機(jī)械式回轉(zhuǎn)式擰瓶機(jī)的設(shè)計(jì)及工程分析【說(shuō)明書(shū)+CAD+UG】
機(jī)械式回轉(zhuǎn)式擰瓶機(jī)的設(shè)計(jì)及工程分析【說(shuō)明書(shū)+CAD+UG】,說(shuō)明書(shū)+CAD+UG,機(jī)械式,回轉(zhuǎn),式擰瓶機(jī),設(shè)計(jì),工程,分析,說(shuō)明書(shū),仿單,cad,ug
英文原文
Applications
4.1 Introduction
This chapter demonstrates the scope of the method developed for the three-dimensional analysis of a screw compressor. The CFD package used in this case was COMET developed by ICCM GmbH Hamburg, today a part of CD-Adapco. The analysis of the flow and performance characteristics of a number of types of screw machines is performed to demonstrate a variety of parameters used for grid generation and calculation.
The first example is concerned with a dry air screw compressor. A common compressor casing is used with two alternative pairs of rotors. The rotors have identical overall geometric properties but different lobe profiles. The application of the adaptation technique enables convenient grid generation for geometrically different rotors. The results obtained by three dimensional modelling are compared with those derived from a one-dimensional model, previously verified by comparison with experimental data..The relative advantages of each rotor profile are demonstrated.
The second example shows the application of three dimensional flow analysis to the simulation of an oil injected air compressor. The results, thus obtained, are compared with test results obtained by the authors from a compressor and test rig, designed and built at City University. They are presented in the form of both integral parameters and a p-indicator diagram. Calculations based on the assumptions of the laminar flow are compared to those of turbulent flow. The effect of grid size on the results is also considered and shown here.
The third example gives the analysis of an oil injected compressor in an ammonia refrigeration plant.This utilises the real fluid property subroutines in the process calculations and demonstrates the blow hole area and the leakage flow through the compressor clearances.
The fourth example presents two cases, one of a dry screw compressor to show the influence of thermal expansion of the rotor on screw compressor performance and one of a high pressure oil-flooded screw compressor to show the influence of high pressure loads upon the compressor performance.
4.2 Flow in a Dry Screw Compressor
Dry screw compressors are commonly used to produce pressurised air, free of any oil. A typical example of such a machine, similar in configuration to the compressor modelled, is shown in Figure 4-1. This is a single stage machine with 4 male and 6 female rotor lobes. The male and female rotor outer diameters are 142.380 mm and 135.820 mm respectively, while their centre lines are 108.4 mm apart. The rotor length to main diameter ratio l/d=1.77. Thus, the rotor length is 252.0 mm. The male rotor with wrap angle =248.40 is driven at a speed of 6000 rpm by an electric motor through a gearbox. The male and female rotors are synchronised through timing gears with the same ratio as that of the compressor rotor lobes i.e. 1.5. The female rotor speed is therefore 4000 rpm. The male rotor tip speed is then 44.7m/s, which is a relatively low value for a dry air compressor. The working chamber is sealed from its bearings by a combination of lip and labyrinth seals.
Each rotor is supported by one radial and one axial bearing, on the discharge end, and one radial bearing on the suction end of the compressor. The bearings are loaded by a high frequency force, which varies due to the pressure change within the working chamber. Both radial and axial forces, as well as the torque change with a frequency of 4 times the rotational speed. This corresponds to 400Hz and coincides with the number of working cycles that occur within the compressor per unit time.
Figure 4-1 Cross section of a dry screw compressor
The compressor takes in air from the atmosphere and discharges it to a receiver at a constant output pressure of 3 bar. Although the pressure rise is moderate, leakage through radial gaps of 150 m is substantial. In many studies and modelling ,procedures, volumetric losses are assumed to be a linear function of the cross sectional area and the square root of pressure difference, assuming that the interlobe clearance is kept more or less constant by the synchronising gears. The leakage through the clearances is then proportional to the clearance gap and the length of the leakage line. However, a large clearance gap is needed to prevent contact with the housing caused by rotor deformation due to the pressure and temperature changes within the working chamber. Hence, the only way to reduce leakage is to minimise the length of the sealing line. This can be achieved by careful design of the screw rotor profile. Although minimising,leakage is an important means of improving a screw compressor efficiency, it is not the only one. Another is to increase the flow area between the lobes and thereby increase the compressor flow capacity, thereby reducing the relative effect of leakage. Modern profile generation methods take these various effects into account by means of optimisation procedures which lead to enlargement of the male rotor interlobes and reduction in the female rotor lobes. The female rotor lobes are thereby strengthened and their deformation thus reduced.
To demonstrate the improvements possible from rotor profile optimisation, a three dimensional flow analysis has been carried out for two different rotor profiles within the same compressor casing, as shown in Figure 4-2. Both rotors are of the “N” type and rack generated.
Figure 4-2‘N’ Rotors, Case-1 upper, Case-2 lower
Case 1 is an older design, similar in shape to SRM “D” rotors. Its features imply that there is a large torque on the female rotor, the sealing line is relatively long and the female lobes are relatively weak.
Case 2, shown on the bottom of Figure 4-2, has rotors optimised for operating on dry air. The female rotor is stronger and the male rotor is weaker. This results in higher delivery, a relatively shorter sealing line and less torque on the female rotor. All these features help to improve screw compressor performance.
The results of these two analyses are presented in the form of velocity distributions in the planes defined by cross-sections A-A and B-B, shown in Figure 4-1.
In the case of this study, the effect of rotor profile changes on compressor integral performance parameters can be predicted fairly accurately with one-dimensional models, even if some of the detailed assumptions made in such analytical models are inaccurate. Hence the integral results obtained from the three-dimensional analysis are compared with those from a one-dimensional model.
4.2.1 Grid Generation for a Dry Screw Compressor
In Case-1, the rotors are mapped with 52 numerical cells along the interlobe on the male rotor and 36 cells along each interlobe on the female rotor in the circumferential direction. This gives 208 and 216 numerical cells respectively in the circumferential direction for the male and female rotors. A total of 6 cells in the radial direction and 97 cells in the axial direction is specified for both rotors. This arrangement results in a numerical mesh with 327090 cells for the entire machine. The cross section for the Case-1 rotors is shown in Figure 4-3. The female rotor is relatively thin and has a large radius on the lobe tip. Therefore, it is more easily mapped than in Case-2 where the tip radius is smaller, as shown in Figure 4-4.
Figure 4-3 Cross section through the numerical mesh for Case-1 rotors
The rotors in Case 2 are mapped with 60 cells along the male rotor lobe and 40 cells along the female lobe, which gives 240 cells along both rotors in the circumferential direction. In the radial direction, the rotors are mapped with 6 cells while 111 cells are selected for mapping along the rotor axis. Thus, the entire working chamber for this compressor has 406570 cells. In this case, different mesh sizes are applied and different criteria are chosen for the boundary adaptation of these rotors. The main adaptation criterion selected for the rotors is the local radius curvature with a grid point ratio of 0.3 to obtain the desired quality of distribution along the rotor boundaries. By this means, the more curved rotors are mapped with only a slight increase in the grid size to obtain a reasonable value of the grid aspect ratio. To obtain a similar grid aspect ratio without adaptation, 85 cells would have been required instead of 60 along one interlobe on the female rotor. This would give 510 cells in the circumferential direction on each of rotors. If the number of cells in the radial direction is also increased to be 8 instead of 6 but the number of cells along axis is kept constant, the entire grid would contain more then a million cells which would, in turn, result in a significantly longer calculation time and an increased requirement for computer memory.
Figure 4-4 Cross section through the numerical mesh for Case-2 rotors
4.2.2 Mathematical Model for a Dry Screw Compressor
The mathematical model used is based on the momentum, energy and mass conservation equations as given in Chapter 2. The equation for space law conservation is calculated in the model in order to obtain cell face velocities caused by the mesh movement. The system of equations is closed by Stoke’s, Fourier’s and Fick’s laws and the equation of state for an ideal gas. This defines all the properties needed for the solution of the governing equations.
4.2.3 Comparison of the Two Different Rotor Profiles
The results obtained for both Case 1 and Case 2 compressors are presented here. To establish the full range of working conditions and to obtain an increase of pressure from 1 to 3 bars between the compressor suction and discharge, 15 time steps were required. A further 25 time steps were then needed to complete the 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 400 MB.
In Figure 4-5 the velocity vectors in the cross and axial sections are compared. The top diagram is given for Case-1 rotors and the bottom one for Case-2. As may be seen, the Case 2 rotors realised a smoother velocity distribution than the Case 1 rotors. This may have some advantage and could have increased the compressor adiabatic efficiency by reduction in flow drag losses. In both cases, recirculation within the entrapped working chamber occurs as consequence of the drag forces in the air as shown in the figure. On the other hand, different fluid flow patterns can be observed in the suction port. The velocities within the working chambers and the suction and discharge ports are kept relatively low while the flow through the clearance gaps changes rapidly and easily reaches sonic velocity.
Figure 4-5 Velocity field in the compressor cross section for Case1 and Case2 rotors
Figure 4-6 Velocity field in the compressor axial section for Case1 and Case2 rotors
These differences are confirmed in the view of the vertical compressor section through the female rotor axis, shown in Figure 4-6. In Case 2, lower velocities are achieved not only in the working chamber but also in the suction and discharge ports. In the suction port, this is significant because of the fluid recirculation which appears at the end of the port. This recirculation causes losses which cannot be recovered later in the compression process. Therefore, many compressors are designed with only an axial port instead of both, radial and axial ports. Such a situation reduces suction dynamic losses caused by recirculation but, on the other hand, increases the velocity in the suction chamber which in turn decreases efficiency. Some of these problems can be avoided only by the design of screw compressor rotors with larger lobes and a bigger swept volume and a shape which allows the suction process to be completed more easily. However, rotor profile design based on existing one-dimensional procedures neglects flow variations in the ports and hence is inferior for this purpose. In such cases, only a full three dimensional approach such as this, will be effective.
中文譯文
應(yīng)用
4.1簡(jiǎn)介
本章介紹了對(duì)螺桿壓縮機(jī)的三維分析開(kāi)發(fā)的方法的范圍。在這種情況下，采用由ICCM GmbH Hamburg開(kāi)發(fā)的CFD軟件，現(xiàn)在是CD-Adapco的一部分。對(duì)一定數(shù)量的螺桿機(jī)器的類(lèi)型的流程和性能特性的分析是用來(lái)展示用于柵格一代和演算的各種各樣的參量。
第一個(gè)例子是關(guān)于一個(gè)干螺桿空氣壓縮機(jī)。一個(gè)常見(jiàn)的壓縮機(jī)外殼是使用兩個(gè)可選雙轉(zhuǎn)子。轉(zhuǎn)子具有相同的整體幾何性質(zhì)但是有不同的葉剖面。適應(yīng)技術(shù)的應(yīng)用可以方便使網(wǎng)格生成幾何不同的轉(zhuǎn)子。三維模型得到的結(jié)果與從一個(gè)一維模型獲得的那些比較，以前被核實(shí)與實(shí)驗(yàn)數(shù)據(jù)相比。演示了每個(gè)轉(zhuǎn)子配置文件的相對(duì)優(yōu)勢(shì)。
第二個(gè)例子顯示了三維流動(dòng)分析模擬注入油空氣壓縮機(jī)的應(yīng)用。如此得到的結(jié)果與從壓縮機(jī)的作者和試驗(yàn)臺(tái)，設(shè)計(jì)和建造城市大學(xué)通過(guò)以下方式獲得的測(cè)試結(jié)果進(jìn)行了比較。他們提出了兩個(gè)積分的形式參數(shù)和一個(gè)p-α示意圖。計(jì)算基于的假設(shè)是層流與湍流流動(dòng)的那些進(jìn)行比較。網(wǎng)格尺寸對(duì)計(jì)算結(jié)果的影響也被認(rèn)為是在這里。
第三個(gè)例子給出了油中注入的制冷壓縮機(jī)的分析。這利用了現(xiàn)實(shí)的流體屬性的過(guò)程中計(jì)算的子程序，并演示吹孔區(qū)域和通過(guò)壓縮機(jī)的間隙泄漏流。
第四個(gè)例子呈現(xiàn)兩種情況，一是顯示的干式螺桿壓縮機(jī)的轉(zhuǎn)子的螺桿式壓縮機(jī)的性能，熱膨脹的影響和高壓油沒(méi)螺桿式壓縮機(jī)中的一個(gè)，以顯示的影響高壓負(fù)荷時(shí)壓縮機(jī)的性能。
4.2干燥螺絲壓縮機(jī)的流程
干燥螺絲壓縮機(jī)是常用的生產(chǎn)被加壓的空氣，不需要任何油。這樣機(jī)器的一個(gè)典型的例子，在配置與被塑造的壓縮機(jī)相似，在表4-1顯示。這是一個(gè)有4個(gè)陽(yáng)性和6個(gè)陰性轉(zhuǎn)子葉單級(jí)機(jī)。陽(yáng)性和陰性的轉(zhuǎn)子外直徑分別為142.380毫米和135.820毫米，而他們的中心線(xiàn)108.4毫米。轉(zhuǎn)子長(zhǎng)度的主直徑比L / D = 1.77。因此，轉(zhuǎn)子長(zhǎng)度252毫米。陽(yáng)轉(zhuǎn)子與包角= 248.40在每分鐘6000轉(zhuǎn)的速度驅(qū)動(dòng)，通過(guò)齒輪箱由一個(gè)電動(dòng)馬達(dá)。陽(yáng)性和陰性的轉(zhuǎn)子通過(guò)定時(shí)齒輪同步與壓縮機(jī)轉(zhuǎn)子裂片即1.5的相同比率。因此，陰性的轉(zhuǎn)子轉(zhuǎn)速為每分鐘4000轉(zhuǎn)。陽(yáng)轉(zhuǎn)子葉尖速度然后44.7米/ s，這是相對(duì)低的值，為干燥的空氣壓縮機(jī)。工作腔密封從它的軸承，由唇，迷宮式密封的組合。每個(gè)轉(zhuǎn)子是由一個(gè)徑向和軸向軸承和一個(gè)徑向軸承在放電結(jié)束后吸入端的壓縮機(jī)。軸承是由一個(gè)高頻力加載，它會(huì)因在工作腔的壓力變化而變化。徑向和軸向的力，以及頻率的旋轉(zhuǎn)速度的4倍的轉(zhuǎn)矩變化。這對(duì)應(yīng)于400Hz和發(fā)生在壓縮機(jī)內(nèi)的每單位時(shí)間的工作周期數(shù)一致。
壓縮機(jī)以空氣從大氣排到一個(gè)接收器3個(gè)恒定的輸出壓力。雖然壓力上升是溫和的，經(jīng)過(guò)150徑向間隙泄漏是巨大的。在許多研究和建模過(guò)程中，容積損失被認(rèn)為是一個(gè)線(xiàn)性函數(shù)的橫截面積和壓差的平方根假設(shè)葉片間間隙保持或多或少不變的同步齒輪。然后，通過(guò)該間隙的泄漏間隙和泄漏管路的長(zhǎng)度成比例。然而，一個(gè)大的間隙是必要的，以防止轉(zhuǎn)子變形，由于工作腔內(nèi)的壓力和溫度的變化所造成的與殼體接觸。因此，減少泄漏的唯一方法是將密封線(xiàn)長(zhǎng)度。這可以通過(guò)仔細(xì)的螺桿轉(zhuǎn)子型線(xiàn)設(shè)計(jì)實(shí)現(xiàn)。盡管最小化泄漏是一個(gè)重要的手
圖4-1 干式螺桿壓縮機(jī)的截面
段,提高了螺桿壓縮機(jī)效率,卻不是唯一的一個(gè)。另一個(gè)是提高葉流之間的區(qū)域,從而提高壓氣機(jī)葉流量,從而減少了相對(duì)效應(yīng)的泄漏。現(xiàn)代配置生成方法把這些不同的影響考慮通過(guò)優(yōu)化程序，導(dǎo)致擴(kuò)大陽(yáng)轉(zhuǎn)子葉片和減少陰性轉(zhuǎn)子葉。陰性的轉(zhuǎn)子葉是加強(qiáng)及其變形從而降低。為了證明可能從轉(zhuǎn)子齒形優(yōu)化，改善已進(jìn)行了三維流場(chǎng)計(jì)算在兩個(gè)不同的轉(zhuǎn)子型線(xiàn)在同一個(gè)壓縮機(jī)殼體，如圖4-2所示。生成兩個(gè)轉(zhuǎn)子的“N ”型和機(jī)架。例1是一個(gè)比較老的設(shè)計(jì)，形狀類(lèi)似SRM “D”的轉(zhuǎn)子。它的特點(diǎn)意味著陰轉(zhuǎn)子上，有一個(gè)大的轉(zhuǎn)矩，密封線(xiàn)是比較長(zhǎng)的相對(duì)較弱陰性葉。顯示在圖4-2的底部，例2的轉(zhuǎn)子的優(yōu)化操作在干燥的空氣。陰性的轉(zhuǎn)子是強(qiáng)大而陽(yáng)性的轉(zhuǎn)子是較弱的。這結(jié)果在較高的輸送，一個(gè)相對(duì)較短的密封線(xiàn)和扭矩少陰轉(zhuǎn)子上。所有這些特點(diǎn)有助于提高螺桿壓縮機(jī)的性能。
這兩個(gè)分析結(jié)果中的橫截面定義的平面A-A、B-B速度分布的形式出現(xiàn)，如圖4-1所示。
在本研究的情況下，轉(zhuǎn)子型線(xiàn)的變化對(duì)壓縮機(jī)的整體性能參數(shù)的影響可以相當(dāng)準(zhǔn)確地預(yù)測(cè)的一維模型，即使在這樣的分析模型作了詳細(xì)的假設(shè)是不正確的。因此，從三維分析得到的積分結(jié)果與一維模型的比較。
4.2.1用于干式螺桿壓縮機(jī)的網(wǎng)格生成。
在例1中，轉(zhuǎn)子被映射52個(gè)數(shù)值細(xì)胞沿葉片間的陽(yáng)轉(zhuǎn)子和36個(gè)細(xì)胞沿著每個(gè)葉片間的陰轉(zhuǎn)子的圓周方向。這給出了分別在圓周方向上的208和216的數(shù)值的單元格的陽(yáng)性和陰性的轉(zhuǎn)子?？偣灿?個(gè)細(xì)胞在徑向方向上，并在軸向方向上的97個(gè)細(xì)胞被指定為兩個(gè)轉(zhuǎn)子。這種安排導(dǎo)致整個(gè)機(jī)器327090細(xì)胞的數(shù)值嚙合。例1轉(zhuǎn)子的截面如圖4-3所示。陰性轉(zhuǎn)子比較薄，在葉頂大半徑上。因此，它是更容易比分析映射在尖端半徑越小，如圖4-4所示。
圖4–2 轉(zhuǎn)子'N ' ，例1上，例2下
圖4-3 通過(guò)案例1轉(zhuǎn)子截面數(shù)值網(wǎng)格
在例2中的轉(zhuǎn)子被映射60細(xì)胞沿凸轉(zhuǎn)子突齒40細(xì)胞沿陰性葉瓣這給沿兩個(gè)轉(zhuǎn)子在圓周方向上的240個(gè)細(xì)胞。在徑向方向上，轉(zhuǎn)子被映射到與6個(gè)細(xì)胞，111細(xì)胞被選擇為沿轉(zhuǎn)子軸的映射。因此，該壓縮機(jī)的整個(gè)工作腔有406570個(gè)細(xì)胞。在這種情況下，不同的大小被應(yīng)用，并且這些轉(zhuǎn)子的邊界適應(yīng)不同標(biāo)準(zhǔn)的選擇。轉(zhuǎn)子的主要適應(yīng)選擇的標(biāo)準(zhǔn)是與某個(gè)網(wǎng)格點(diǎn)分布，以獲得所需的質(zhì)量比為0.3，沿轉(zhuǎn)子的邊界的局部曲率半徑。通過(guò)這種方式,更多的彎曲轉(zhuǎn)子映射只有一個(gè)輕微增加網(wǎng)格大小來(lái)獲得一個(gè)合理的價(jià)值網(wǎng)格的長(zhǎng)寬比。為了獲得一個(gè)類(lèi)似85個(gè)細(xì)胞所需要的網(wǎng)格長(zhǎng)寬比，而不是沿著一個(gè)葉片間的60陰轉(zhuǎn)子的。這將給510細(xì)胞在圓周方向上每個(gè)轉(zhuǎn)子。如果細(xì)胞的數(shù)量在徑向方向也增加到8代替6但數(shù)量的細(xì)胞沿軸不變，則整個(gè)網(wǎng)格將包含更多，然后一百萬(wàn)細(xì)胞，從而反過(guò)來(lái)導(dǎo)致計(jì)算時(shí)間大大延長(zhǎng)，增加計(jì)算機(jī)內(nèi)存要求。
通圖4 -4 過(guò)數(shù)值網(wǎng)格橫截面為例2轉(zhuǎn)子
4.2.2干式螺桿壓縮機(jī)的數(shù)學(xué)模型
所用的數(shù)學(xué)模型是基于在2章給出了動(dòng)量，能量和質(zhì)量守恒方程。這個(gè)方程計(jì)算空間的保護(hù)模型是為了獲得細(xì)胞面速度引起的嚙合運(yùn)動(dòng)。方程系統(tǒng)是封閉的斯托克城，傅立葉和菲克的法律和理想氣體狀態(tài)方程。這是定義控制方程解決方案所需的所有屬性。
4.2.3兩個(gè)不同的轉(zhuǎn)子的比較
獲得的結(jié)果對(duì)兩例1和例2壓縮機(jī)介紹如下。建立完整的范圍的工作條件和獲得增加壓力從1到3酒吧的壓縮機(jī)吸、排之間，15次步驟是必需的。一個(gè)進(jìn)一步的25次步驟然后需要完成完整的壓縮機(jī)循環(huán)。每個(gè)時(shí)間步需要大約30分鐘，運(yùn)行時(shí)間在一個(gè)800 MHz的AMD Athlon處理器。計(jì)算機(jī)內(nèi)存要求約400 MB。
圖4-5的速度矢量在十字架和軸向部分進(jìn)行比較。前圖給出案例1和轉(zhuǎn)子底部一個(gè)案例2。如可以看到的，在第2種情況的轉(zhuǎn)子實(shí)現(xiàn)更平滑的速度分布比第1種情況的轉(zhuǎn)子。可以增加壓縮機(jī)絕熱效率，減少流動(dòng)阻力損失。在這兩種情況下，再循環(huán)在裹入工作腔發(fā)生的后果在空中拖曳力如圖。 另一方面，不同的流體流動(dòng)模式可觀(guān)測(cè)到吸入口。工作腔和吸入閥和排出端口的速度范圍內(nèi)保持相對(duì)低的，而流過(guò)的間隙間隙的變化迅速，方便達(dá)到聲速。
圖4-5 壓縮機(jī)截面的例1和案例2轉(zhuǎn)子速度場(chǎng)
圖4-6 壓縮機(jī)軸向部分案例1和案例2轉(zhuǎn)子速度場(chǎng)
這些區(qū)別根據(jù)垂直的壓縮機(jī)部分被證實(shí) 通過(guò)陰性電動(dòng)子軸，顯示在圖4-6上。在例2中，低速度達(dá)到不僅在工作室還在入口及出口的港口。在吸入口，這是很重要的,因?yàn)橐后w再循環(huán),結(jié)束時(shí)出現(xiàn)端口。這個(gè)循環(huán)造成損失是無(wú)法恢復(fù)后的壓縮過(guò)程。因此，許多壓縮機(jī)的設(shè)計(jì)只有一個(gè)軸向港口而不是兩個(gè)港口，徑向和軸向港口。這種情況下減少吸入動(dòng)態(tài)而造成損失的再循環(huán)，但另一方面，增加的速度，吸入腔從而降低效率。這些問(wèn)題中的某些問(wèn)題，可避免僅由螺桿壓縮機(jī)的轉(zhuǎn)子的設(shè)計(jì)中具有較大的葉和更大的掃過(guò)容積的形狀，這使得更容易地完成吸入過(guò)程。然而，電動(dòng)子根據(jù)現(xiàn)有的一維做法的外形設(shè)計(jì)忽略在口岸上的流程變化并且為此下等。在這種情況下,只有一個(gè)完整的三維方法如此，這將是有效的。
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