管殼式換熱器速度場(chǎng)及其振動(dòng)情況分析畢業(yè)論文外文翻譯 (2)

1、<p><b>　　中文3155字</b></p><p>　　Velocity distribution and vibration excitation in tube bundle heat exchangers</p><p>　　Abstract—Design criteria for tube bundle heat exchangers, to

2、 avoid fluidelastic instability, are based on stability criteria for ideal bundles and uniform flow conditions along the tube length. In real heat exchangers, a non-uniform flow distribution is caused by inlet nozzles, i

3、mpingement plates, baffles and bypass gaps. The calculation of the equivalent velocities, according to the extended stability equation of Connors, requires the knowledge of the mode shape and the assumption of a realisti

4、</p><p>　　(1) for equivalent velocity distributions, based on partial constant velocities, and</p><p>　　(2) for the calculation of the critical volume flow in practical design applications. With

5、 computational fluid dynamic (CFD) programs it is possible to calculate the velocity distribution in real tube bundles, and to determine the most endangered tube and thereby the critical volume flow. The paper moreover p

6、resents results and design equations for the inlet section of heat exchangers with variations of a broad range of geometrical parameters, e.g., tube pitch, shell diameter, nozzle diameter, s</p><p>　　INTRODU

7、CTION</p><p>　　For a safe design of real heat exchangers, to avoid damages caused by fluid-elastic instability, the effective velocity distribution over the entire tube length should be known, particularly i

8、n the section with nozzle inlet and in the baffle windows. Up to now only rough assumptions. Computational fluid dynamic (CFD) programs enable the calculation of the flow field in tube bundle heat exchangers . By paramet

9、er studies the influence of the geometry can be investigated. Correlating the calculated</p><p>　　By variation of the inlet geometry, it becomes possible to derive simple correlations for equivalent velocity

10、 distributions and corresponding flow areas in tube bundle heat exchangers. The three-dimensional steady-state flow field on the shell side of heat exchangers with rigid tubes is calculated using the commercial CFD progr

11、am STARCD.</p><p>　　The program solves the well know 3D Navier–Stokes equations for incompressible turbulent flow by using the standard k- model.</p><p>　　INVESTIGATION OF THE INLET SECTION GEOM

12、ETRY</p><p>　　The flow distribution in different inlet sections of tube bundle heat exchangers has been investigated. The tubes first are supported in two fixed bearings, so the support length is equal to th

13、e tube length L. The investigation of a one-pass section is justified, since the velocity distribution in the inlet section is independent of the flow in the following sections of a multi-span heat exchanger; designing r

14、eal heat exchangers. Calculating the steady-state flow field, a constant volume flow ra</p><p>　　and the critical velocity for the 30_ tube array, the second K_.60_/ value is determined with the two average

15、equivalent velocities in transversal flow direction and a critical value ucr:, which is estimated as a linear relation of the critical velocities ucr:.30_/ and ucr:.60_/, depending on the angle of the flow direction. The

16、 basis of this procedure was confirmed by analysing the experimental Figure 2. Stability ratio for the tubes in the first three rows approached by flow in a bundle with a</p><p>　　In figure 1 the endangered

17、tubes for both K_ values in the second tube row approached by flow are shown. Figure 2 shows as an example the K_ values for all Velocity distribution and vibration excitation in tube bundle heat exchangers tubes in the

18、first three tube rows approached by flow at a reduced distance b_ D 0:73, that means, without the tube row No. 1. The volume flow rate was VP D 1:93 m3_s􀀀1.The correction factors cn used in this case are listed

19、in table II.</p><p>　　Experimental data by Jahr [5] show that in homogeneous flow and in ideal bundles, the second row becomes first critical, the first rowonly at about 50%higher throughput, depending on _

20、. The reason is the lower upstream velocity of the first row and thereby a lower force on the tube, even though the gap velocities are the same in the first and the second tube row. The highest K_ values were taken in th

21、e second row. This value determines the value of the critical volume flow rate.The maximum valu</p><p>　　critical volume flow rates are too high. In method B it is assumed that the flow toward the bundle and

22、 the second row occurs only in the nozzle-projection area.</p><p>　　The design by method B achieves values being too low by a factor of 2–3. Surely, these considerable differences for one-through-flow will b

23、ecome lower in real heat exchangers, depending on the number of flow sections. It is the object of this investigation to find a combination of the two methods A and B, getting a safe prediction of the “measured” critical

24、 volume flow rates.</p><p>　　MODEL FOR THE VELOCITY DISTRIBUTION AND THE FLOW AREAS</p><p>　　The model does not describe the true velocity distribution, but the equivalent velocities, i.e. the e

25、xcitation force on themost endangered tube will be approximated. In figure 5, the basic data of the model for determining the distribution of the equivalent velocities and the corresponding flow areas in the second tube

26、row are shown. L is the support length of the tubes and s is length of the chord in the second tube row. The model has been developed and tested for a central position of the inle</p><p>　　maximum flow under

27、 the inlet nozzle with the cross-sectional area Fq1 and the equivalent velocity u1. All other velocities are referred to this highest value, i.e. the velocity ratio lower positive flow with the cross-sectional area Fq2 a

28、nd the velocity ratio possible negative recirculation flow with the cross-sectional area Fq3 and 3 D u3=u1. The flow rate in the partial section III is about 2􀀀10% of the total flow rate, but the equivalent velo

29、city is negligible, since the mode shape function </p><p>　　In order to determine the jet expansion factors X, the reduced volume flows are plotted over all variants of X. Figure 8 shows an example. The sear

30、ched solutions are found for such values of the jet expansion factor, where the volume flow rate, calculated by the model, is equal to the critical volume flow rate, calculated by STAR-CD. There are two solutions, but th

31、e solutions at the lower X values are not always plausible . For this reason the higher X values were taken. In figure 8 it can be ob</p><p>　　CONCLUSION</p><p>　　The presented method produces e

32、quivalent velocity distributions and corresponding cross-sectional areas in real heat exchanger bundles and enables the designer to predict the vibration excitation by fluid-elastic instability more accurate than before.

33、 The derived equations are valid for the inlet section in the second row of normal .30_/ and in the third row of rotated .60_/ triangular arrays of both single- and multispan tubes heat exchangers, considering the influe

34、nce of the different energy r</p><p>　　為了使流體流動(dòng)穩(wěn)定，管殼式換熱器的設(shè)計(jì)要基于穩(wěn)定條件下在理想管道長(zhǎng)度下的勻速流動(dòng)。在實(shí)際生產(chǎn)的換熱器中，流體流動(dòng)不均勻會(huì)造成進(jìn)口噴嘴，撞擊管板及支路產(chǎn)生位移偏差等情況。根據(jù)Connors穩(wěn)定性方程的延伸，得出等效的速度計(jì)算方法，需要了解到換熱器找截流面的速度分布和振動(dòng)情況。本次研究的目標(biāo)是調(diào)查推導(dǎo)出其相關(guān)性及為這種計(jì)算方法提出意見(jiàn)和建

35、議。</p><p>　　基于局部的恒速度，對(duì)應(yīng)著相同的速度分布。</p><p>　　計(jì)算臨界體積流量在實(shí)際設(shè)計(jì)中的應(yīng)用，利用流體流動(dòng)力學(xué)軟件都可以計(jì)算實(shí)際管路中速度場(chǎng)的分布情況，從而確定流動(dòng)的臨界條件，進(jìn)而出臨界的體積流量。其對(duì)換熱器的設(shè)計(jì)過(guò)程中進(jìn)口段得分析及研究結(jié)果應(yīng)用廣泛，包括如下的幾何參數(shù)，如：管殼直徑，噴嘴直徑、跨度、噴管出口與管殼之間的距離等。</p><

36、p><b>　　簡(jiǎn)介</b></p><p>　　根據(jù)換熱器的安全設(shè)計(jì)準(zhǔn)則，為了避免由于流體不穩(wěn)定流動(dòng)造成的損失，整個(gè)管道長(zhǎng)度的流體流速分布都需要計(jì)算出來(lái)，特別是剖面與噴嘴擋板進(jìn)口及出口處。</p><p>　　但是到目前為止只有不成熟的假設(shè)，流體力學(xué)計(jì)算軟件在管殼式換熱器的流場(chǎng)計(jì)算中得到廣泛的應(yīng)用</p><p>　　對(duì)幾何參數(shù)對(duì)相關(guān)的

37、計(jì)算流速和振型函數(shù)的影響進(jìn)行了試驗(yàn)研究，及遵循相關(guān)設(shè)計(jì)標(biāo)準(zhǔn)。</p><p>　　通過(guò)幾何的變化，就能夠通過(guò)管殼式換熱器進(jìn)口參數(shù)推導(dǎo)出簡(jiǎn)單的關(guān)聯(lián)式和相應(yīng)的流量等效速度分布。管殼式換熱器殼程剛性管的流場(chǎng)的三維穩(wěn)態(tài)計(jì)算公式是商用流體力學(xué)計(jì)算軟件的STARCD程序。了解程序中的k-w模型可以解決不可壓縮湍流的三維n-s方程。</p><p>　　入口的界面幾何參數(shù)的研究</p>&

38、lt;p>　　換熱器管束的不同區(qū)域的入口流量的影響因素不同，管子首先是由兩個(gè)固定軸承支撐，所以支撐的管道長(zhǎng)度為L(zhǎng) 。我們所研究的部分是合理的，因?yàn)樗俣确植荚谌肟谔幨仟?dú)立的流量，而這一條件在換熱器其它部分的入口條件是通用的。</p><p>　　在流場(chǎng)計(jì)算條件的穩(wěn)態(tài)下，流體的體積和流量是固定的，為了確定軸向速度在管內(nèi)的分部差距，流體在殼程中的正常條件下，對(duì)每只管殼的縫隙和鄰近管進(jìn)行速度分析，對(duì)每個(gè)缺口運(yùn)用運(yùn)處

39、理的等效速度方程的均方根值得等效速度差距的情況是相反的。這三個(gè)平均等效速度的方法定義兩個(gè)穩(wěn)定的流向比率。第一階30K定義為正常流動(dòng)方向，是三個(gè)使用的方法中最大的等效速度和管殼的臨界速度。它決定了第二階60K與這兩個(gè)平均等效速度的橫向流動(dòng)方向及其臨界值。這是一個(gè)線性的估計(jì)情況，30K和60K取決于流體流動(dòng)的角度。在這個(gè)程序基礎(chǔ)上，通過(guò)圖2可以看出管殼的穩(wěn)定性比前三排困在一起有所降低。</p><p>　　對(duì)所有值K

40、流速分布及振動(dòng)激勵(lì)在管殼式換熱器前三排管束有所降低。實(shí)驗(yàn)研究數(shù)據(jù)表明，流體流動(dòng)是否均勻，主要取決于捆綁化得程度，第二排的流速大約要比第一排高出50%左右。這是因?yàn)樯嫌瘟黧w流動(dòng)的速率要低于第一排，從而獲得較低的流體力。盡管在第一排和第二排之間的管束內(nèi)流體流動(dòng)速率之間的差值相同。第二排的K值能量最高，這絕對(duì)了臨界容積流量。最大的能量值K穩(wěn)定在第一排管束與第二排管束之間。管殼式換熱器的管束布局應(yīng)該避免這種情況的發(fā)生。流體以正三角形的流動(dòng)趨勢(shì)向

41、第二排管束靠近。在b=0.73的情況下，第一排管束的最大實(shí)際計(jì)算值要略微高于第二排管束的最大實(shí)際計(jì)算值，這是由于噴嘴的橫向出口速度。此外，圖三中所顯示的橫向流動(dòng)方向并不明顯。根據(jù)實(shí)際問(wèn)題的研究表明，兩個(gè)線性函數(shù)之間所表達(dá)的穩(wěn)定性之比為2:5.關(guān)于穩(wěn)定性之比2:5的數(shù)據(jù)是經(jīng)過(guò)反復(fù)驗(yàn)證得到的經(jīng)驗(yàn)值。研究報(bào)告在圖四中給出了流體流動(dòng)穩(wěn)定性的臨界體積流動(dòng)速率。這種研究結(jié)果不同于之前的簡(jiǎn)單的設(shè)計(jì)方法，在均勻流場(chǎng)中的有效方法應(yīng)該是截面積法。因?yàn)樵谶@種

42、情況下的預(yù)測(cè)臨界流量的誤差會(huì)很大，所以這種方法通常不被研究人員所采用。在第二種方法中流量捆扎的情況只發(fā)生在no</p><p>　　模型和流動(dòng)速度的分布情況</p><p>　　該模型描述不出真正的流體流動(dòng)的速率分布，特別是等效速度。激振所產(chǎn)生的力對(duì)管束的影響很大?；A(chǔ)數(shù)據(jù)確定的模型等效速度分布區(qū)域和相應(yīng)的流管如圖五所示。支撐長(zhǎng)度為L(zhǎng)，第二排管束的長(zhǎng)度為S。用該模型對(duì)中央位置的進(jìn)口噴嘴進(jìn)行

43、開(kāi)發(fā)和測(cè)試得到一個(gè)對(duì)稱(chēng)陣型的函數(shù)，這是基于流體影響最大的中央管束的噴嘴部分進(jìn)行研究的。三個(gè)流體流動(dòng)部分不斷被分為三個(gè)不同的速度模型。進(jìn)口噴嘴的主要參數(shù)有流體流動(dòng)的最大流量、橫截面積及等效速度。所有的其他的速度總稱(chēng)為最高價(jià)值。例如，積極流速比低流量的截面積和速度之比可能帶來(lái)校級(jí)循環(huán)流量。其中，第三部分的流量是管殼式換熱器流量總數(shù)的10%，但等效速度是可以忽略的，自振函數(shù)幾乎為零。</p><p>　　而通流面積的減

44、小對(duì)換熱器的換熱效率的影響尤為重要。正因如此，在等效速度下，u3被假設(shè)為零。第一個(gè)條件是假設(shè)的，真正的模型速度在不同部分應(yīng)該產(chǎn)生相同的激振影響。例如，圖五所示的等效速度計(jì)算方法考慮了流速分布。如果降低噴嘴入口的直接，就不會(huì)發(fā)生有規(guī)律得流體流量分配。在圖六所示的速度分布在管隙之間，噴嘴的直徑大小與流體流動(dòng)速率成幾何關(guān)系。標(biāo)記線顯示速度分布在受激振影響最大的管束之間的部分。只有與這兩條線相交在中間節(jié)點(diǎn)上，長(zhǎng)度才能充分的看到。在管束間隙中配置

45、文件，二號(hào)和三號(hào)的情況極為相似，但二號(hào)的傳熱效率比三號(hào)的傳熱效率稍微低一些。對(duì)該模型進(jìn)行流體流動(dòng)力學(xué)的計(jì)算方程可以看出，在四號(hào)管束放置外的區(qū)域，表現(xiàn)出顯著較低的速度值，從而進(jìn)一步降低了五號(hào)管束與六號(hào)管束的傳熱效率，結(jié)果為減少管束流體流動(dòng)區(qū)域的長(zhǎng)度如圖七所示。為了避免其再循環(huán)，噴嘴直徑必須大于0.33.以下情況可以通過(guò)減小管殼式換熱器管束中流體流動(dòng)長(zhǎng)度來(lái)獲得：</p><p>　　為了確定射流的擴(kuò)展，體積減小的因素

46、會(huì)造成全流變化。圖八給出了設(shè)計(jì)實(shí)例，找出了搜索解決這些問(wèn)題的關(guān)鍵因素，在射流膨脹體積流量、計(jì)算模型、等效臨界體積流量等。有兩種解決這個(gè)問(wèn)題的方案，但是不同方法的應(yīng)用要對(duì)應(yīng)不同的情況而定。因此，增加噴氣距離管殼式換熱器的膨脹系數(shù)也會(huì)隨之增加。</p><p>　　分析表明，距離越大，X的增加值就越多。由于考慮到影響噴嘴直徑和定義的參數(shù)，距離b是需要設(shè)定最合適的參數(shù)。無(wú)論管殼式換熱器第一排管束的噴嘴入口處流動(dòng)面積是否

47、重疊，必須保證b的距離。通過(guò)這個(gè)函數(shù)而導(dǎo)出的計(jì)算方程，可以計(jì)算出截面尺寸。圖十所示三維流體流動(dòng)速率與二維流體流動(dòng)速率略有差距。在第二計(jì)算方程中，根據(jù)射流擴(kuò)展因子X(jué)的減少而確定合適的b值。如圖十所示，根據(jù)兩種情況的速度比長(zhǎng)度略微有所差異，大于L。第二種情況的管殼式換熱器的換熱系數(shù)在某些條件下具有較大的偏差。然而實(shí)際上，繪制傳熱系數(shù)的計(jì)算值中第二種方法是瀕危管束，因此在相同的橫截面積下他們具有最大的傳熱效率。如圖六所示，換熱器殼程需要具有一

48、定的徑向速度。因此，就必須要采取較低的擬合曲線，從而在降低差異之間，使L=0.5得以實(shí)現(xiàn)。在應(yīng)用模型的計(jì)算中，速度比可以被繪制成一個(gè)線性函數(shù)，臨界流體流量達(dá)到百分之一百的模擬值，</p><p>　　如圖十一所示，流體流動(dòng)提及流量計(jì)算模擬方程可以推導(dǎo)出管殼式換熱器的殼體直徑。關(guān)于速度場(chǎng)得分布，只對(duì)相對(duì)價(jià)值有必要的部分才可以被人知道。為了這個(gè)目的，驗(yàn)證計(jì)算流體力學(xué)軟件的結(jié)果是非常有必要的。</p>&

49、lt;p><b>　　結(jié)語(yǔ)</b></p><p>　　所提出的研究方法可以產(chǎn)生等效的流體流動(dòng)速度場(chǎng)分布和相應(yīng)的截面尺寸，在實(shí)際的換熱器管束中，設(shè)計(jì)師需要估測(cè)出管束內(nèi)的振動(dòng)激動(dòng)，使計(jì)算結(jié)果產(chǎn)生的誤差更小，推導(dǎo)出的計(jì)算方程，有效的管束第二個(gè)入口段得流體流動(dòng)速率情況正常，考慮到不同能量對(duì)換熱器的傳熱系數(shù)的影響，對(duì)多管式換熱器進(jìn)行分不同部分進(jìn)行研究。到目前為止只有中中央位置的進(jìn)口噴嘴進(jìn)行了評(píng)