外文翻譯--脈沖管式冷卻器中制冷器的開發(fā)和設(shè)計（英文）

1、AI AA-2000-2868 DEVELOPMENT AND DESIGN OF COLD HEAT EXCHANGER OF PULSE TUBE COOLER Wei Dong, Marco Lucentini, Vincenzo Naso Dr. Wei Dong - dong@uniromal .it Dr. Marco Lucentini - luce@uniromal .it Prof. Vincenzo Na

2、so Ph.D. - naso@uniromal .it University of Rome “La Sapienza” Dept. of Mechanical and Aeronautical Engineering Via Eudossiana. 18 - 00 184 - Rome - ITALY Fax: +39-06-488 1759 ABSTRACT During the past fifteen ye

3、ars heat transfer in oscillating flows become the subject of increasing interest in the engineering community. Applications of oscillating flows include, for example, the cooling electric equipment or alternative, en

4、vironmentally safe refrigeration technologies, such as pulse tubes or Stirling refrigerators. Important components of these refrigerators are their heat exchangers. In such devices the working fluid is subject to osc

5、illatory forcing which is a key part of the process, as opposed to situations where oscillations are generated with the aim to enhance heat transfer. Heat transfer in oscillating, and often compressible flow, has not

6、 yet been completely understood, and the lack of design methodologies for heat exchangers in such flow is one reason that efficiencies of these devices are limited. A practical design method of cold heat exchanger wa

7、s developed in this paper. The influence of the workmg gas oscillating on heat transfer was considered in this method. The design parameters related to gas oscillation can be calculated using known methods. NOTATION

8、 Af cp specific heat D plate to plate spacing d,, hydraulic diameter H a space of cooler h heat transfer coefficient L k thermal conductivity m mass flow rate n number of plate P pressure Pr Prandtl

9、number q heat transfer rate R e ,amplitude of the instantaneous Reynolds number Re, dimensionless frequency cross - sectional area for flow length of cold heat exchanger Tel: +39-06-44585271/258 T temperature U

10、 , ,maximum velocity X ,a coefficient of thermal diffusivity c1 dynamic viscosity V kinetic viscosity P density Z ,0 angular frequency tidal displacement of gas parcel - averaged shear stress of the gas INTRODUCT

11、ION Heat exchangers are devices which enhance the transfer of the heat and are vital components of every cryocooler. They exist in a wide variety of types, shapes, sizes and arrangements and are made of all kinds of

12、 materials. In order to exploit the thermoacoustic effect for heat pumping in pulse tube refrigerator (PTR), heat exchangers are attached both ends of the pulse tube. The cold heat exchanger removes heat fiom a co

13、ld temperature reservoir and the hot exchanger rejects the pumped heat and absorbed acoustic work to the environment at temperature. Here we analyze only the cold heat exchanger of PTR. In general, the design methods

14、 of the heat exchanger are for steady flow and the heat is exchanged between two gas streams owing unidirectionally with steady volume flow rates. These methods are not suitable for oscillating flow. For this r

15、eason, the available design method was developed with taking into account the heat transfer correlation for oscillating flow conditions in heat exchanger. The proposed design method allows independent optimizatio

16、n of heat exchanger of PTR. 1. HEAT TRANSFER PROCESS OF WORKING GAS IN COLD HEAT EXCHANGER A closer look into the cold heat exchanger, illustrated in the magnified region of Figure 1, reveals the mechanism responsib

17、le for thermoacoustic heat 420 (6) In conclusion, in the limit D+O, the total cooling rate decreases as D2. This trend is illustrated qualitatively as curve (a) in Figure 3. b) The boundary layerflow l i m i t .In the

18、 opposite extreme, D-m, the boundary layer that lines on one surface becomes “distinct”. In other words, each channel looks like the entrance region to parallel plate duct. The overall pressure drop is fixed at AP. T

19、he overall force balance on the control volume H x L requires AP-H =n.2.ZW .L (7) in which n is the number of the channels and Z , is the L averaged shear stress of gas 1 12 2, = 0.1 328Re+ - pU3 ( 8 ) Combin

20、ed, equation (7) and (8) yield (9) The total heat transfer rate from one of the L long surfaces ( 4 ; ) can be calculated by recognizing the overall Nusselt number for Pr> 0.5; (10) which leads to q; = q”L = k

21、(T, - T-)0.664PF E ( V 1 x (1 1) The total heat transfer rate released by the entire stack is 2n times larger than x q; = 2nq; = 2nk(T, - T- )0.644Pf: (T] (12) In view of the n and U expression listed in eq, the

22、 total heat transfer rate becomes (13) The second conclusion we reached is that in the large D limit, the total heat transfer rate decreases as D ’. This second trend has been added as curve (b) to the same graph (F

23、igure 3.) DISCUSSION AND CONCLUSION According to above limit conditions, maximum of actual (unknown) curve q‘(D) can only occur at an optimal spacing Do,,, that is of the order as the D value obtained by intersecting the

24、 asymptotes qa and q . It is easy to show that the order of magnitude statement. q, - qb yields the following spacing: L (14) This estimate agrees very well with the more exact result obtained by the maximum of act

25、ual q’(D) sketched in figure 3. The order of magnitude of the maximum package heat transfer rate that corresponds to D = D,, is obtained by combining eqs (6) and (14): (15) The brief scale analysis represented in th

26、is section can be repeated for the situation in which only one surface of the board is Joule-heated to T,, and the other surface can be modeled as adiabatic. The only change is that 2n is replaced by n in eq. (12), s

27、o that results become -- z2.10 - L ( APL2 pa r q m a x’ - < 037( e)% Pr Hc, (T, - T-) (17) It is obviously that the change in the thermal boundary conditions of one plate to plate channel affects only the

28、 value of numerical coefficient in the expressions for Dopt and q max . A practical design approach of cold heat exchanger was discussed in above. The influence of the working gas oscillating on heat transfer was co