齒輪泵的結(jié)構(gòu)改進(jìn)設(shè)計(jì)含開(kāi)題及16張CAD圖

齒輪泵的結(jié)構(gòu)改進(jìn)設(shè)計(jì)含開(kāi)題及16張CAD圖,齒輪泵,結(jié)構(gòu),改進(jìn),改良,設(shè)計(jì),開(kāi)題,16,cad 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. 中文譯文 4.3 噴油螺桿壓縮機(jī)的流量 圖4-27計(jì)算比較湍流和層流壓力變化 如圖4-28為在計(jì)算吸氣和排氣的質(zhì)量流量功能軸角中獲得的壓縮機(jī)流從層流和湍流差異?？傮w而言，湍流模型比流從層流高4%，無(wú)論是在平均值和振幅，壓縮機(jī)的排出端是最大的，通過(guò)計(jì)算近似結(jié)果獲得間隙顯著的湍流輸送的重。在吸入端獲得的流量差異的平均值，小于3％。這表明，具有大的吸入端的壓縮機(jī)吸氣開(kāi)口沒(méi)有任何顯著的動(dòng)力損失，雖然在壓縮機(jī)低壓域存在湍流。這是預(yù)期的層流和湍流之間的差異計(jì)算將提高排氣壓力和減小壓縮機(jī)速度。 圖4-28根據(jù)流體的流動(dòng)比較螺桿式壓縮機(jī)的入口和出口 從層流和湍流數(shù)值模型的積分獲得的參數(shù)，如表4-2中。根據(jù)這些結(jié)果，可以得出結(jié)論，在湍流的螺桿式壓縮機(jī)上有一定的影響。其效果是在壓力越小，流速越大。從第k湍流模型獲得的結(jié)果的更詳細(xì)的分析，可以發(fā)現(xiàn)在以下四個(gè)數(shù)字，如圖4-29的湍流的動(dòng)能。圖4-30，圖4-31動(dòng)蕩對(duì)粘度和無(wú)量綱距離墻Y +耗散率，如圖4-32。 圖4-29螺桿壓縮機(jī)內(nèi)的湍流動(dòng)能 圖4-30螺桿式壓縮機(jī)內(nèi)的損耗率 圖4-31螺桿壓縮機(jī)內(nèi)的湍流粘度 圖4-32從墻壁內(nèi)壓縮機(jī)的量綱距離通過(guò)吸入閥和排出側(cè)的壓縮機(jī)的結(jié)果列于所有這些圖中，在通過(guò)轉(zhuǎn)子軸的吸入閥和排出的橫截面的垂直剖面上的吹孔區(qū)域的水平部分。動(dòng)蕩，耗散，湍流粘度和y+的動(dòng)能都是高暴露在吸域葉上，所有這些逐漸消亡走向放電。耗散率非常高，轉(zhuǎn)子之間的間隙差距，如圖4-30所示，而在其他領(lǐng)域，它是顯著較低。另一方面，如圖4-32所示，+小的間隙中，在主電源處于吸入它具有較高的值。 4.3.5 網(wǎng)格大小對(duì)計(jì)算精度的影響 在計(jì)算這本書(shū)中的大部分平均30個(gè)細(xì)胞的數(shù)量沿一個(gè)和類(lèi)似用于轉(zhuǎn)子之間旋轉(zhuǎn)兩位置的數(shù)量的選擇步驟嚙合。在這個(gè)例子中包括約45萬(wàn)個(gè)細(xì)胞數(shù)值網(wǎng)格，其中約322,000數(shù)字單元格定義轉(zhuǎn)子域。這是用于這項(xiàng)研究為了方便使用的細(xì)胞數(shù)量與PC電腦的Athlon800處理器和1GB的RAM，雖然網(wǎng)格上，得到的結(jié)果似乎是令人滿(mǎn)意的，并與實(shí)驗(yàn)數(shù)據(jù)相同，但在康秀紅，杜強(qiáng)，李殿中，李依依的調(diào)查中，影響網(wǎng)格尺寸的計(jì)算精度的到的結(jié)果是可靠的。本次調(diào)查由45個(gè)數(shù)字單元格沿轉(zhuǎn)子的數(shù)值t網(wǎng)。這給了整個(gè)壓縮機(jī)515,520細(xì)胞轉(zhuǎn)子和637,000細(xì)胞領(lǐng)域。這是最大的數(shù)值的網(wǎng)格，可以在裝入光盤(pán)交換過(guò)程中溶液沒(méi)有可用的計(jì)算機(jī)內(nèi)存。圖4-33中介紹這在轉(zhuǎn)子軸垂直的截面中的三個(gè)數(shù)值的嚙合。圖4-37獲得不同的網(wǎng)目尺寸和比功率的積分流量。圖36中給出的是作為旋轉(zhuǎn)角度的函數(shù)的排出流，粗網(wǎng)格顯示振蕩流，但是，所有三個(gè)網(wǎng)目尺寸仍然是流量的平均值，如在圖4-37所示，從先前得到的值計(jì)算比功率。它顯示了輕微的下降值，計(jì)算增加的細(xì)胞數(shù)目。得到的結(jié)果是在研究壓縮機(jī)的平均面積為25至30數(shù)值RO涵道。出于這個(gè)原因，產(chǎn)生相同的壓縮機(jī)的是兩個(gè)額外的嚙合。產(chǎn)生一個(gè)較小以20分沿的轉(zhuǎn)子，這兩個(gè)轉(zhuǎn)子給了19萬(wàn)個(gè)細(xì)胞，而其它的壓縮機(jī)部件幾乎相同被最初的細(xì)胞數(shù)量映射。數(shù)值細(xì)胞的總?cè)藬?shù)為353,000左右。在較低的數(shù)字的單元格的幾何形狀，這并不精確地說(shuō)，按照轉(zhuǎn)子的形狀和轉(zhuǎn)子之間的互連，連接在轉(zhuǎn)子上的結(jié)果就可能是不恰當(dāng)?shù)摹＿@個(gè)數(shù)值的細(xì)胞的數(shù)量可能是最低的，得到的結(jié)果是可靠的。本次調(diào)查由45個(gè)數(shù)字單元格沿轉(zhuǎn)子的數(shù)值t網(wǎng)。這給了整個(gè)壓縮機(jī)515,520細(xì)胞轉(zhuǎn)子和637,000細(xì)胞領(lǐng)域。這是最大的數(shù)值的網(wǎng)格，可以在裝入光盤(pán)交換過(guò)程中溶液沒(méi)有可用的計(jì)算機(jī)內(nèi)存。圖4-33中介紹這在轉(zhuǎn)子軸垂直的截面中的三個(gè)數(shù)值的嚙合。 圖4-33網(wǎng)格大小相同的三鐘不同的壓縮機(jī) 在圖4-34中壓力角圖的計(jì)算結(jié)果，圖4-36中的排放流角圖。第一個(gè)圖表顯示如何計(jì)算研究所有三個(gè)門(mén)控網(wǎng)目尺寸的工作壓力。最低的細(xì)胞數(shù)量給出了工作腔的最高壓力，反之亦然。消耗功率略有變化，從42千瓦獲得的最小的最大網(wǎng)格，略小宇1千瓦。兩者之間的差異小于3%。這種情況如圖4-35所示，該圖顯示了在周期內(nèi)所述壓縮機(jī)的排放區(qū)的最大的差異。這些差異也顯示在圖的中間區(qū)域，這是泄露流器RO-小網(wǎng)格之間獲得的結(jié)果。在這方面，可能是網(wǎng)格捕捉太粗以致所有的震蕩出現(xiàn)流動(dòng)。 圖4-34 三種不同網(wǎng)格大小的P-阿爾法圖 圖4-35 三種不同的網(wǎng)格尺寸壓縮機(jī)功率計(jì)算 圖4-36 不同網(wǎng)格尺寸放電流速 圖4-37 獲得的不同網(wǎng)目尺寸和比功率的積分流量 圖36中給出的是作為旋轉(zhuǎn)角度的函數(shù)的排出流，粗網(wǎng)格顯示振蕩流，但是，所有三個(gè)網(wǎng)目尺寸仍然是流量的平均值，如在圖4-37所示，從先前得到的值計(jì)算比功率。它顯示了輕微的下降值，計(jì)算增加的細(xì)胞數(shù)目。得到的結(jié)果是在研究壓縮機(jī)的平均面積為25至30數(shù)值RO-器的細(xì)胞沿網(wǎng)格進(jìn)行計(jì)算三種不同的網(wǎng)目尺寸的壓縮機(jī)是足夠準(zhǔn)確的。