管壳式换热器速度场及其振动情况分析毕业论文外文翻译

1、中文3155字velocity distribution and vibration excitation in tube bundle heat exchangersabstractdesign criteria for tube bundle heat exchangers, to avoid fluidelastic instability, are based on stability criteria for ideal bundles and uniform flow conditions along the tube length. in real heat exchangers

2、, a non-uniform flow distribution is caused by inlet nozzles, impingement 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 realistic velocity distri

3、bution in each flow section of the heat exchanger. it is the object of this investigation to derive simple correlations and recommendations,(1) for equivalent velocity distributions, based on partial constant velocities, and(2) for the calculation of the critical volume flow in practical design appl

4、ications. with 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 presents results and design equations for the inlet section of heat

5、 exchangers with variations of a broad range of geometrical parameters, e.g., tube pitch, shell diameter, nozzle diameter, span width, distance between nozzle exit and tube bundle.introductionfor a safe design of real heat exchangers, to avoid damages caused by fluid-elastic instability, the effecti

6、ve velocity distribution over the entire tube length should be known, particularly in 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 par

7、ameter studies the influence of the geometry can be investigated. correlating the calculated velocities with the mode shape function, and regarding the design criteria accepted for ideal bundles, the vibration excitation can be simulated for each tube in the complex geometry, described by the stabil

8、ity ratio k_, as defined in equation.by variation of the inlet geometry, it becomes possible to derive simple correlations for equivalent velocity distributions and corresponding flow areas in tube bundle heat exchangers. the three-dimensional steady-state flow field on the shell side of heat exchan

9、gers with rigid tubes is calculated using the commercial cfd program starcd.the program solves the well know 3d navierstokes equations for incompressible turbulent flow by using the standard k- model.investigation of the inlet section geometrythe flow distribution in different inlet sections of tube

10、 bundle heat exchangers has been investigated. the tubes first are supported in two fixed bearings, so the support length is equal to the 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 follow

11、ing sections of a multi-span heat exchanger; designing real heat exchangers. calculating the steady-state flow field, a constant volume flow rate vp was fixed, in order to determine the axial velocity distribution in the tube gaps. the velocities in the six gaps of each tube with the neighbouring tu

12、bes are analysed. the fluid at the shell-side is air at normal conditions. by applying the extended connors equation, the equivalent velocities for each gap are achieved. the root mean square values of the equivalent velocities of the opposite gaps are determined. with these three average equivalent

13、 velocities it is possible to define the approach flow direction and two stability ratios: the first k_.30_/ value is defined for the normal approach flow direction, using the maximum of the three equivalent velocitiesand the critical velocity for the 30_ tube array, the second k_.60_/ value is dete

14、rmined with the two average 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 basis of this procedure was confirmed by analysing

15、the experimental figure 2. stability ratio for the tubes in the first three rows approached by flow in a bundle with a reduced distance b_ d 0:73 and a volume flow rate vp d 1:93 m3.in figure 1 the endangered tubes for both k_ values in the second tube row approached by flow are shown. figure 2 show

16、s as an example the k_ values for all velocity distribution and vibration excitation in tube bundle heat exchangers tubes in the 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_s1.the correction f

17、actors cn used in this case are listed in table ii.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 _ . the reason is the lower upstream velocity of the first row and t

18、hereby 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 the second row. this value determines the value of the critical volume flow rate.the maximum values of the stability ratios k_ of the tubes in the f

19、irst and the second actual tube row are plotted in figure 3 as a function of the reduced distance b_ for the described bundles. the highest value of k_ d 0:8 appears for normal triangular flow direction (k_.30_/) on the tube number 1 of tube row no. 2, when the shell is completely filled out with tu

20、bes. this tube layout should be avoided. the highest k_ values are achieved for the normal triangular flow direction in the second tube row approached by flow. only at b_ d 0:73, the calculated k_ value in the first actual tube row is a little bit higher than the maximum value in the second actual t

21、ube row, and the k_.60_) values get up to the value of k_max; that is due to the peak transversal velocity at the nozzle exit. moreover, in figure 3 it is shown that the transversal flow direction is not critical. this is true for all investigated bundles with a pitch ratio of _ d 1:28. two nearly l

22、inear functions for the stability ratio between 0 _ b_ _ 2:5 and for b_ 2:5 can be determined. the value of about b_ d 2:5 seems to be a good choice, but b_ should not be lower than 1.the results for the k_.60_/ values of the other investigated pitch ratios will be presented in 11. in the further se

23、ctions of the paper all results are presented for the calculated stability ratios of the normal triangular flow direction k_.30_/. in figure 4 the critical volume flow rates, calculated by the described method with star-cd, are plotted over the reduced distance b_. these results are compared with tw

24、o different simple design methods. in method a a uniform flow in the cross-sectional area is supposed. that is not admissible in this case, because the predictedcritical volume flow rates are too high. in method b it is assumed that the flow toward the bundle and the second row occurs only in the no

25、zzle-projection area.the design by method b achieves values being too low by a factor of 23. surely, these considerable differences for one-through-flow will become lower in real heat exchangers, depending on the number of flow sections. it is the object of this investigation to find a combination o

26、f the two methods a and b, getting a safe prediction of the “measured” critical volume flow rates.model for the velocity distribution and the flow areasthe model does not describe the true velocity distribution, but the equivalent velocities, i.e. the excitation force on themost endangered tube will

27、 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 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 dev

28、eloped and tested for a central position of the inlet nozzle and for a symmetrical mode shape function. it assumes that the most endangered tube is located in the center of the nozzle. three flow sections with partial constant velocities have been distinguished in the model:maximum flow under the in

29、let 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 and the velocity ratio possible negative recirculation flow with the cross-sectional are

30、a fq3 and 3 d u3=u1. the flow rate in the partial section iii is about 210% of the total flow rate, but the equivalent velocity is negligible, since the mode shape function _ is nearly zero. the reduction of the flow area by fq3 is more important. so, the equivalent velocity u3 was assumed to be zer

31、o. the first condition is the assumption, that the true velocities u.z/ and the model values ur should produce the same excitation force in each partial section r. for example, the equivalent velocity u1 in figure 5 is calculated by taking into account the velocity distribution ucr:.z/ and the mode

32、shape function _.z/ only in the partial section i along the length l1: if the reduced nozzle diameter ds=l is high, then no recirculation occurs and l2 d l. the length l2 can be determined by the “measured” flow distribution. in figure 6 the velocity distribution in the tube gaps, refered to the vel

33、ocity in the nozzle vs, for one geometry is illustrated. the marked line shows the velocity distribution for the most endangered tube in the tube gap no. 1. at the intersection of this line with the neutral point line, the length l2 can be found. the profiles of the velocities in the tube gaps nos.

34、2 and 3 are very similar; the values are negligibly lower. the tube gap no. 4, placed outside the nozzle-projection area shows significant lower velocity values, which reduce further for the tube gaps nos. 5 and 6. figure 6 contains also the values of the length l1=l and the equivalent velocities u1

35、=vs and u2=vs, calculated by the model equations. the results for the reduced flow length l2=l of all investigated tube bundles are illustrated in figure 7. to avoid recirculation, the nozzle diameters have to exceed values ds _ .0:33 to 0:5/l, depending on the pitch ratio . the following functions

36、for the reduced length l2 could be obtained: 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 searched solutions are found for such values of the jet expansion factor, where the volume flow rate, calculated

37、 by the model, is equal to the critical volume flow rate, calculated by star-cd. there are two solutions, but the 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 observed that near the minima the influence of the jet ex

38、pansion factor is not significant, except the noncommendable case b_ d 0:375. the curves are clear graduated, so, increasing the reduced distance b_, the jet expansion factor also increases. the analysis showed that the distance a was not suitable correlating the values of x. the distance b is the a

39、ppropriate parameter considering the influence of the nozzle diameter and to define, whether the first row overlaps the nozzle flow area, which should be ensured choosing b_ _ 1:0. in figure 9 all values of the jet expansion factor x for dm d 600 and 700 mm are plotted over the reduced distance b_.

40、for a safe design the following function can be define (20) for values b_ 0:4, i.e. those shells, which are completely filled with tubes, there is x d 1:0. by this function and by the derived equations, the cross-sectional areas fq1 and fq2 can be calculated. in figure 10 the velocity ratio 2 d u2=u

41、1 is plotted over the reduced difference length .l_2 l1/=l; l_2 being calculated by equation (15), also when l_2 should be figure 9. jet expansion factor x depending on the reduced distance b_. figure 10. velocity ratio 2 depending on the reduced difference length. greater than l. the values of 2 sh

42、ow larger deviations and in some cases a dependence on b_. in reality, the plotted values 2 are the values calculated for the most endangered tube, therefore they are the maximum values in the cross-sectional area fq2. considering the radial velocity distribution shown in figure 6, average values fo

43、r u2 are needed. therefore, it makes sense to take the lower fitting line. thereby values between at a reduced difference length of 0.5 are achieved. the velocity ratio 2 can be described by a linear function: applying the model calculation, critical volume flow rates are achieved on the safe side b

44、etween 80 and 100% of the simulated values, that corresponds to a deviation of _10% from an average value. in figure 11 the volume flow rate, calculated by the derived model equations, for shell diameters dm d 600 mm and dm d 700 mm are shown. the results for the shell diameters dm d 400 mm and 1200

45、 mm are not shown here, but join together with the presented results. the calculated absolute values of u1 and vpcr: have been validated againstmeasurements in single-span tubes heat exchangers. as for the velocity distribution, only the relative values for are necessary to be known. for this purpos

46、e, a validation against cfd results is adequate.conclusionthe presented method produces equivalent 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 tha

47、n before. 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 influence of the different energy ratios in the partial sections 1. up to now

48、 only cases with a central position of the inlet nozzle has been evaluated; so, the investigation is not yet concluded. the following problems have to be clarified: _ influence of the transversal flow direction in a normal triangular array (30_) at higher pitch ratios 11, _ no symmetrical mode shape

49、 function and no central position of the inlet nozzle, _ model accommodations for other tube arrangements, _ development of a model, describing the partial flow rates in the baffle windows of real heat exchangers 管壳式换热器速度场及其振动情况分析为了使流体流动稳定，管壳式换热器的设计要基于稳定条件下在理想管道长度下的匀速流动。在实际生产的换热器中，流体流动不均匀会造成进口喷嘴，撞击管

50、板及支路产生位移偏差等情况。根据connors稳定性方程的延伸，得出等效的速度计算方法，需要了解到换热器找截流面的速度分布和振动情况。本次研究的目标是调查推导出其相关性及为这种计算方法提出意见和建议。 基于局部的恒速度，对应着相同的速度分布。 计算临界体积流量在实际设计中的应用，利用流体流动力学软件都可以计算实际管路中速度场的分布情况，从而确定流动的临界条件，进而出临界的体积流量。其对换热器的设计过程中进口段得分析及研究结果应用广泛，包括如下的几何参数，如：管壳直径，喷嘴直径、跨度、喷管出口与管壳之间的距离等。简介根据换热器的安全设计准则，为了避免由于流体不稳定流动造成的损失，整个管道长度的流

51、体流速分布都需要计算出来，特别是剖面与喷嘴挡板进口及出口处。但是到目前为止只有不成熟的假设，流体力学计算软件在管壳式换热器的流场计算中得到广泛的应用对几何参数对相关的计算流速和振型函数的影响进行了试验研究，及遵循相关设计标准。通过几何的变化，就能够通过管壳式换热器进口参数推导出简单的关联式和相应的流量等效速度分布。管壳式换热器壳程刚性管的流场的三维稳态计算公式是商用流体力学计算软件的starcd程序。了解程序中的k-w模型可以解决不可压缩湍流的三维n-s方程。入口的界面几何参数的研究换热器管束的不同区域的入口流量的影响因素不同，管子首先是由两个固定轴承支撑，所以支撑的管道长度为l 。我们所研究

52、的部分是合理的，因为速度分布在入口处是独立的流量，而这一条件在换热器其它部分的入口条件是通用的。在流场计算条件的稳态下，流体的体积和流量是固定的，为了确定轴向速度在管内的分部差距，流体在壳程中的正常条件下，对每只管壳的缝隙和邻近管进行速度分析，对每个缺口运用运处理的等效速度方程的均方根值得等效速度差距的情况是相反的。这三个平均等效速度的方法定义两个稳定的流向比率。第一阶30k定义为正常流动方向，是三个使用的方法中最大的等效速度和管壳的临界速度。它决定了第二阶60k与这两个平均等效速度的横向流动方向及其临界值。这是一个线性的估计情况，30k和60k取决于流体流动的角度。在这个程序基础上，通过图2

53、可以看出管壳的稳定性比前三排困在一起有所降低。对所有值k流速分布及振动激励在管壳式换热器前三排管束有所降低。实验研究数据表明，流体流动是否均匀，主要取决于捆绑化得程度，第二排的流速大约要比第一排高出50%左右。这是因为上游流体流动的速率要低于第一排，从而获得较低的流体力。尽管在第一排和第二排之间的管束内流体流动速率之间的差值相同。第二排的k值能量最高，这绝对了临界容积流量。最大的能量值k稳定在第一排管束与第二排管束之间。管壳式换热器的管束布局应该避免这种情况的发生。流体以正三角形的流动趋势向第二排管束靠近。在b=0.73的情况下，第一排管束的最大实际计算值要略微高于第二排管束的最大实际计算值，这是由于喷嘴的横向出口速度。此外，图三中所显示的横向流动方向并不明显。根据实际问题的研究表明，两个线性函数之间所表达的稳定性之比为2:5.关于稳定性之比2:5的数据是经过反复验证得到的经验值。研究报告在图四中给出了流体流动稳定性的临界体积流动速率。这种研究结果不同于之前的简单的设计方法，在均匀流场中的有效方法应该是截面积法。因为