, turbulent viscosity and y+ are all high for the lobes exposed to the suction domains. All these gradually die out towards discharge. The dissipation rate is extremely high in the clearance gaps between the rotors, as shown in Figure 4-30, while in the other domains it is significantly lower. On the other hand, y+ is small in the clearance gaps while in the main do-mains at suction it has higher values, as shown in Figure 4-32. 4.3.5 The Influence of the Mesh Size on Calculation Accuracy Most calculations in this book are presented for numerical meshes with an average number of 30 cells along one interlobe and a similar number of time steps selected for the rotor to rotate between two interlobe positions. The numerical mesh for thecompressor in this example consists of about 450,000 cells of which About 322,000 numerical cells define the rotor domains. This was a convenient numberof cells to use with a PC computer with an ATHLON 800 processor and 1GB of RAM, which was used for this study. Although the results obtained on that mesh appeared to be satisfactory and agreed well with the experimental data, an investi-gation of the influence of the mesh size on the calculation accuracy had to be con-ducted. For that reason, two additional meshes were generated for the same com-pressor. A smaller one was generated with 20 points along the rotor interlobe, which gave 190,000 cells on both rotors while the other compressor parts were mapped with almost the same number of cells as originally. The overall number of numerical cells was about 353,000. A lower number of cells on the rotors results in a geometry, which does not follow the rotor shape precisely, and the intercon-nection between rotors would possibly become inappropriate. This number of nu-merical cells is probably the lowest for which reliable results can be obtained. Thelargest numerical mesh generated for this investigation consists of 45 numerical cells along the rotor interlobe. That gave 515,520 cell on the rotors and 637,000 cells for the entire compressor domain. This was the biggest numerical mesh that could be loaded into the available computer memory without disc swapping dur-ing the solution. These three numerical meshes are presented in Figure 4-33 in the cross section perpendicular to the rotor axes. Figure 4-33 Three different mesh sizes for the same compressor The results of the calculations are presented in Figure 4-34 in the form of a pres-sure-angle diagram, and in Figure 4-36 as a discharge flow-angle diagram. The first diagram shows how the calculated working pressures for all three investi-gated mesh sizes agree with the measurements. The lowest number of cells gives the highest pressure in the working chamber and vice versa. As a result of that, the consumed power is changed slightly, from 42 kW obtained with the smallest mesh to slightly less then 41 kW for the largest mesh. The difference between the two is less then 3%. This situation is shown in Figure 4-35. The diagram shows the larg-est difference within the cycle to be in the discharge area of the compressor. Some difference is also visible in the middle area of the diagram which seems to be a consequence of the leakage flows obtained with smaller meshes between the ro-tors. In that area, the mesh is probably too coarse to capture all the oscillations which appear in the flow. Figure 4-34 P-alpha diagrams for three different mesh sizes Figure 4-35 Compressor power calculated with three different mesh sizes 4.3 Flow in an Oil Injected Screw Compressor Figure 4-36 Discharge flow rates for different mesh sizes Figure 4-37 Integral flow rate and Specific power obtained with different mesh sizes Diagrams of discharge flow as a function of rotation angle are given in Figure4-36. The coarser mesh shows less oscillation in the flow then the finer meshes. However, the mean value of the flow remained the same for all three mesh sizes, as shown in Figure 4-37. Specific power is calculated from the values obtained previously. It shows a slight fall in value as the number of computational cells is increased. The results obtained with the three different mesh sizes for the compressor in-vestigated here give the impression that the calculation conducted for the com-pressor on an average size of the mesh with 25 to 30 numerical cells along the ro-tor interlobe is sufficiently accurate. 中文译文 4.3 喷油螺杆压缩机的流量 图 4-27 计算比较湍流和层流压力变化 如图 4-28为在计算吸气和排气的质量流量功能轴角中获得的压缩机流从层流和湍流差异。总体而言，湍流模型比流从层流高 4%，无论是在平均值和振幅，压缩 机的排出端是最大的，通过计算近似结果获得间隙显着的湍流输送的重。在吸入端获得的流量差异的平均值，小于 3。这表明，具有大的吸入端的压缩机吸气开口没有任何显着的动力损失，虽然在压缩机低压域存在湍流。这是预期的层流和湍流之间的差异计算将提高排气压力和减小压缩机速度。 图 4-28 根据流体的流动比较螺杆式压缩机的入口和出口 从层流和湍流数值模型的积分获得的参数，如表 4-2中。根据这些结果，可以得出结论，在湍流的螺杆式压缩机上有一定的影响。其效果是在压力越小，流速越大。从第 k湍流模型获得的结果的更详细的分析，可以 发现在以下四个数字，如图 4-29的湍流的动能。图 4-30，图 4-31动荡对粘度和无量纲距离墙 Y +耗散率，如图 4-32。 图 4-29 螺杆压缩机内的湍流动能 图 4-30 螺杆式压缩机内的损耗率 图 4-31 螺杆压缩机内的湍流粘度 图 4-32从墙壁内压缩机的量纲距离通过吸入阀和排出侧的压缩机的结果列于所有这些图中，在通过转子轴的吸入阀和排出的横截面的垂直剖面上的吹孔区域的水平部分。动荡，耗散，湍流粘度和 y+的动能都是高暴露在吸域叶上，所有这些逐渐消亡走向放电。耗散率非常高，转子之间的间隙差距，如图 4-30所示，而在其他领域，它是显着较低。另一方面，如图 4-32所示， +小的间隙中，在主电源处于吸入它具有较高的值。 4.3.5 网格大小对计算精度的影响 在计算这本书中的大部分平均 30个细胞的数量沿一个和类似用于转子之间旋转两位置的数量的选择步骤啮合。在这个例子中包括约 45万个细胞数值网格，其中约 322,000数字单元格定义转子域。这是用于这项研究为了方便使用的细胞数量与 PC电脑的 Athlon800处理器和 1GB的 RAM，虽然网格上，得到的结果似乎是令人满意的，并与实验数据相同，但在康秀红，杜强，李殿中 ，李依依的调查中，影响网格尺寸的计算精度的到的结果是可靠的。本次调查由 45个数字单元格沿转子的数值 t网。这给了整个压缩机 515,520细胞转子和 637,000细胞领域。这是最大的数值的网格，可以在装入光盘交换过程中溶液没有可用的计算机内存。图 4-33中介绍这在转子轴垂直的截面中的三个数值的啮合。图 4-37获得不同的网目尺寸和比功率的积分流量。图 36中给出的是作为旋转角度的函数的排出流，粗网格显示振荡流，但是，所有三个网目尺寸仍然是流量的平均值，如在图 4