【機械類畢業(yè)論文中英文對照文獻翻譯】雙螺旋槽端面密封結構參數(shù)的優(yōu)化設計,機械類畢業(yè)論文中英文對照文獻翻譯,機械類,畢業(yè)論文,中英文,對照,對比,比照,文獻,翻譯,雙螺旋,端面,密封,結構,參數(shù),優(yōu)化,設計 Science in China Series E: Technological Sciences ? 2007 SCIENCE IN CHINA PRESS Springer Received January 5, 2005; accepted February 26, 2007 doi: 10.1007/s11431-007-0058-5 ? Corresponding author (email: liuying@) Sci China Ser E-Tech Sci | Aug. 2007 | vol. 50 | no. 4 | 448-453 Optimization design of main parameters for double spiral grooves face seal LIU Zhong, LIU Ying ? (b) assembly of static ring module. 2 Theoretical models 2.1 Optimization model The optimization model includes three elements: optimization targets, optimization variables and constraint conditions. Much less leak is the first requirement for the seal. And next, the stiffness of the fluid film between the sealing rings must be large enough to reduce the possible film pertur- bation and guarantee the sealing stability. So optimization targets should be that the fluid film axial stiffness ( s k ) reaches its maximum when the flow leakage (Q) is the least. The optimization function can be expressed as max ( ) / ssp f kdFdh=? and 3 min ( ) , 12 hp fQ rη ? = ? where h is the film thickness between the static and rotational rings, and sp F is the opening force produced by the fluid film between the sealing gap to open the seal. Optimization variables include the operating conditions (rotating speed, sealing pressure, fluid viscosity) and the structural parameters (outer and inner radius of the sealing rings, ratio of the groove width to the land width, ratio of the sealing dam width to the face width, groove depth, groove number, spiral angle (α), etc). The possible range of the optimization variables is limited. And such limitations are regarded as the constraint conditions of the optimization. The main constraint conditions are 2 mh μ≥ and 0≤α≤90°. At present, the general idea of the optimization method for the single spiral groove face seal is as follows: select a fixed and small value of the gap, and keep one of the parameters of groove depth, groove number, ratio of the groove width to the land width and spiral angle as variable and keep others invariable, and then deduce the variation trend of the sealing performance [4] . Using such method, only the individual effect of every element is considered, and mutual action to sealing performance between all elements is neglected. Doing by this way cause errors, and does not re- flect the real running condition under which the closing force of the face seals is usually a fixed value. Parameters optimization with a fixed closing force should be consistent with practical con- ditions. In addition, the film thickness varies with the optimization variables, so mutual effects of synchronous variations of the optimization variables and the film thickness should be taken into 450 LIU Zhong et al. Sci China Ser E-Tech Sci | Aug. 2007 | vol. 50 | no. 4 | 448-453 consideration. The optimization process is shown in Figure 2. Figure 2 Design frame for parameters optimization. 2.2 Fundamental equations in calculation of sealing performance The governing equation for the force balance is given as follows (Figure 3): 22 2 2 222 2 11 ()() 44 SP m P ππ FFF Ddp dDp=++ ? + ? , (1) where D 1 is the inner radius of the stator, D 2 the outer radius of the stator, d 1 , d 2 the inner and outer radius of the corrugated pipe, F m the closing force produced by compressed corrugated pipe, F p the axial force caused by the fluid pressure passing through the corrugated pipe, p 1 the pressure of the leakage space and p 2 the pressure of the seal space. Figure 3 Sketch map of the force balance in steady-state. The opening force is composed of the static pressure and dynamic pressure of the sealing fluid. The governing equation for the fluid dynamics is a two-dimension steady-state turbulent Reynolds equation 33 1 ()0, 2 r hp rhp r h rk rkr φ ρρω ρ φηφ η φ ?? ?? ???? ? + ?=?? ?? ????? ?? ?? (2) LIU Zhong et al. Sci China Ser E-Tech Sci | Aug. 2007 | vol. 50 | no. 4 | 448-453 451 and its dimensionless form is 33 1 ()0 /12 /12 r HP RHP H RkR RRk R φ ρρ ρ φφ φ ???? ????? + ?Λ =?? ?? ???? , (3) where r is the radius, ρ the fluid density, h the film thickness, η the fluid viscosity, φ the radian, p 0 the pressure at outer radius of the sealing ring, p the dynamic pressure, ω the rotating speed, r 1 the inner radius of the sealing ring, and 0 h the sealing gap; dimensionless radius 1 ,Rrr= dimen- sionless pressure 0 ,P pp= dimensionless gap 0 ;Hhh= k φ and r k the turbulent coeffi- cients, and 22 10 6/ o rphηωΛ= the compressing coefficient. The calculation region of eq. (3) is a cirque between the inner radius and the outer radius of the sealing ring, which can be divided into three parts: inner cirque, outer cirque and middle deep groove in which the fluid pressure is supposed to be equal everywhere. According to the conden- sability of the involved fluid, the fluid density and viscosity in a tiny cell are supposed to be con- sistent, but different among bordering cells. Periodicity of the spiral grooves in the sealing face causes periodic distribution of the fluid pressure. So the pressure of the whole calculation region may be gained by only considering one spiral groove, one land and its adjacent sealing dam. A coordinate switch, ln ,uR= is adopted, which transforms the quadrangle with spiral and arc borders under R φ? coordinate to parallelogram under u φ? coordinate. The finite element analysis method with quadrangle equivalence cell is adopted to calculate fluid pressure in the sealing gap. Two boundary conditions are as follows: 1) In radial direction, at inner and outer radius, the pressure boundary condition is 110 1 /( 1),PP ppRR== == 0001 1( / ).P PRRrr== == 2) In circumferential direction, the periodic pressure boundary condition is (2/,) (,).P zR P Rθ π?+ = Newton-Raphson method with relaxation technique is used to solve discrete eq. (3) in order to gain the pressure distribution in the sealing gap. Then the following steady-state sealing per- formances could be obtained. 1) Opening force , SP A Fpdrφ= ∫∫ (4) where A stands for the whole region of the sealing face, and p the dynamic pressure. 2) Flow leakage 3 . 12 hp Q rη ? = ? (5) 3) The pressure loss through feeding holes 4 128 , QL p d η π Δ= (6) where Q denotes the flow leakage in volume, L the length of feeding holes, and d the diameter of feeding holes. 4) Steady-state axial stiffness of film 452 LIU Zhong et al. Sci China Ser E-Tech Sci | Aug. 2007 | vol. 50 | no. 4 | 448-453 SP /. s kdFdh=? (7) 2.3 Optimization result for the structural parameters of double spiral grooves face seal The influences of the structural parameters on the sealing performance, such as groove depth, groove number, ratio of groove width to land width, spiral angle, are shown in Figure 4 in which 0 /Kkk= is dimensionless axial stiffness of the film, and 0 k is the axial stiffness of a benchmark seal. The parametric optimization is conducted under a specific operating state, 20000n = r/min, 0 1.8p = MPa, the fluid viscosity and closing force for seal are fixed. Figure 4 The influence of main structural parameters on sealing performance. (a) The influence of groove depth; (b) the in- fluence of groove number; (c) the influence of ratio of groove width to land width; (d) The influence of spiral angle. ―■― , Sealing gap (μm); ―▲― , dimensionless film stiffness, K; ―▼― , dimensionless leakage, Q; ―●― , ratio of film stiffness to leakage, K/Q. Figure 4(a) shows that both the sealing gap and leakage increase, but the stiffness of the fluid film decreases as the groove depth grows under the condition of fixed closing force. The dynamic effect is stronger with larger groove depth, but larger groove depth (above 20 μm) tends to infinity without dynamic effect and can only be treated as boundary condition. Only with smaller groove depth (below 20 μm), sealing performance, e.g. film stiffness, flow leakage, is better. Figure 4(b) indicates that the effect of the groove number on the film stiffness is little in a wide range (8―20). Additionally, other optimization results show that the curve of the film stiffness to the groove number varies with different fluid viscosity [5] . Figure 4(c) describes the changing trend of the sealing gap, flow leakage and film stiffness with the increase of the ratio of the groove width to the land width. The dynamic effect along the LIU Zhong et al. Sci China Ser E-Tech Sci | Aug. 2007 | vol. 50 | no. 4 | 448-453 453 circumferential direction of the seal ring grows up with an increase in the ratio of the groove width to the land width, which introduces large sealing gaps and flow leakage. On the contrary, the film stiffness decreases smoothly during this process. Larger film stiffness is obtained when the ratio of the groove width to the land width is between 0.2―0.5. The circular velocity of the seal ring can be decomposed into shear and normal velocity. Let us consider a single calculation cell. The larger the spiral angle is and the smaller the intersection angle between shear velocity and circle velocity is, the stronger the effect of the spiral angle is. But when the spiral angle becomes much larger, the area of the calculation cell will be narrower, and the interference of the flow between adjacent cells will also become more serious which will weaken the dynamic pressure effect under some condition on the contrary. This will obviously appear under the working conditions with high rotating speed and high viscosity. Anyway, the spiral angle and the seal gap influence the performance of the face seal under the fixed closure force condition. Therefore, Figure 4(d) shows that the smaller the spiral angle (below 50 degree), the higher the film axial stiffness and the larger the ratio of the film stiffness to the flow leakage. On the other hand, when the spiral angle is above 60 degrees, the sealing gap is wider. These results suggest that the sealing gap is the main factor exerting the dynamic effects. The bumping effect of the spiral reaches the maximum when the spiral angle is approaching 90 degrees. At the same time, the sealing gap and the flow leakage augment rapidly with the growing dynamic effect. The film axial stiffness and the ratio of the film stiffness to the flow leakage arrive at their maxima with the spiral angle be- tween 75―80 degrees. 4 Conclusions (1) The interaction of main structural parameters and the influence of the seal gap on the per- formance of the double spiral grooves face seal are investigated simultaneously with the fixed closure force. The optimization results of the groove depth, groove number, ratio of the groove width to the land width and spiral angle are obtained under the working condition of high speed, high pressure and ultra-low temperature. These results are different from those of single spiral groove gas face seal. The trend of optimization curves with fixed closure force is also dissimilar with that assuming fixed seal gap. (2) The optimization results show that much better sealing performance can be obtained when the ratio of the groove width to the land width is 0.5 and the spiral angle is about 75 degrees. But the influence of the groove number on the sealing performance is not obvious. (3) A small error is caused by the assumption of parallel sealing gap, which ignores the effect of coning film due to heat and elastic distortion. But it does not affect the reliability of parametric optimization results. 1 Peng X D, Jiao Y R, Ye Z W, et al. The hotspots in mechanical face seals. Gener Mech (in Chinese), 2003, 3: 54―57 2 Liu Z, Liu Y, Liu X F. Static performance analysis of a new double spiral groove face seals. Lubric Engin (in Chinese), 2005, 167: 63―65 3 Zheng X Q, Berard G. Large diameter spiral groove face seal development, PerkinElmer fluid sciences. Centurion Me- chanical Seals, 2000, 107―134 4 Hu D M, Hao M M, Peng X D, et al. Geometry optimization of spiral groove upstream pumping mechanical seals. Lubric Engin (in Chinese), 2003, (1): 35―41 5 Liu Z, Liu Y, Liu X F. Effect of low temperature fluid viscosity on geometry optimization of double spiral groove mechanical seals. Lubric Engin (in Chinese), 2006, 182: 79―81 雙螺旋槽端面密封結構參數(shù)的優(yōu)化設計 劉 忠 劉 瑩 * 劉向鋒 (清華大學摩擦學國家重點實驗室 , 北京 100084) 摘要 在定閉合力假設條件下, 采用有限元方法對工作在高速、高密封腔壓力 和超低溫工況下的一種雙螺旋槽端面密封的主要結構參數(shù), 如: 槽深、槽數(shù)、槽 臺寬比和螺旋角等進行了優(yōu)化設計. 結果表明, 當槽臺寬比為 0.5, 螺旋角約 75° 時, 可以獲得最大的密封油膜剛度, 而槽數(shù)對密封性能的影響并不顯著. 關鍵詞 雙螺旋槽 端面密封 優(yōu)化設計 1 概述 螺旋槽端面密封因其具有低磨損、低泄漏、低能耗等優(yōu)點 , 目前廣泛應用于石油、化工、 航空和航天等領域. 但是隨著高新技術和工業(yè)的高速發(fā)展 , 對流體密封的性能也提出更高的 要求, 如 : 環(huán)保、最大限度降低能耗、長壽命( 或重復使用性) 等 [1] , 因此, 研究和設計新型密封 以滿足高性能要求, 成為流體密封研究領域的新熱點, 其中一種新型雙螺旋槽端面密封結構 (圖 1)受到人們的關注. 它與傳統(tǒng)的單螺旋槽密封相比除具有低泄漏的特點外 , 在高速、高壓等 極端環(huán)境中還具有更好的穩(wěn)定性 (密封膜剛度較大 , 并可以依靠雙邊螺旋槽結構實現(xiàn)對外擾動 的自平衡) [2,3] . 資料表明 [3] , 此類雙螺旋密封結構在國外已經應用于航天、國防等重要領域 . 圖 1 雙螺旋端面密封槽型與裝配結構示意圖 (a) 動環(huán)端面; (b) 靜環(huán)組件 本文作者采用等閉合力條件對雙螺旋端面密封的主要結構參數(shù)進行了優(yōu)化設計 , 為密封 端面槽型設計選擇提供了理論基礎 . 2 密封結構參數(shù)優(yōu)化設計的理論基礎 2.1 優(yōu)化設計模型 優(yōu)化設計模型包括 3 個要素: 目標函數(shù)、設計變量和約束條件 . 螺旋槽端面密封首先要求密封性 , 即泄漏量要小 ; 其次, 應保證密封間隙內的液膜軸向剛 度足夠大 , 這樣才能減小外界干擾帶來的膜厚偏差 , 提高密封運行的穩(wěn)定性 . 因此 , 優(yōu)化目標 選為在泄漏量( Q)較小時的最大液膜軸向剛度 ( s k ). 優(yōu)化的目標函數(shù)可以表示為 max ( ) d / d , ssp f kFh=? 3 min ( ) 12 hp fQ rη ? = ? , 式中 h 為靜環(huán)與動環(huán)之間的液膜間隙 ; sp F 為密封開啟力. 設計變量包括兩類 : 工況參數(shù)( 轉速、壓力比、黏度) 和結構參數(shù)( 內外徑、槽臺寬比、密封 壩寬度比、槽深、槽數(shù)、螺旋角等 ). 設計變量一般不允許任意取值而是受到一些限制, 這些限制稱為約束條件. 優(yōu)化方程的 主要約束條件是: 2 μm,h≥ 090.α