VIEWS: 4 PAGES: 6 POSTED ON: 10/12/2012 Public Domain
HEAT TRANSFER AND HEAT EXCHANGERS INTRODUCTION The engineer is frequently called upon to transfer energy from one fluid to another. The transfer is most often effected by a heat exchanger. A steam power plant, for example, includes several major heat exchangers: boiler, superheater, economizer, and condenser. It also includes minor heat exchangers such as the gland exhaust condenser and the lube oil cooler. The principal heat exchangers in an air conditioning system are the evaporator and the condenser. Other applications abound throughout the mechanical industry. THEORETICAL CONSIDERATIONS AND THE HEAT TRANSFER PROBLEM The typical heat transfer problem involves energy transfer from a fluid stream at an assumed average temperature, through a series of “resistances”, to a destination fluid stream also at an assumed average temperature. For transfer through a pipe wall, the resistances consist of, in order, a thermal boundary layer (TBL), a scale or soot layer, the pipe wall, another scale layer, and the destination TBL. Each TBL is characterized by a convection (or film) coefficient designated h. Newton’s Law of Cooling models heat transfer through the TBL: q hT Ts where: q " = heat flux or rate of heat transfer per unit area Btu / h ft 2 h = convection (film) coefficient Btu / h ft 2 0 R T = average temperature of the fluid stream Ts = surface temperature of the pipe Fourier’s Law governs transfer by conduction through the pipe wall: dT q k dr where: k = thermal conductivity ( Btu / ft h 0 R ) The negative sign results from the decreasing temperature gradient. Integration yields: 1 2kLT1 T2 q ln r2 r1 where: q = heat transfer rate Btu / h L = length of pipe through which heat transfer occurs ft An experimentally determined fouling factor R''f ft 2 h 0 R / Btu accounts for scale resistance.1 THE OVERALL HEAT TRANSFER COEFFICIENT Heat transfer through the various layers can be represented by an overall heat transfer coefficient, U, such that: q UAD where: A = heat transfer surface area based on either the inside or outside pipe surface U = overall heat transfer coefficient (Btu/h-ft2-R) D = log mean temperature difference, derived below In analogy with electrical resistance, we can express the Newton and Fourier laws as: T1 T2 T1 T2 q conv q cond Rconv Rcond where: 1 ln r2 r1 Rconv Rcond hA 2kL We can then write: T1 T 2 T1 T 2 q Rtotal 1 ln r2 r1 Rf 2 R f1 1 A1 h1 A1 2kL A2 A2 h2 Selecting either the inside or outside area as a basis, factoring out 1/A1 (for example), and letting A1 = 2r1L, we have: 1 See, e.g., Incropera, p 585, for representative fouling factors. 2 q AUD 1 1 r ln r2 r1 r1 r U R1 1 f R2 1 f h1 k r2 r2 h2 LOG MEAN TEMPERATURE DIFFERENCE What temperature difference is to be used for D? Because the temperature varies on each side of the heat exchanger as heat transfer occurs, we must derive a sort of average temperature difference. Consider a counterflow heat exchanger, which we represent as shown in the figure below: Let: T = temperature of the hot fluid t = temperature of the cold fluid h = hot fluid c= cold fluid C= heat capacity rate mc i = entrance condition o = outlet condition D = log mean temperature difference Ch T dq T+dT t+dt t Cc Ti T1 dT dq To to dt T2 ti x q qc q h q h C h To Ti q c C c t o t i dq C h dT C c dt UTdA 3 dq dq 1 1 d T dT dt C UTdA h Cc C h Cc d T 1 1 2 2 T 1 U C dA h Cc 1 T T Ti t o t i UA ln 2 UA o T q q To ti Ti t o UA T2 T1 1 h qc q T2 T1 D T ln 2 T 1 For multipass and cross-flow heat exchangers, the expression for D is modified by a multiple F, which depends upon heat exchange geometry, number of tube passes and whether the fluid is mixed (flow transverse to the fluid direction is not prevented) or unmixed (transverse flow ix prevented by fins or separate tubes). The F-factor can be found in the technical literature.2 EXTENDED SURFACES (FINS) Extended surfaces, or fins, are frequently used to increase the heat transfer rate. Fins can be justified economically when fin effectiveness, defined by the below approximate equation, is greater than about 2. 1 kP 2 hA 2 f c where: k = thermal conductivity P = perimeter of fin h = convection coefficient Ac = cross sectional area of fin Because the temperature varies from the base of the fin to its tip, an adjustment is made to the total heat transfer surface area, which consists of the uncovered base area plus the fin area. 2 See, e.g., , Incropera and DeWitt, 4 ed, pp. 592-94. 4 Multiplying the total area, At, by an overall surface efficiency, o,3 which is less than unity, gives: o qt qt 1 NA f 1 f qmax hAt b At where: N = number of fins Af = area of a single fin At = total area including base f = fin efficiency qt = total heat transfer rate from At h = convection coefficient = Tb - T Tb = base temperature T = bulk temperature Fin efficiency is a function of fin geometry, convection coefficient and thermal conductivity. Values of fin efficiency for various shapes can be found in the technical literature. 4 Having found both fin and overall surface efficiency, the overall heat transfer coefficient is modified by multiplying each area (fin plus uncovered base) by the overall surface efficiency. This step results in the following relation for U: 1 1 R r ln r2 r1 r1 R2 r1 U f1 1 f o1h1 o1 k r2 o 2 o 2 r2 h2 From the definition of fin effectiveness, it is deduced that fins with a high thermal conductivity, low ratio of perimeter to cross-sectional area and small convection coefficient are desirable. Because a small convection coefficient improves fin performance, fins are frequently placed on the gas side of a gas-to-liquid heat exchanger. Automobile radiators and air conditioning evaporators and condensers are common examples. Fins are often placed on economizers in boiler designs. THE CONVECTION HEAT TRANSFER COEFFICIENT 3 See Incropera, Section 3.6.5. 4 See, e.g., Incropera, p 123 5 (FILM COEFFICIENT) Convection coefficients are determined from empirical correlations found in the literature. Incropera, for example, includes summaries of convection coefficients for external and internal flow in the back of chapters 7 and 8. The classical correlation for internal flow is the Colburn equation: NuD 0.023Re 4 / 5 Pr1/ 3 D hD where: NuD (the Nusselt number) k = thermal conductivity of fluid k VD Re (the Reynolds number) = kinematic viscosity k Pr (the Prandtl number) (thermal diffusivity) cp A cautionary note. The engineer must exercise care in choosing and using a correlation. Particular attention must be given to limits of applicability. Is the flow fully developed? Laminar or turbulent? What temperature should be used to find properties? What are the applicable ranges for the Reynolds and Prandtl numbers and the length to diameter ratio? DESIGN PROCEDURE Select an appropriate heat exchanger type. Select appropriate convection coefficients. Determine needed thermal properties. Solve the fundamental equations. Iteration may be required. qh mc p T qc mc p T q UAFD 6