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中文譯文
4.3 在噴油螺桿壓縮機(jī)的流量
4.3.1 網(wǎng)格生成的油潤(rùn)滑壓縮機(jī)
陽(yáng)極和陰極的轉(zhuǎn)子有40個(gè)數(shù)值細(xì)胞沿各葉片間的圓周方向，6細(xì)胞在徑向和軸向方向上的112。這些形式為轉(zhuǎn)子和殼體444830細(xì)胞總數(shù)。為了避免需要增加網(wǎng)格點(diǎn)的數(shù)量，如果一個(gè)更精確的計(jì)算是必需的，一個(gè)適應(yīng)的方法已應(yīng)用于邊界的定義。
時(shí)間變化的數(shù)量為25，在這種情況下，一個(gè)內(nèi)部循環(huán)。的對(duì)陽(yáng)極的轉(zhuǎn)子轉(zhuǎn)一圈所需的時(shí)間步驟的總數(shù)是那么125。在轉(zhuǎn)子中的細(xì)胞數(shù)為每個(gè)時(shí)間步長(zhǎng)保持相同。以實(shí)現(xiàn)這一目標(biāo)，一個(gè)特殊的網(wǎng)格移動(dòng)程序開(kāi)發(fā)中的時(shí)間通過(guò)壓縮機(jī)轉(zhuǎn)速的確定步驟，正如4章解釋。對(duì)于初始時(shí)間步長(zhǎng)的數(shù)值網(wǎng)格圖4-15提出。
圖4數(shù)值網(wǎng)格噴油螺桿壓縮機(jī)444830細(xì)胞
4.3.2數(shù)學(xué)模型的油潤(rùn)滑壓縮機(jī)
數(shù)學(xué)模型的動(dòng)量，能量，質(zhì)量和空間方程問(wèn)題，如第2.2節(jié)所描述的，但一個(gè)額外的方程的標(biāo)量屬性油的濃度的增加使石油對(duì)整個(gè)壓縮機(jī)性能的影響進(jìn)行計(jì)算。本構(gòu)關(guān)系是一樣的前面的例子。石油是一種被動(dòng)的物種在模型處理，這不混合液體-空氣的背景。對(duì)空氣的影響占通過(guò)物質(zhì)和能量的來(lái)源是加上或減去的主要流模型相應(yīng)的方程。在這種情況下，動(dòng)量方程通過(guò)拖曳力的影響如前所述。
建立工作條件和從吸氣開(kāi)始全方位1巴壓力獲得6，7壓力的增加，8和9條近450000細(xì)胞放電，數(shù)值網(wǎng)格對(duì)于每一種情況下只有25時(shí)間步驟來(lái)獲得所需的工作條件，其次是進(jìn)一步的25的時(shí)間的步驟來(lái)完成一個(gè)完整的壓縮機(jī)循環(huán)。每個(gè)時(shí)間步所需的約30分鐘的運(yùn)行時(shí)間在一個(gè)800 MHz的AMD 速龍?zhí)幚砥饔?jì)算機(jī)內(nèi)存需要約450 MB。
4.3.3對(duì)油的數(shù)值模擬和實(shí)驗(yàn)結(jié)果的比較—淹沒(méi)式壓縮機(jī)
在壓縮機(jī)中的腔室，在壓縮機(jī)內(nèi)的循環(huán)的實(shí)驗(yàn)得到的壓力歷史和測(cè)得的空氣流量和壓縮機(jī)功率的情況下，測(cè)量的速度場(chǎng)擔(dān)任了寶貴的基礎(chǔ)，以驗(yàn)證CFD計(jì)算的結(jié)果。要獲得這些值， 5/6噴油壓縮機(jī)中，已經(jīng)描述的，測(cè)試安裝在壓縮機(jī)實(shí)驗(yàn)室在城市大學(xué)倫敦，如圖4-16上的鉆機(jī)。
4-16噴油螺桿空氣壓縮機(jī)5 / 6-128mm（= 90mm）在測(cè)試床
4.3流的噴油螺桿壓縮機(jī)
該試驗(yàn)臺(tái)滿足螺桿壓縮機(jī)的接受所有pneurop /程序的要求試驗(yàn)。壓縮機(jī)是根據(jù)ISO 1706和交付流程測(cè)試測(cè)定了BS 5600。高質(zhì)量的壓力傳感器測(cè)量的壓力，與在入口帶到壓縮機(jī)的讀數(shù)，從壓縮機(jī)排出和在分離器。溫度是通過(guò)熱電偶測(cè)量FeCo入口和放電從壓縮機(jī)、油分離器后。測(cè)量透射電子顯微鏡—溫度也被兩個(gè)，油和冷卻水的入口端油冷卻器。從冷卻器和壓縮機(jī)的油流量的計(jì)算能量和質(zhì)量平衡。通過(guò)實(shí)驗(yàn)室型轉(zhuǎn)矩儀傳感器測(cè)量扭矩的IML色氨酸—500連接發(fā)動(dòng)機(jī)和壓縮機(jī)驅(qū)動(dòng)軸之間。壓縮機(jī)是由一個(gè)100千瓦的柴油發(fā)動(dòng)機(jī)的最大輸出驅(qū)動(dòng)，這可能在可變速度操作。測(cè)得的是壓縮機(jī)的轉(zhuǎn)速頻率計(jì)、信號(hào)轉(zhuǎn)換為電流后，轉(zhuǎn)移到一個(gè)數(shù)據(jù)記錄器。
圖4-17電腦屏幕上的壓氣機(jī)試驗(yàn)臺(tái)的測(cè)量程序
壓縮機(jī)流量測(cè)量到BS 5600與所述的孔板通過(guò)壓力換能器的PDCR 120/35WL超過(guò)壓差測(cè)量經(jīng)營(yíng)范圍為0?200千帕所有相關(guān)的脈動(dòng)量的測(cè)量值被用于獲得的熱力學(xué)循環(huán)的細(xì)節(jié)。
這些，在截留容積的壓力應(yīng)用是最重要的，因?yàn)樗枰L制機(jī)器的PV圖。因此，從開(kāi)發(fā)建設(shè)的整個(gè)光伏圖僅需4離散點(diǎn)在機(jī)器外殼的壓力變化的記錄。ENDEVCO壓阻式傳感器， E8180B被用于測(cè)量瞬時(shí)同時(shí)壓縮機(jī)中的絕對(duì)壓力值。每個(gè)傳感器重新有線的壓力在一個(gè)葉片空間。從開(kāi)始的吸入端， 4反式生產(chǎn)者被定位在所述壓縮機(jī)殼體的變化記錄在每個(gè)連續(xù)葉片空間。當(dāng)繪制順序，他們給了壓力 - 時(shí)間整個(gè)壓縮機(jī)工作循環(huán)的圖。在兩個(gè)壓縮機(jī)的橫截面圖4-18速度矢量
圖4 18速度矢量在兩個(gè)壓縮機(jī)橫截面
前截面由不得通過(guò)吸入口,底部——截面B-B
所有測(cè)量值被自動(dòng)記錄和轉(zhuǎn)移到個(gè)人電腦通過(guò)一個(gè)高速InstruNet數(shù)據(jù)記錄器。 數(shù)據(jù)采集系統(tǒng)啟用高速測(cè)量的頻率以超過(guò)2千赫。 收購(gòu)和測(cè)量程序的電腦是寫給這在Visual Basic,允許在線測(cè)量和計(jì)算,壓縮機(jī)工作參數(shù)。 一個(gè)電腦屏幕上記錄的測(cè)量程序給出了圖4 17。在圖4-18中，在兩個(gè)橫截面的速度矢量。其中一個(gè)這些是通過(guò)進(jìn)氣口和油噴射管，另一個(gè)是靠近排出。圖4-19示出了在通過(guò)壓縮機(jī)的垂直截面中的速度。高的速度值的差距，兩者之間的轉(zhuǎn)子和他們的住房和兩個(gè)轉(zhuǎn)子之間，所產(chǎn)生的尖銳的壓力梯度通過(guò)的間隙。這些有清楚區(qū)別的速度在葉片間區(qū)域其中的流體流動(dòng)相對(duì)緩慢。引起的流體流有僅由運(yùn)動(dòng)的數(shù)值網(wǎng)格，這是產(chǎn)生的方式，以跟隨的運(yùn)動(dòng)在時(shí)間上的轉(zhuǎn)子。最上方的圖顯示了通過(guò)的吸入口和油噴射開(kāi)口的橫截面。再循環(huán)吸入口是巨大的，因?yàn)橛偷奈恢茫坪跏歉邍娚淇?。如果油注入已進(jìn)一步向下游的位置，再循環(huán)已經(jīng)減少。底部的圖，它示出了橫靠近排放口部分，表明更多的再循環(huán)環(huán)存在于葉片與較低壓力下，如在該圖的頂部可見(jiàn)。在高壓區(qū)域進(jìn)行平滑處理的速度相對(duì)較低的值，類似的壁的速度在一定程度上。在軸向截面C-C速度場(chǎng)，它穿過(guò)轉(zhuǎn)子沿轉(zhuǎn)子內(nèi)尖，在圖4-19所示
圖4-19速度矢量在壓縮機(jī)軸向截面CC
平滑的速度是在高壓力區(qū)域中可見(jiàn)的右端的圖像。在壓縮機(jī)的上部，其中，低壓力和低氣壓梯度時(shí)，流態(tài)多彎曲，從而表明流漩渦。也有在吸入口的遠(yuǎn)端再循環(huán)的同時(shí)，在同時(shí)，流經(jīng)端口的軸向的一部分是更密集
在截面A-A的油分布和壓力場(chǎng)被顯示在頂部和底部圖分別如圖4-20所示。如前所述，一些流體再循環(huán)從工作腔的吸入口通過(guò)壓縮機(jī)間隙。圖4-20表示，與空氣一起，油從逸出加壓工作腔室的吸入口，通過(guò)轉(zhuǎn)子到轉(zhuǎn)子漏路徑。在吸入口的油的存在下也肉眼觀察期間這種壓縮機(jī)的測(cè)試。然而，沒(méi)有測(cè)量，用其制成的。
圖4-20截面通過(guò)入口和噴油口A-A油頂–質(zhì)量濃度，底壓力分布
一些有限的結(jié)果，在油分布的實(shí)驗(yàn)研究興等人（2001 ）公布的螺桿式壓縮機(jī)。在這種情況下，油流觀察到通過(guò)使由透明材料制成的壓縮機(jī)殼體。雖然作者沒(méi)有完整地記錄了他們的結(jié)果，它似乎從什么他們出版的3-D計(jì)算所得到的油流模式在他們的實(shí)驗(yàn)中獲得的那些類似。在吸入口的熱油的存在下，雖然有益的轉(zhuǎn)子的潤(rùn)滑，增加了氣體的工作腔室的溫度，然后再關(guān)閉。這減少了被困的質(zhì)量因此壓縮機(jī)的容量，是另一個(gè)的影響不由螺桿壓縮機(jī)的過(guò)程的一維模型，建模。圖4-21顯示了在壓縮機(jī)內(nèi)的壓力分布與陽(yáng)極轉(zhuǎn)子轉(zhuǎn)速為5000rpm 。這個(gè)數(shù)字表示內(nèi)的壓力的每個(gè)工作腔幾乎是均勻的，并且其可以被視為例如幾乎所有的計(jì)算和比較。由于這個(gè)原因，所得到的結(jié)果的3-D計(jì)算可以與從測(cè)量得到的那些相比。
圖4-21兩個(gè)轉(zhuǎn)子之間的軸向部分 - 壓力分布
在工作腔的內(nèi)壓力的變化，如圖4-22所示，作為一個(gè)陽(yáng)轉(zhuǎn)子軸角度的功能。這里的壓力軸角圖與從壓縮機(jī)測(cè)試結(jié)果相比。結(jié)果顯示放電的壓力是 6，7， 8和9巴絕對(duì)壓力在軸速度為5000rpm 。在所有情況下，進(jìn)氣壓力為1巴。預(yù)測(cè)和之間的協(xié)議測(cè)量值是合理的，尤其是在壓縮過(guò)程中。一些差異被記錄在吸入和排出區(qū)。那些在抽吸區(qū)域是可能的后果，在圖中可見(jiàn)的流量波動(dòng)4-19 ，這表明，在抽吸過(guò)程中的流動(dòng)和在最開(kāi)始的壓縮還沒(méi)有這樣衰減。另一方面，壓阻式傳感器用于測(cè)量壓力進(jìn)行在較低的壓力更高的錯(cuò)誤確保接近零在這些領(lǐng)域的差異，這是。記錄的差異在高壓端，在放電過(guò)程中，可能產(chǎn)生的被導(dǎo)致的無(wú)法捕捉真正geometryaccurately的。計(jì)算出的放電端口簡(jiǎn)化了從真實(shí)的。它也映射到具有相對(duì)低的
細(xì)胞數(shù)。的計(jì)算精度上的網(wǎng)目尺寸的影響是分析在第4.3.5節(jié)中更詳細(xì)地說(shuō)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 ‘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 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-D calculations may be compared with those obtained from measurements.
Figure 4-21 Axial section between two rotors - Pressure distribution
The 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