雙螺桿空氣壓縮機結(jié)構(gòu)設(shè)計【全套含11張CAD圖紙】
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附錄A譯文3.1 螺桿壓縮機性能的計算內(nèi)部能量守恒 (3.1)其中是角度的旋轉(zhuǎn)的主旋翼h =h( )的比焓,m =m ( )是質(zhì)量流率p = ( ) ,工作腔的控制體積中的流體壓力, Q = Q( )的流體之間的熱傳遞和壓縮機周圍, V = V ( ) ,壓縮機工作腔中的本地卷。在上述方程中,輸入和輸出的下標(biāo)表示的流體流入及流出。 流體的總焓流入由以下組件: (3.2)其中,下標(biāo)L,G表示泄漏增益SUC ,抽吸條件,和油為石油。 流體總流出焓包括: (3.3)指數(shù)升, l表示泄漏損耗和dis表示放電條件與m顯示表示放電注入的油或其它液體污染的氣體的質(zhì)量流率 右手法側(cè)的能量方程由模型的下列術(shù)語流體和壓縮機的螺桿轉(zhuǎn)子和殼體,并通過它們的周邊,由于氣體的溫度的差異,上述殼體和轉(zhuǎn)子的表面之間的熱交換的傳熱系數(shù)求值表達式= 0.023, RE0占.8 。通過主轉(zhuǎn)子的外徑和內(nèi)徑之間的差異為特征長度的雷諾數(shù)和努塞爾數(shù)。這可能不是用于此目的的最合適的尺寸,但出現(xiàn)的特征長度在0.2的指數(shù)部分的傳熱系數(shù)的表達式,因此,只要它表征壓縮機的體積,它仍然在同一個數(shù)量級,作為其他特征尺寸的影響不大的機器。特征速度為Re數(shù)的計算從本機的質(zhì)量流量和橫截面面積。這里的表面,在其上進行熱交換,以及壁溫,依靠的主旋翼的旋轉(zhuǎn)角度 。上述所表示的商品的大量攝入量和其平均焓由于工作體積的氣體流入的能量增益決定。因此,能量的流入的旋轉(zhuǎn)角變化。在吸入期間,等于氣體進入工作容積帶來的平均氣體焓。3.1.1螺桿壓縮機性能的計算然而,在吸入口關(guān)閉時,一定量的壓縮氣體通過間隙泄漏到壓縮機工作腔 。該氣體的質(zhì)量,以及其焓在氣體泄漏方程的基礎(chǔ)上確定。工作體積充滿了氣體,由于泄漏,只有當(dāng)工作體積周圍的空間中的氣體壓力較高,否則無泄漏,或它是在相反的方向,即從對其他壓力通風(fēng)系統(tǒng)的工作腔??偭魅腱蔬M一步校正的焓的量帶入工作腔注入的油。由于從工作體積的氣體流出的能量損失是指由商品質(zhì)量的流出和平均氣體焓。在工作過程中,這是進入排放氣室,被壓縮的氣體的同時,在擴展的情況下,由于不適當(dāng)?shù)呐懦鰤毫Γ@是通過在較低壓力下工作體積到鄰近的空間的間隙泄漏的氣體。如果工作腔中的壓力低于在排出室,排放口是打開的,該流程將在相反的方向,即從排出氣室進入工作腔。質(zhì)量的變化,有一個負號 其假定的焓等于壓力腔中的平均氣體焓。供給的工作氣體在壓縮過程中的熱力學(xué)表示由術(shù)語PdV d 。這個術(shù)語是從本地的壓力和體積變化率進行評估。后者被定義產(chǎn)生瞬時工作體積和其旋轉(zhuǎn)角度的變化的螺桿運動學(xué)的關(guān)系得到的。事實上,術(shù)語的dV /差d可確定瞬時interlobe區(qū),捕獲和重疊區(qū)域校正。如果油或其它流體注入上述壓縮機的工作腔,油質(zhì)量的流入和其焓應(yīng)包括在流入條款 而事實,盡管在混合物中的油的質(zhì)量分?jǐn)?shù)顯著的體積流率時,其效果是不明顯的,因為油的體積分?jǐn)?shù)通常是非常小的。總流出的流體的質(zhì)量,還包括注入的油,其中的較大部分仍然與工作流體混合。氣體之間的熱傳遞和油滴描述由一個一階微分方程確定。質(zhì)量連續(xù)性方程 (3.4)質(zhì)量連續(xù)性方程 (3.5)質(zhì)量的流出率包括: (3.6)質(zhì)量流率的每一個方程滿足連續(xù)性方程 (3.7)其中W m/s表示流體速度, - 流體密度和A - 流體截面區(qū)域。得到的瞬時密度 = ()被困在控制量與相應(yīng)的瞬時體積V的大小從瞬時的質(zhì)量為m ,密度為 =m/ V 。3.1.2吸氣和排氣口從壓縮機的幾何形狀的橫截面面積A得到的旋轉(zhuǎn)角度,它可以被認(rèn)為是周期函數(shù)。吸氣口區(qū)域被定義為: (3.8)SUC裝置上面的吸氣口開口,并且ASUC的時刻開始的值,0表示為在吸入口的橫截面面積的最大值。 如果未指定不同的旋轉(zhuǎn)角度的基準(zhǔn)值,吸入口關(guān)閉時,假設(shè)在吸管末端 = 0。排放口區(qū)同樣被定義為: (3.9)其中下標(biāo)e表示放電結(jié)束, c表示排出口的橫截面面積的最大值壓縮和ADIS , 0表示結(jié)束。吸入和排出端口流體速度 (3.10)其中,為吸入/排放孔的流量系數(shù),而下標(biāo)1和2表示所考慮的端口的上游和下游 ,在計算機代碼中提供計算,如果H2 H1反向流動。3.1.3氣體泄漏 泄漏量的主要部分是總流速的螺紋機,因此發(fā)揮了重要作用,因為它們影響的過程都影響了壓縮機的質(zhì)量流率或壓縮機送貨,即容積效率和壓縮工作的熱力學(xué)效率。對于實際計算時壓縮機的過程中泄漏的影響,這是方便區(qū)分的兩種類型泄漏,根據(jù)他們的方向方面的工作室:增益和損失的泄漏。增益來自排放氣室,并從相鄰的工作腔室獲得,其中有一個較高的壓力泄漏。虧損泄漏離開吸氣室和鄰近腔室向具有較低的壓力的腔室流動。泄漏速度的計算如下考慮的流體流過的間隙。該過程本質(zhì)上是絕熱的Fanno流。為了簡化計算,該流程是有時被假設(shè)為在恒定的溫度條件下,而不是在等焓。此處出發(fā)從當(dāng)時的絕熱條件下進行分析以差的形式,小的旋轉(zhuǎn)角的變化來表示,即在本模型中,泄漏只有很輕微的影響。本模型只考慮氣體泄漏,沒有嘗試考慮到泄漏的氣 - 液混合物中,可摻入適當(dāng)減少間隙的間隙油膜的影響效果。一個理想化的間隙被假定為具有矩形形狀,并漏出的液體的質(zhì)量流量的連續(xù)性方程所表達: (3.11) = (Re,Ma) ,其中r和w是泄漏氣體的密度和速度,Ag= lgg表示間隙的橫截面面積, lg代表泄漏間隙的長度,封口線,g表示泄漏間隙的寬度或間隙,泄漏流排放系數(shù)。在螺桿式壓縮機:領(lǐng)先的尖端密封線之間形成的主柵極的轉(zhuǎn)子指向尖端和套管,落后的頂端密封線主柵極反向尖端和套管之間形成四個不同的密封線來區(qū)分,前部之間的密封線排出轉(zhuǎn)子正面殼體和轉(zhuǎn)子之間的密封線inter lobe 。所有密封線有間隙差距形成泄漏區(qū)域。此外,葉頂間隙泄漏區(qū)域伴隨著通過吹孔區(qū)。據(jù)的類型和位置的泄漏間隙, 5個不同的泄漏可以被識別,即:通過后前端密封和通過領(lǐng)先和前密封的密封和收益損失。第五,的“ through leakage ”不直接影響在工作腔內(nèi)的過程,而是通過從排放氣室向吸入口。泄漏的氣體速度是來自動量方程,粘液壁的摩擦 (3.12)其中f (Re,Ma)的摩擦系數(shù),這是依賴于雷諾數(shù)和馬赫數(shù), Dg是間隙的有效直徑, Dg 2g和dx的長度增量。從連續(xù)性方程,并假設(shè)T常量來消除壓力的氣體密度,該方程可以被集成在壓力從位置2處的高壓側(cè)到低壓側(cè)的間隙,得到1位: (3.13)其中, = fLg / Dg+ 泄漏流電阻的特點, Lg表示間隙泄漏流方向,f表示摩擦系數(shù)和局部阻力系數(shù),代表間隙長度。 可以評價為每個間隙為一個函數(shù),它的尺寸和形狀和流動特性。 a是聲音的速度。全部程序需要的模式,包括摩擦和阻力系數(shù)的雷諾數(shù)和馬赫數(shù)條件為每個不同的間隙。同樣地,在工作流體的摩擦損耗也可以被定義為方程中的局部摩擦因子和流體的速度相關(guān)的葉尖速度,密度,和小摩擦面積。本模型采用為每個特定的壓縮機的類型和用途,用一個簡單的函數(shù)的值作為輸入?yún)?shù)來確定。這些方程引入到所述壓縮機的型號的泄漏流率計算的每個間隙在當(dāng)?shù)氐男D(zhuǎn)角度為 。3.1.4油或液體注射注射油或其他液體起到潤滑,冷卻,密封的目的,修改螺桿式壓縮機的熱力學(xué)過程中 增大油噴射的影響,下面的段落概述了設(shè)計計算程序。相同的程序,可以應(yīng)用到任何其它液體注射治療。特殊效果,如氣體或其縮合物混合溶解在注入液體或反之亦然應(yīng)單獨核算,如果他們預(yù)期會影響的過程。一個程序納入到模型中的這些現(xiàn)象將概述如下。用一個方便的參數(shù)來定義所注入的油質(zhì)量 - 氣質(zhì)量比,/ ,通過打開的油口流入。這是假設(shè)為均勻分布,從該油流入可以評價為 (3.14)作為輸入?yún)?shù)的油 - 氣質(zhì)量比預(yù)先規(guī)定。注入到壓縮機中的油除了潤滑的主要目的是用于冷卻氣體。為了提高冷卻效率使用將油霧化成細小液滴的噴霧裝置,其中的氣體和油之間的接觸表面增加。通過使用特殊設(shè)計的噴嘴,或通過簡單的高壓噴射的霧化,液滴尺寸的分布可以被定義在對于一個給定的油噴射系統(tǒng)的油 - 氣質(zhì)量流量和速度比。此外,每個不同的目標(biāo)油滴的大小,直到它擊中后面液滴,可以在拉格朗日框架內(nèi)解決,會計重力慣性,阻力,和其他油滴的動力學(xué)方程,每個液滴的轉(zhuǎn)動半徑或套管內(nèi)壁半徑的大小。該溶液的液滴平行動量方程能量方程應(yīng)產(chǎn)生的量與其周圍的氣體的熱交換。在本模型中,一個簡單的程序中,通過氣體的熱交換而與周圍的氣體和油液滴的之間瞬時的熱傳遞由微分方程確定。假設(shè)液滴保留一個球形的形式,與預(yù)定的假設(shè)平均液滴直徑dS的液滴和氣體之間的熱交換可以表示在一個簡單的冷卻條件(),其中敖是液滴表面, ,是從索特平均直徑的液滴和浩是液滴的表面上的表達式確定的傳熱系數(shù)。熱交換必須采取平衡的熱量的變化率或每單位時間的液滴送入量,Q=,其中是油的比熱和下標(biāo)o表示油滴。油滴溫度的變化率可以表示為 (3.15)傳熱系數(shù)可獲得: Nu=2+0.6Re (3.16)方程在兩個時間/角度步驟整合產(chǎn)生新的油滴溫度在每一個新的時間/角度步驟: (3.17)其中, p是在以前的時間步長的油滴溫度, k是液滴的無量綱的時間常數(shù),K = /T = / , = / 的實時恒定的墨滴。對于給定的油滴平均直徑, dS的時間常數(shù)的無量綱的形式 (3.18)進一步假設(shè)派生的液滴溫度來表示油的平均溫度,即,這是進一步用于計算的氣 - 油混合物的焓。上述方法是基于這樣的假設(shè),油液滴的時間常數(shù)小于通過氣體的行進時間之前,擊中的轉(zhuǎn)子或殼體的壁,或到達壓縮機排出口的液滴。這意味著,熱交換是通過氣體在壓縮過程中的微滴行進時間內(nèi)完成。這個先決條件獲得霧化油注入。這將產(chǎn)生足夠小的液滴尺寸,給出了通過選擇一個適當(dāng)?shù)膰娮旖嵌鹊男∫旱蔚臅r間常數(shù),并且,在一定程度上,決定初始油噴霧速度。在獨立計算的溶液液滴動量方程的基礎(chǔ)上,針對不同的液滴的軌跡平均直徑和初始速度進行計算。目前在使用中,大部分螺桿式壓縮機都具有典型的尖之間的20和50米/秒的速度,很好的滿足這個條件是直徑小于50微米的油滴。欲了解更多詳細信息,請參閱Stosic等, 1992年。因為液滴動態(tài)包含一個完整的模型,將包含復(fù)雜的計算機代碼,其結(jié)果將總是依賴于油注入噴嘴的設(shè)計和角度,本計算代碼使用上述的簡單的方法。這被認(rèn)為是完全令人滿意的范圍內(nèi)的不同的壓縮機。輸入?yún)?shù)是只索特平均直徑的油滴, Ds和油性能 - 密度,粘度和比熱。3.1.5計算流體屬性在一個理想的氣體,內(nèi)部熱的氣 - 油混合物中的能量由下式給出: (3.19)式中,R是氣體常數(shù), 為絕熱指數(shù)因此,在壓縮機工作腔中的流體的壓力或溫度,可以顯示計算由輸入油的油溫T的方程為: (3.20)如果k趨于0,即高傳熱系數(shù)或小油滴的條件下,油溫快速接近的氣體溫度在一個真正的氣體的情況下,情況比較復(fù)雜,因為不能明確計算的溫度和壓力。然而,由于內(nèi)部的能量可以表示為溫度的函數(shù)和特定的體積。只有被簡化的計算過程可以通過采用內(nèi)部能量作為一個因變量,而不是焓,通常的做法是:狀態(tài)方程P = F1 (T, V)和特定的內(nèi)部能量U = F2 ( T,V )方程通常脫鉤。因此,溫度可以從已知的特定的內(nèi)能和得到的微分方程的解的特定體積計算。附錄B外文文獻3.1 Calculation of Screw Compressor Performance The Conservation of Internal Energy (3.1) where is angle of rotation of the main rotor, h = h() is specific enthalpy, m = m () is mass flow rate p = p(), fluid pressure in the working chamber control volume, Q = Q(), heat transfer between the fluid and the compressor surrounding, V = V () local volume of the compressor working chamber. In the above equation the subscripts in and out denote the fluid inflow and outflow.The fluid total enthalpy inflow consists of the following components: (3.2) where subscripts l, g denote leakage gain suc, suction conditions, and oil denotes oil. The fluid total outflow enthalpy consists of: (3.3) where indices l, l denote leakage loss and dis denotes the discharge conditions with m dis denoting the discharge mass flow rate of the gas contaminated with the oil or other liquid injected. The right hand side of the energy equation consists of the following terms which are model The heat exchange between the fluid and the compressor screw rotors and casing and through them to the surrounding, due to the difference in temperatures of gas and the casing and rotor surfaces is accounted for by the heat transfer coefficient evaluated from the expression Nu = 0.023 Re0.8. For the characteristic length in the Reynolds and Nusselt number the difference between the outer and inner diameters of the main rotor was adopted. This may not be the most appropriate dimension for this purpose, but the characteristic length appears in the expression for the heat transfer coefficient with the exponent of 0.2 and therefore has little influence as long as it remains within the same order of magnitude as other characteristic dimensions of the machine and as long as it characterizes the compressor size. The characteristic velocity for the Re number is computed from the local mass flow and the cross-sectional area. Here the surface over which the heat is exchanged, as well as the wall temperature, depend on the rotation angle of the main rotor. The energy gain due to the gas inflow into the working volume is represented by the product of the mass intake and its averaged enthalpy. As such, the energy inflow varies with the rotational angle. During the suction period, gas enters the working volume bringing the averaged gas enthalpy, which dominates in the suction chamber. However, during the time when the suction port is closed, a certain amount of the compressed gas leaks into the compressor working chamber through the clearances. The mass of this gas, as well as its enthalpy are determined on the basis of the gas leakage equations. The working volume is filled with gas due to leakage only when the gas pressure in the space around the working volume is higher, otherwise there is no leakage, or it is in the opposite direction, i.e. from the working chamber towards other plenums. The total inflow enthalpy is further corrected by the amount of enthalpy brought into the working chamber by the injected oil. The energy loss due to the gas outflow from the working volume is defined by the product of the mass outflow and its averaged gas enthalpy. During delivery, this is the compressed gas entering the discharge plenum, while, in the case of expansion due to inappropriate discharge pressure, this is the gas which leaks through the clearances from the working volume into the neighbouring space at a lower pressure. If the pressure in the working chamber is lower than that in the discharge chamber and if the discharge port is open, the flow will be in the reverse direction, i.e. from the discharge plenum into the working chamber. The change of mass has a negative signand its assumed enthalpy is equal to the averaged gas enthalpy in the pressure chamber. The thermodynamic work supplied to the gas during the compression process is represented by the term pdV d . This term is evaluated from the local pressure and local volume change rate. The latter is obtained from the relationships defining the screw kinematics which yield the instantaneous working volume and its change with rotation angle. In fact the term dV/d can be identified with the instantaneous interlobe area, corrected for the captured and overlapping areas.If oil or other fluid is injected into the working chamber of the compressor, the oil mass inflow and its enthalpy should be included in the inflow terms. In spite of the fact that the oil mass fraction in the mixture is significant, its effect upon the volume flow rate is only marginal because the oil volume fraction is usually very small. The total fluid mass outflow also includes the injected oil, the greater part of which remains mixed with the working fluid. Heat transfer between the gas and oil droplets is described by a first order differential equation.The Mass Continuity Equation (3.4) The mass inflow rate consists of: (3.5) The mass outflow rate consists of: (3.6) Each of the mass flow rate satisfies the continuity equation (3.7)where wm/s denotes fluid velocity, fluid density and A the flow crosssectionarea. The instantaneous density = () is obtained from the instantaneous mass m trapped in the control volume and the size of the corresponding instantaneous volume V , as = m/V .3.1.2 Suction and Discharge Ports The cross-section area A is obtained from the compressor geometry and it may be considered as a periodic function of the angle of rotation . The suction port area is defined by: (3.8)where suc means the starting value of at the moment of the suction port opening, and Asuc, 0 denotes the maximum value of the suction port crosssection area. The reference value of the rotation angle is assumed at the suction port closing so that suction ends at = 0, if not specified differently. The discharge port area is likewise defined by: (3.9)where subscript e denotes the end of discharge, c denotes the end of compression and Adis, 0 stands for the maximum value of the discharge port crosssectional area. Suction and Discharge Port Fluid Velocities (3.10)where is the suction/discharge orifice flow coefficient, while subscripts 1 and 2 denote the conditions downstream and upstream of the considered port. The provision supplied in the computer code will calculate for a reverse flow if h2 h1.3.1.3 Gas Leakages Leakages in a screw machine amount to a substantial part of the total flow rate and therefore play an important role because they influence the process both by affecting the compressor mass flow rate or compressor delivery, i.e. volumetric efficiency and the thermodynamic efficiency of the compression work. For practical computation of the effects of leakage upon the compressor process, it is convenient to distinguish two types of leakages, according to their direction with regard to the working chamber: gain and loss leakages. The gain leakages come from the discharge plenum and from the neighbouring working chamber which has a higher pressure. The loss leakages leave the chamber towards the suction plenum and to the neighbouring chamber with a lower pressure. Computation of the leakage velocity follows from consideration of the fluid flow through the clearance. The process is essentially adiabatic Fanno-flow. In order to simplify the computation, the flow is is sometimes assumed to be at constant temperature rather than at constant enthalpy. This departure from the prevailing adiabatic conditions has only a marginal influence if the analysis is carried out in differential form, i.e. for the small changes of the rotational angle, as followed in the present model. The present model treats only gas leakage. No attempt is made to account for leakage of a gas-liquid mixture, while the effect of the oil film can be incorporated by an appropriate reduction of the clearance gaps. An idealized clearance gap is assumed to have a rectangular shape and the mass flow of leaking fluid is expressed by the continuity equation: (3.11)where r and w are density and velocity of the leaking gas, Ag = lgg the clearance gap cross-sectional area, lg leakage clearance length, sealing line, g leakage clearance width or gap, = (Re, Ma) the leakage flow discharge coefficient.Four different sealing lines are distinguished in a screw compressor: the leading tip sealing line formed between the main and gate rotor forward tip and casing, the trailing tip sealing line formed between the main and gate reverse tip and casing, the front sealing line between the discharge rotor front and the housing and the interlobe sealing line between the rotors. All sealing lines have clearance gaps which form leakage areas. Additionally, the tip leakage areas are accompanied by blow-hole areas. According to the type and position of leakage clearances, five different leakages can be identified, namely: losses through the trailing tip sealing and front sealing and gains through the leading and front sealing. The fifth, “throughleakage” does not directly affect the process in the working chamber, but it passes through it from the discharge plenum towards the suction port.- 14 -
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