外文翻译冷凝器和蒸发器的热力设计制定和应用.docx

1、外文翻译冷凝器和蒸发器的热力设计制定和应用附录B：参考英文文献与译文Thermodynamic design of condensers and evaporators:Formulation andapplicationsChristian J.L. Hermesa b s t r a c t This paper assesses the therm-hydraulic design approach introduced in a previous publication (Hermes, 2012) for condensers and evaporators aimed at min

2、imum entropy generation. An algebraic model which expresses the dimensionless rate of entropy generation as a function of the number of transfer units, the uid properties, the thermal-hydrauliccharacteristics, and the operating conditions is derived. Case studies are carried out with different heat

3、exchanger congurgitations for small-capacity refrigeration applications. The theoretical analysis led to the conclusion that a high effectiveness heat exchanger does not necessarily provide the best thermal-hydraulic design for condenser and evaporator coils, when the rates of entropy generation due

4、 to heat transfer and uid friction are of the sameorder of magnitude. The analysis also indicated that a high aspect ratio heat exchanger produces a lower amount of entropy than a low aspect ratio one.Conception thermodynamiccondenser et desse evaporate: formulation et applications.Keywords: floatin

5、g head;heat exchanger;design;industry1. IntroductionCondensers and evaporators are heat exchangers with fairly uniformwall temperature employed in a wide range of HVACR products, spanning from household to industrial applications. In general, they are designed aiming at acplishing a certain heat tra

6、nsfer duty at the penalty of pumping power.There are two well-established methods available for the thermal heat exchanger design, the log-mean temperature difference (LMTD) and the effectiveness/number of transfer units (-NTU) approach (Kakac and Liu, 2002; Shah andSiliculose, 2003). The second has

7、 been preferred to the former for the sake of pact heat exchanger design as the effectiveness (), dened as the ratio between the actual heattransfer rate and the maximum amount that can be transferred, provides a 1st-law criterion to rank the heat exchanger.performance, whereas the number of transfe

8、r units (NTU) pares the thermal size of the heat exchanger with its capacity of heating or cooling fluid. Furthermore, the -NTU approach avoids the cumbersome iterative solution required by the LMTD for outlet temperature calculations.Nonetheless, neither -NTU or LMTD approaches are suitable to addr

9、ess the heat transfer/pumping power trade-off,which is the crux for a balanced heat exchanger design. For this purpose, Bajan (1987) established the so-called thermodynamic design method, later renamed as entropy generation minimization method (Bajan, 1996), which balances the thermodynamic irrevers

10、ibilities due to the heat transfer with a nite temperature difference to those associated with the viscous uid ow, thus providing a 2nd-law criterion that has been widely used for the sake of heat exchanger design and optimization (San and Jan, 2000; Leprous et al., 2005; Achaean and Wongwises, 2008

11、; Mishap et al., 2009; Kotcioglu et al.,2010; Pussoli et al., 2012; Hermes et al., 2012). However, the models adopted in those studies do not provide a straightforward indication of howthe design parameters (geometry, uid properties, working conditions) affect the rate of entropygeneration. They als

12、o require plex numerical solutions,being therefore not suitable for back-of-the-envelope calculations in the industrial environment.In a recent publication, Hermes (2012) advanced an explicit, algebraic formulation which expresses the dimensionless rate of entropy generation as a function of the num

13、ber of transfer units, the uid properties, the thermal hydraulic characteristics ( j and f curves), and the operating.conditions (heat transfer duty, core velocity, and coil surface temperature) for heat exchangers with uniform wall temperature. An expression for the optimum heat exchanger effective

14、ness, based on the working conditions, heat exchanger geometry and uid properties, was also presented. The present paper is therefore aimed at assessing the formulation introduced by Hermes (2012) for designing condensers and evaporators for refrigeration systems spanning from house-hold application

15、, which amounts w10% of the electrical energy consumed worldwide (Malo and Silva, 2010).2. Mathematical formulationIn general, condensers and evaporators for refrigeration applications are designed considering the coil ooded with two-phase refrigerant, and also a wall temperature equal to the refrig

16、erant temperature (Barbarossa and Hermes, 2008), in such a way as the temperature proles along the streams are those represented in Fig. 1. In addition, the outer (e.g., air,water, brine) side heat transfer coefcient and the physical properties are assumed to be constant. Therefore, the heat transfe

17、r rate if calculated from:1where is the mass ow rate, Ti, To and Ts are the inlet, outlet and surface temperatures, respectively, Q hAs(TseTm) is the heat transfer rate, Tm is the mean ow temperature over the heat transfer area, As, and is the heat exchanger effectiveness, calculated from (Kays and

18、London, 1984):2where NTU hAs/mcp is the number of transfer units. The pressure drop, on the other hand, can be calculated from(Kays and London, 1984):3where f is the friction factor, uc is the velocity in the minimum ow passage, Ac, and the subscripts “i and “o refer to the heat exchanger inlet and

19、outlet ports, respectively. Oneshould note that Eqs. (1) and (3) can be linked to each other through the following approximation for the Gibbs relation,4where Tmz(Ti To)/2, and the entropy variation, soesi,is calculated from the 2nd-law of Thermodynamics,5where the rst term in the right-hand side ac

20、counts for the reversible entropy transport with heat ( _Q=Ts), whereas _Sg is the irreversible entropy generation due to both the heattransfer with nite temperature difference and the viscous ow. Substituting Eqs. (1), (3) and (5) into Eq. (4), it follows that: NS 6where NS is the dimensionless rat

21、e of entropy generation. The errors associated to the approximation used in Eq. (4) are marginal: noting that DTm 20 K in most small-capacityrefrigeration applications, it follows that the difference between the exact and approximated mean temperature never exceeds 1 K, which in turn affects the dim

22、ensionless entropy generation by less than 1%. Now noting that both condensers and evaporators are designed to provide a heat transfer duty subjected to ow rate and face area constraints Eq. (6) can be re-written as follows (Hermes, 2012):7And Q(ToeTi)/Ts a dimensionless temperature difference with

23、both To and Ti known from the application. One should note that the rst and second terms of the right-hand side of Eq. (7)stand for the dimensionless entropy generation rates associated with the heat transfer with nite temperature difference and the viscous ow, respectively. The optimum heat exchang

24、er design NTUopt that minimizes the rate of entropy generation is obtained from (Hermes, 2012):8drop effects, which rule the entropy generation for the low aspect ratio designs, are attenuated for low NTU values where the entropy generation due to nite temperature difference isDominant.4. Case studi

25、esFor the sake of heat exchanger design, Eq. (8) has to be solved concurrently with (ToeTi)/(TseTi) as the coil surface temperature, Ts,must be free to vary thus ensuring that Q (andso _Q and _ m) is constrained. However, the solution is implicit.for Ts, thus requiring an iterative calculation proce

26、dure:a guessed Ts value is needed to calculate the effectiveness and NTU eln(1e), which is used in Eq. (8) with j j(Re) and f f(Re) curves, and also with the dimensionless core velocity to e outwith Q,which in turn is used to recalculate Ts until convergence is achieved. Firstly consider an air-supp

27、lied tube-n condenser for small-capacity refrigeration appliances running under the following working conditions: _ Q 1kW, _ V 1000 m3h1Ti 300 K (Waltrich et al., 2011; Hermes et al., 2012). Lets assume two heat exchanger congurations: (i) circular tubes with at ns (i.e., Kays and Londons surface 8.

28、0-3/8T), whose thermal-hydraulic characteristics are j 0.16$Re0.4,tubes and ns (Kays and Londons surface CF-8.72), whose thermal-hydraulic characteristics are j 0.22$Re0.4 f 0.20$Re0.2, s 0.524 and Dh 3.93 mm. Also note that Pr z 0.7 for air. Fig. 4 pares the performance characteristics ( j and f cu

29、rves) of surfaces 8.0-3/8T and CF-8.72 as functions of Re rucDh/m. Fig. 5 pares the dimensionless entropy generation Observed for both surfaces as a function of NTU. A curve of (NTU), which the same for both surfaces, is also plotted to be used as a reference. It can be clearly seen that the (, NTU)

30、 design which minimizes the rate of entropy generation is (0.61, 0.95) for surface 8.0-3/8T and (0.57, 0.81) for surface CF-8.72. It can also be noted that the circular-n surfaceshowed a higher rate of entropy generation for all NTU span,which is mostly due to the viscous uid ow effect assurface CF-

31、8.72 has a higher friction factor than surface 8.0-3/8T for the same Reynolds number (see Fig. 4). For low NTU values, where the entropy generation is ruled by NS,DT, both surfaces showed similar NS values as their j-curves are close (see Fig. 4).Fig. 6 pares three different condenser designs consid

32、- ering surface 8.0-3/8T and face areas varying from 0.025 to 0.1 m2 running under the same working conditions. The heat exchanger length was also varied in order to acmodate the heat transfer surface area for different face areas. For a vertical, constant NTU line (i.e. same heat transfer area), it can be clearly observed that a heat exchanger design with high aspect ratio (higher face area, smaller length in the ow direction) produces a signicantly lower amount of entropy in parison t