The optimal dimensions of convective-radiating circular fins

The optimal dimensions of convective-radiating circular fins with variable profile, heat-transfer coefficient and thermal conductivity, as well as internal heat generation are obtained. A profile of the form y=(w/2) [1+(r _o/r) ⁿ ] is studied, while variation of thermal conductivity is of the form k=k _o[1+ɛ((T−T _∞)/ (T _b−T _∞)) ^m ]. The heat-transfer coefficient is assumed to vary according to a power law with distance from the bore, expressed as h=K[(r−r _o)/(r _e−r _o)]^λ. The results for λ=0 to λ=1.9, and −0.4≤ɛ≤0.4, have been expressed by suitable dimensionless parameters. A correlation for the optimal dimensions of a constant and variable profile fins is presented in terms of reduced heat-transfer rate. It is found that a (quadratic) hyperbolic circular fin with n=2 gives an optimum performance. The effect of radiation on the fin performance is found to be considerable for fins operating at higher base temperatures, whereas the effect of variable thermal conductivity on the optimal dimensions is negligible for the variable profile fin. It is also observed, in general, that the optimal fin length and the optimal fin base thickness are greater when compared to constant fin thickness. Received on 22 February 1999     

Boiling on a straight pin fin with variable thermal conductivity

This work theoretically investigated the thermal performance and stability characteristics of a straight pin fin subject to boiling considering a temperature-dependent thermal conductivity of fin, k=k _sat(1+b(T−T _sat)). Steady-state temperature distribution and the associated fin base heat flow were for the first time analytically found, whose stability characteristics were evaluated by linear stability analysis. A positive temperature coefficient b will raise both the fin's temperature and base heat flow. The corresponding stability for stable fin boiling was enhanced. A negative b results in an opposite trend. The use of a mean thermal conductivity in fin boiling calculations is discussed. Received on 3 November 1997     

Computation of conjugate heat transfer in the turbulent mixed convection regime in a vertical channel with multiple heat sources

This paper presents the results of a comprehensive numerical study to analyze conjugate, turbulent mixed convection heat transfer from a vertical channel with four heat sources, uniformly flush-mounted to one of the channel walls. The results are presented to study the effect of various parameters like thermal conductivity of wall material (k _s), thermal conductivity of flush-mounted discrete heat source (k _c), Reynolds number of fluid flow (Re _s), modified Richardson number (Ri ⁺) and aspect ratio (AR) of the channel. The standard k-ε turbulence model, modified by including buoyancy effects with physical boundary conditions, i.e. without wall functions, has been used for the analysis. Semi-staggered, non-uniform grids are used to discretise the two dimensional governing equations, using finite volume method. A correlation, encompassing a wide range of parameters, is developed for the non-dimensional maximum temperature (T ^*) using the asymptotic computational fluid dynamics (ACFD) technique.     

Conjugate natural convection in a square porous cavity

 Steady-state conjugate natural convection in a square cavity filled with a porous medium is studied numerically in this paper. The enclosure consists of two horizontal conductive walls of finite thickness and two vertical walls at different uniform temperatures. The focus is on the role of solid-fluid conductivity ratio, k, on the flow and heat transfer characteristics and the average Nusselt number, , over the vertical hot and cold walls of the cavity for a limited set of particular parameters. It was shown that the interface temperature, θ_w, along the top of the solid wall decreases with the increase in the wall conductivity k. Also, the values of decreases with the increase of the values of the parameter k. Comparison with known results from the open literature when the wall thickness of the horizontal solid walls is neglected (non-conjugate problem) is excellent. Received on 4 April 2000     

Natural convection in shallow enclosures with multiple conducting partitions

Laminar natural convection and conduction in shallow enclosures having multiple partitions with finite thickness and conductivity have been studied. An approximate analytical solution is obtained by using the parallel flow approximation in horizontal shallow enclosures heated isothermally at two vertical ends while adiabatic on horizontal end walls. The same problem is solved also using a finite difference formulation and the control volume method. The study covers the range ofRa from 105 to 107,A=H/L0.2, C=1/L from 0 to 0.15, and the thermal conductivity ratio of partition to fluidk _r from 10^–4 to 10¹¹. The partition numberN was varied from 0 to 5. The Prandtl number was 0.72 (for air). The results are reduced in terms ofNu as a function ofRa, k, and various geometrical parameters (A, C). The streamlines and isotherms are produced to visualize the flow and temperature fields.Es wird der kombinierte Einfluß von laminarer Naturkonvektion und Leitung in flachen Behältern mit mehreren Trennwänden endlicher Dicke und Leitfähigkeit untersucht. Eine analytische Näherungslösung läßt sich über die Parallelstromapproximation bezüglich horizontaler flacher Behälter finden, deren zwei vertikale Begrenzungswände isotherm beheizt sind, während die Horizontalflächen adiabat sein sollen. Das selbe Problem wird unter Verwendung eines Differenzverfahrens und der Kontrollvolumen-Methode gelöst und zwar für die Parameterbereiche 105 Ra 10⁷;A=H/L<0.2;>C=1/L 0.15; 10^–4k_r 1011, wobei der letzte Parameter das Verhältnis der Leitfähigkeit von Trennwand und Fluid bezeichnet. Die Zahl der TrennwändeN variierte Zwischen 0 und 5, die Prandtl-Zahl betrug 0.72 (Luft). Die Ergebnisse werden in dimensionsloser Form gemäß der BeziehungNu =f (Ra, k _r ,A, C) mitgeteilt bzw. durch Diagrammdarstellungen der Stromlinien- und Isothermenfelder veranschaulicht.Financial support from the Natural Sciences and Engineering Council Canada is acknowledged. Financial support to A. Kangni from Canadian Fellowship Program For French Speaking Countries is also acknowledged.     

Treatment of moisture condensation on fins

An analysis of natural convection from a vertical plate fin when the fin base temperature is below the dew point of the surrounding air is presented in this paper. The analytical solution derived is based upon a constant heat and mass transfer coefficient and is also valid for forced convection. The results of this simplified theory are compared with a numerical solution where the coupling of convection and conduction is taken into account. An experimental verification of the results is also shown.

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Aus Kondensation von Feuchtigkeit an Rippen

Zusammenfassung Es wird eine Analyse der freien Konvektion an einer vertikalen plattenförmigen Rippe dargestellt, bei der die Temperatur im Anfangsbereich der Rippe unterhalb des Taupunktes der umgebenden Luft liegt. Die abgeleitete analytische Lösung beruht auf einem konstanten Wärme- und Stoffübergangskoeffizienten und gilt auch für die erzwungene Konvektion. Die Resultate dieser vereinfachten Theorie werden mit einer numerischen Lösung verglichen, in der die Verbindung von Konvektion und Wärmeleitung in Betracht gezogen wird. Angeführt wird auch eine experimentelle Bestätigung der Resultate.

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Nomenclature a _f thermal diffusivity of air - A, B constants in Eq. (7) - c constant defined in Eq. (3) - D diffusion coefficient - f an arbitrary function ofT andx in Eq. (12) - F ₁,F ₂ coefficients in differential Eq. (13) - g gravitational acceleration - h heat transfer coefficient - h _m mass transfer coefficient - k thermal conductivity of fin - k _f thermal conductivity of air - l latent heat of moisture condensation - L total length of fin - L _w length of wet fin - m parameter, (h/kt)^1/2 - m _l dimensionless parameter, 1+ B/T _r - m _y parameter,m m _l ^1/2 - p pressure of surrounding air - p _ws saturation pressure of water vapor - p _w partial pressure of water vapor in air - Pr Prandtl number,/a _f - q total heat fluxl - q _c convective heat flux - q _m heat flux - q _r radiative heat flux - R parameter in Eq. (14) - R _w specific gas constant of water vapor - t half thickness of fin - T temperature - T _b base temperature of wet fin - T _c base temperature of dry fin=saturation temp. of vapor - T _r reference temperature defined in Eq. (15) - T temperature of surrounding air - T temp, difference between fin surface and surroundings - v initial temperature for quasilinearization - x vertical coordinate, see Fig. 1 - y horizontal coordinate, see Fig. 1 - coefficient of thermal expansion - emissivity - dimensionless parameter in Eq. (14) - ø _d heat flux of dry fin - ø _tot total heat flux of dry-wet fin - kinematic viscosity - Stefan-Boltzman coefficient - relative humidity of air     

Heat transfer for laminar flow in internally finned pipes with different fin heights and uniform wall temperature

An analysis is presented for fully developed laminar convective heat transfer in a pipe provided with internal longitudinal fins, and with uniform outside wall temperature. The fins are arranged in two groups of different heights. The governing equations have been solved numerically to obtain the velocity and temperature distributions. The results obtained for different pipe-fins geometries show that the fin heights affect greatly flow and heat transfer characteristics. Reducing the height of one fin group decreases the friction coefficient significantly. At the same time Nusselt number decreases inappreciably so that such reduction is justified. Thus, the use of different fin heights in internally finned pipes enables the enhancement of heat transfer at reasonably low friction coefficient.Nomenclature A_f dimensionless flow area of the finned pipe, Eq. (8) - a_f flow area of the finned pipe - C_p specific heat at constant pressure - f coefficient of friction, Eq. (12) - H₁, H₂ dimensionless fin height h₁/r_o h₂/r_o - h₁, h₂ fin heights - average heat transfer coefficient at solid-fluid interface - KR fin conductance parameter, k_s/k_f - k_f thermal conductivity of fluid - k_s thermal conductivity of fin - l pipe length - mass flow rate - N number of fins - Nu Nusselt number, Eqs. (15) and (16) - P pressure - Q total heat transfer rate at solid fluid interface - Q_f1, Q_f2 heat transfer rate at fin surface - q_w average heat flux at pipe-wall, Q/(2 r_ol) - R dimensionless radial coordinate r/r_o - Re Reynolds Number, Eq. (13) - r radial coordinate - r_o radius of pipe - r₁, r₂ radii of fin tips - T temperature - T_b bulk temperature - U dimensionless velocity, Eq. (2) - U_b dimensionless bulk velocity - u_z axial velocity - z axial coordinate - angle between the flanks of two adjacent fins - half the angle subtended by a fin - angle between the center-lines of two adjacent fins - angular coordinate - dynamic viscosity - density - dimensionless temperature, Eq. (6) - _b dimensionless bulk temperature     

The Use of Porous Fins for Heat Transfer Augmentation in Parallel-Plate Channels

In this study, a steady, fully developed laminar forced convection heat augmentation via porous fins in isothermal parallel-plate duct is numerically investigated. High-thermal conductivity porous fins are attached to the inner walls of two parallel-plate channels to enhance the heat transfer characteristics of the flow under consideration. The Darcy–Brinkman–Forchheimer model is used to model the flow inside the porous fins. This study reports the effect of several operating parameters on the flow hydrodynamics and thermal characteristics. This study demonstrates, mainly, the effects of porous fin thickness, Darcy number, thermal conductivity ratio, Reynolds number, and microscopic inertial coefficient on the thermal performance of the present flow. It is found that the highest Nusselt number is achieved at fully filled porous duct which requires the highest pumping pressure. The results show that using porous fins requires less pumping pressure with comparable high heat augmentation weight against fully filled porous duct. It is found that higher Nusselt numbers are achieved by increasing the microscopic inertial coefficient (A), the Reynolds number (Re), and the thermal conductivity of the porous substrate k ₂. The results show that heat transfer can be enhanced (1) with the use of high thermal conductivity fins, (2) by decreasing the Darcy number, and (3) by increasing microscopic inertial coefficient.     

Cooling of an electronic board situated in various configurations inside an enclosure: lattice Boltzmann method

Natural convection heat transfer in a square cavity induced by heated electronic board (as a thin plate at constant temperature) is investigated using the lattice Boltzmann method. Lattice Boltzmann simulation of natural convective heat transfer in a cavity in the presence of internal straight obstacle has not been considered completely in the literature and this challenge is generally considered to be an open research topic that may require more study. The present work is an extension to our previous paper (see Nazari and Ramzani in Modares. Mech. Eng. 11(2):119–133, 2011) in which the effects of position and dimensions of obstacle on the flow pattern and heat transfer rate are completely studied. A suitable forcing term is represented in the Boltzmann equation. With the representation, the Navier–Stokes equation can be derived from the lattice Boltzmann equation through the Chapman-Enskog expansion. Top and bottom of the cavity are adiabatic; the two vertical walls of the cavity have constant temperatures lower than the plate’s temperature. The study is performed for different values of Grashof number ranging from 10³ to 10⁵ for different aspect ratios and position of heated plate. The effect of the position and aspect ratio of heated plate on heat transfer are discussed and the position of the obstacle in which the maximum rate of heat transfer is investigated in both vertical and horizontal situation. The obtained results of the lattice Boltzmann method are validated with those presented in the literature.     

Natural convection in a square enclosure with an internal isolated vertical plate

Numerical computations were performed for the average Nusselt number at an internal vertical plate situated in a square enclosure, with the inner plate and the bounding wall of the enclosure maintained at uniform but different temperatures. Natural convection occurred in the air which occupied the enclosure space. The position of the inner vertical plate within the enclosure was varied parametrically. The plate height-cavity height ratio was 0.513. For narrow distance between the inner plate and the bounding wall the inner plate Nusselt number was enhanced. Aside from this, the plate average Nusselt number was remarkably insensitive to the plate position. The effect of the Rayleigh number on the velocity and temperature fields and local Nusselt numbers are also discussed. The agreement between the predicted flow pattern forRa=1.1×10⁶ and the flow visualization result was reasonably good.

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Natürliche Konvektion in einem quadratischen Horizontalschacht, der eine freistehende, senkrechte Platte enthält

Zusammenfassung Eine numerische Untersuchung liefert mittlere Nußelt-Zahlen an einer, in einem quadratischen Horizontalschacht freistehenden, senkrechten Platte, wobei deren Temperatur und die der umgebenden Wände jeweils konstant gehalten werden. Im Luftraum dazwischen stellte sich freie Konvektion ein. Die Position der Platte war veränderlich, ihre Höhe blieb mit 51.3% der Schachthöhe konstant. Rückte die Platte nahe an eine Schachtwand, so erhöhte sich die Nußelt-Zahl auf der dieser zugewandten Seite, während die Gesamt-Nußelt-Zahl bezüglich der Platte fast konstant bleibt. Es wird auch der Einfluß der Rayleigh-Zahl auf das Geschwindigkeitsund Temperaturfeld diskutiert. BeiRa=1.1·10⁶ stimmten die Ergebnisse aus der Berechnung gut mit den experimentellen Befunden einer Strömungsvisualisation überein.

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Nomenclature a distance between vertical plate and side-wall of enclosure thermal diffusivity (in definition ofu _r) - b distance between vertical plate and bottom of enclosure - g gravitational acceleration - G characteristic flow rate - H height of vertical plate - k thermal conductivity - k _f fluid thermal conductivity - K relative thermal conductivity,k/k _f - L width of square enclosure - M _res mass residual - Nu local Nusselt number - Nu _m average Nusselt number - Nu _L local Nusselt number of left side of vertical plate - Nu _R local Nusselt number of right side of vertical plate - Nu _B local Nusselt number of bottom side of vertical plate - Nu _T local Nusselt number of top side of vertical plate - p effective pressure - P dimensionless pressure,P=p/[(Ra Pr)(a/H)²] - Pr Prandtl number - Ra Rayleigh number,Ra=gTH ³ Pr/ ² - T temperature - T _i temperature of internal plate - T _o temperature of enclosure surface - u, v velocity components inx-, y-direction - U, V dimensionless velocities,U=u/u _r, V=v/u_r - u _r reference velocity,u _r=(Ra Pr)^1/2/(a/H) - X, {iyY} dimensionless coordinates,X=x/H, Y=y/H Greek symbols heat transfer coefficient - volume expansion coefficient - thickness of plate - kinematic viscosity - density - dimensionless temperature, (T _i–T)/(T _i–T_o)     

Hypersonic tail-fin heat transfer

The flow and heat transfer on the windward surface of tail fins has been experimentally investigated for Mach numbersM =5 and 8 and Re_L=(0.6–1.1)·10⁶ (L is the length of the central chord of the wing on which the fins are mounted). Two lines of flow divergence and, consequently, two zones of enhanced heat transfer on the surface of the fin have been detected. The angle of inclination of the fin to the wing surface, the angle of attack of the wing and the radius of the wing-fin junction were varied.Translated from Izvestiya Rossiiskoi Akademii Nauk, Mekhanika Zhidkosti i Gaza, No.2, pp. 18–25, March–April, 1993.The authors wish to thank S. D. Fonov and T. A. Ershova for the digital analysis of the photographs obtained by the thermal indicator coating and laser knife-edge methods.     

Forced convection heat transfer at the boundary layer of a packed bed

Numerical investigations of the nature of the fluid flow pattern and heat transfer at the boundary layer of a packed bed are reported. A volume averaged Navier-Stokes equation is used to predict the fluid flow and a volume averaged heat balance equation the heat transfer. A variable porosity in the packing is assumed in the region near the wall. Simulations are performed using a modified penalty Galerkin finite element method. The case of fully developed hydrodynamic flow and developing thermal flow is studied. The Nusselt number is found to depend on the Reynolds number, Graetz number and ratio of thermal conductivity of the solid and fluid phases. Comparison is made to some experimental literature values.Nomenclature A constant - [A] Navier-Stokes type matrix - B constant - [B] divergence matrix - C _p constant pressure heat capacity - d characteristic length - D _p particle diameter - D _t tube diameter - {F} solicitation vector - Gz Graetz number, z D _t ^–1 Pr _f Re _p - k permeability term - k _f Thermal conductivity of the fluid phase - k _s Thermal conductivity of the solid phase - [K] coefficient matrix for temperature equation - n normal vector - P pressure - Pr _f Prandtl number for the fluid _f C _p k _f ^-1 - r radial coordinate - R _t tube radius - R residual - R _m residual - Re _p Reynolds number for particle, - t tortuosity factor - T temperature - interstitial velocity - z axial coordinate - effective thermal conductivity - penalty parameter - boundary of solution domain - porosity - viscosity - density - test function - solution domain - test function     

Convection and heat transfer of elliptical tubes

Convection heat transfer (including natural and forced convection) of elliptical tubes had been studied system-atically. The experienced formula of heat transfer had been given. It presents fin efficiency of rectangular finned elliptical tube and optimized fin geometry (i.e. length/width ratio) and fin spacing for rectangular fin.

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Konvektion und Wärmeübergang an elliptischen Rohren

Zusammenfassung Konvektion und Wärmeübergang (sowohl bei freier als auch bei erzwungener Konvektion) an elliptischen Rohren wurden systematisch untersucht. Es wird eine aus dem Experiment abgeleitete Beziehung für den Wärmeübergang angegeben, die den Gütegrad elliptischer Rohre mit Rechteckrippen unter optimierter Rippengeometrie (Längen- zu Breitenverhältnis und Rippenabstand) beinhaltet. Stichworte: Konvektiver Wärmeübergang, Rippengütegrad, Elliptische Rohre mit Rechteckrippe.

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Nomenclature A length of a rectangular fin - A _F area of a fin - A _f overall fin area of per length - A _r tube surface of per length - B width of a rectangular fin - Gr Grashof number - N fin number of per length - Nu =hl/ Nusselt number - Pr Prandtl number - Ra =Gr·Pr Rayleigh number - Re =wl/ Reynolds number - S ₁ transverse tube pitch - S ₂ longitudinal tube pitch - w fluid velocity - a long axis of a ellipse - b short axis of a ellipse - c =a/b shape factor - d _e equivalent diameter - g acceleration of gravity - h heat exchange coefficient - l characteristic length - t fin spacing Greek symbols coefficient of thermal expansion - fin thickness - _fm area average fin efficiency - coefficient of fluid thermal conductivity - _f fin's thermal conductivity - kinematic viscosity     

A spectral‐element discontinuous Galerkin thermal lattice Boltzmann method for conjugate heat transfer applications

We present a spectral‐element discontinuous Galerkin thermal lattice Boltzmann method for fluid–solid conjugate heat transfer applications. Using the discrete Boltzmann equation, we propose a numerical scheme for conjugate heat transfer applications on unstructured, non‐uniform grids. We employ a double‐distribution thermal lattice Boltzmann model to resolve flows with variable Prandtl (Pr) number. Based upon its finite element heritage, the spectral‐element discontinuous Galerkin discretization provides an effective means to model and investigate thermal transport in applications with complex geometries. Our solutions are represented by the tensor product basis of the one‐dimensional Legendre–Lagrange interpolation polynomials. A high‐order discretization is employed on body‐conforming hexahedral elements with Gauss–Lobatto–Legendre quadrature nodes. Thermal and hydrodynamic bounce‐back boundary conditions are imposed via the numerical flux formulation that arises because of the discontinuous Galerkin approach. As a result, our scheme does not require tedious extrapolation at the boundaries, which may cause loss of mass conservation. We compare solutions of the proposed scheme with an analytical solution for a solid–solid conjugate heat transfer problem in a 2D annulus and illustrate the capture of temperature continuities across interfaces for conductivity ratio γ > 1. We also investigate the effect of Reynolds (Re) and Grashof (Gr) number on the conjugate heat transfer between a heat‐generating solid and a surrounding fluid. Steady‐state results are presented for Re = 5?40 and Gr = 10⁵?10⁶. In each case, we discuss the effect of Re and Gr on the heat flux (i.e. Nusselt number Nu) at the fluid–solid interface. Our results are validated against previous studies that employ finite‐difference and continuous spectral‐element methods to solve the Navier–Stokes equations. Copyright © 2016 John Wiley & Sons, Ltd.     

Thermodynamic analysis of the one-dimensional heat conduction with inhomogeneities in the heat flux direction

Starting from the results of Li, Prigogine and others about the one-dimensional heat conduction with constant temperature boundary conditions, the aim of this paper is to study, according to the methods and the purposes of the generalized thermodynamics, the more general case of one-dimensional heat conduction in systems, whose conductivity is function of both temperature and the coordinate in the heat flux direction and presents a finite number of discontinuities.

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Thermodynamische Analyse für eindimensionale Wärmeströmung mit Ungleichartigkeiten in der Wärmeflußleitung

Zusammenfassung Ausgehend von den Ergebnissen von Li, Prigogine und anderen, für eindimensionale Wärmeströmung mit konstanten Temperaturen an den Grenzen, versucht diese Arbeit, gemäß den Methoden und den Zielen der verallgemeinerten Thermodynamik, den allgemeineren Fall der eindimensionalen Wärmeströmung mit Ungleichartigkeiten zu examinieren.

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Nomenclature c volumetric specific heat - G discriminating parameter [G (t) =k ₂ (t)t ² orG _i (t)=k_i(t) t², when the separation of the variables for thermal conductivity can be done, or in general: G(x, t)=k(x,t)t ²] - J heat transfer rate (generalized flux) [W] - J ₀ heat flux [W/m²] - k thermal conductivity - k ₁ component of thermal conductivity depending upon the coordinate in the heat flux direction - k ₂ component of thermal conductivity depending upon temperature - k _i component of the thermal conductivity of a homogeneous layer (i) depending upon temperature - L ₁,L ₂ extreme coordinates - Lip Lipschitz's function - P entropy production rate - (P(T))_min temperature distribution in a system corresponding to the minimum of entropy production rate - p ₁ thermokinetic potential - (P ₁ (T))_min temperature distribution in a system corresponding to the minimum of thermokinetic potential - P ₂ generalized force potential - Q differential form - dQ total differential form - ¯R set of all real numbers - S(x) area of the isothermal surface corresponding to the coordinatex - T system temperature distribution - t absolute temperature - ¯T set of all the possible system temperature distributions - x symbol of Cartesian product - x coordinate in the heat flux direction - X local generalized force - B,Y ₀,Y ₀ ^*,Z sets (defined in this paper) - function representing the time evolution of the temperature distribution in a system - time - ₀ reference time interval     

Measurements of heat transfer coefficients between turbulent liquid jets and a radially finned heated disk placed at the bottom of an open vertical circular cavity

 Experiments have been performed to assess the impact of an extended surface on the heat transfer enhancement for axisymmetric, turbulent liquid jet impingement on a heated round disk. The disk, with an array of integral radial fins mounted on its surface, is placed at the bottom of an open vertical circular cavity. Hydrodynamic and heat transfer data were obtained for a dielectric fluorocarbon liquid FC-77. For a fixed circular heater of diameter D=22.23 mm, several geometric parameters were tested: the nozzle diameter (4.42≤d≤9.27 mm), the confining wall diameter of the vertical cavity (22.23≤D _c≤30.16 mm), and the nozzle-to-heater spacing (0.5≤S/d≤5.0). The FC-77 flow rates varied from =0.2 to 11.0 l/min producing Reynolds numbers in the wide interval 700≤Re _d ≤44,000. For d=4.42 mm, the heat transfer response to the separation distance S/d was small but increased gradually with increasing nozzle diameter up to d=9.27 mm. The thermal resistance R _th increased with the confining wall diameter D _c and also with the nozzle diameter d. A minimum value of the thermal resistance of R _th,min=0.4 cm² K/W was attained for a combination of d=4.42 mm, D _c=22.23 mm, S/d=1, and =7.5 l/min. Based on a simplified heat transfer model, reasonable agreement was obtained between measured values of the thermal resistance and the R _th-predictions. The total fin effectiveness ɛ_f was shown to increase with increasing nozzle diameter, but was invariant with the flow rate (or the jet exit velocity). More than a three-fold heat transfer enhancement was realized through the addition of the array of integral radial fins on the heated round disk. Received on 30 August 2000 / Published online: 29 November 2001     

Control of a submerged jet in a thin rectangular cavity

An innovative method is presented for control of an oscillatory turbulent jet in a thin rectangular cavity with a thickness to width ratio of 0.16. Jet flow control is achieved by mass injection of a secondary jet into the region above the submerged primary jet nozzle exit and perpendicular to the primary nozzle axis. An experimental model, a 2-D and a 3-D computational fluid dynamics (CFD) model are used to investigate the flow characteristics under various secondary injection mass flow rates and injection positions. Two-dimensional laser Doppler anemometry (LDA) measurements are compared with results from the CFD models, which incorporate a standard k–ε turbulence model or a 2-D and 3-D realisable k–ε model. Experimental results show deflection angles up to 23.3° for 24.6% of relative secondary mass flow are possible. The key to high jet control sensitivity is found to be lateral jet momentum with the optimum injection position at 12% of cavity width (31.6% of the primary nozzle length) above the primary nozzle exit. CFD results also show that a standard k–ε turbulence closure with nonequilibrium wall functions provides the best predictions of the flow.     

Numerical computation of high Rayleigh number natural convection and prediction of hot radiator induced room air motion

The two-dimensional stationary turbulent buoyant flow and heat transfer in a cavity at high Rayleigh numbers was computed numerically. The k– turbulence model was used. The time-averaged equations for momentum, energy and continuity, which are coupled to the turbulence equations, were solved using a finite difference formulation. In order to validate the computer code, a comparison exercise was carried out. The test results are in good agreement with the internationally accepted benchmark solution. Grid-refinement shows the necessity of a very fine grid at high Rayleigh numbers with especially small grid-distances in the near-wall region. The computed boundary layer velocity profiles are in excellent agreement with available experimental data. The local heat transfer in the turbulent part of the boundary layers is predicted 20% too high. Computations were carried out for the natural convective flow in a room induced by a hot radiator and a cold window. Various radiator configurations and types of thermal boundary conditions were applied including thermal radiation interaction between surfaces.Nomenclature a thermal diffusivity (m²/s) - C constant in _t expression - D cavity dimensions (m) - g acceleration of gravity (m/s²) - G _k production/destruction of k by buoyancy (kg/ms³) - h enthalpy (J/kg) - IX index of grid point - k turbulent kinetic energy (m²/s²) - m dimensionless stratification parameter - Nu overall Nusselt number - Nu _y local Nusselt number - NX total number of grid points - p pressure (N/m²) - P _k production of k by shear stress (kg/ms³) - Q heat flux through wall (W/m) - Ra overall Rayleigh number - Ra _y local Rayleigh number - Re _t turbulent Reynolds number - S source term in -equation (kg/ms⁴) - S source term for - T _c, T _h temperatures of cold and hot walls (K) - T _s (y) stratification temperature on vertical mid-line (K) - T ₀ mean cavity temperature (K) - u, v horizontal and vertical velocity components (m/s) - u ₀ Brunt-Vaisälä velocity scale (m/s) - x, y horizontal and vertical coordinates (m) - non-linearity parameter for grid - coefficient of thermal expansion (l/K) - jet angle (°) - diffusivity for - S dissipation rate for turbulent kinetic energy (m²/s³) - variable to be solved - thermal conductivity (W/mK) - , _t kinematic and eddy viscosities (m²/s) - stream function (kg/ms) - density (kg/m³) - _k, , _t constants in k– model     

Influence of thermophysical properties on the thermal stability of insulated composite superconducting tape

The effect of thermal conductivity and heat capacity on thermal stability of Nb-Ti tape superconductor stabilized with copper and subjected to transient thermal disturbance, was numerically investigated. The problem was solved by using the three- dimensional heat conduction equation. The results show that the anisotropy of thermophysical properties of the superconductor have significant effect on the thermal stability. It is found that the thermal stability of the tape is improved by increasing the heat capacity and decreasing the thermal conductivity. The best limits for anisotrpy factors α _k and α _c are (1.0; 1.5) and (2.0; 2.5), respectively.     

The effects of span position of winglet vortex generator on local heat/mass transfer over a three-row flat tube bank fin

The naphthalene sublimation method was used to study the effects of span position of vortex generators (VGs) on local heat transfer on three-row flat tube bank fin. A dimensionless factor of the larger the better characteristics, JF, is used to screen the optimum span position of VGs. In order to get JF, the local heat transfer coefficient obtained in experiments and numerical method are used to obtain the heat transferred from the fin. A new parameter, named as staggered ratio, is introduced to consider the interactions of vortices generated by partial or full periodically staggered arrangement of VGs. The present results reveal that: VGs should be mounted as near as possible to the tube wall; the vortices generated by the upstream VGs converge at wake region of flat tube; the interactions of vortices with counter rotating direction do not effect Nusselt number (Nu) greatly on fin surface mounted with VGs, but reduce Nu greatly on the other fin surface; the real staggered ratio should include the effect of flow convergence; with increasing real staggered ratio, these interactions are intensified, and heat transfer performance decreases; for average Nu and friction factor (f), the effects of interactions of vortices are not significant, f has slightly smaller value when real staggered ratio is about 0.6 than that when VGs are in no staggered arrangement. A cross section area of flow passage [m²] - A _mim minimum cross section area of flow passage [m²] - a width of flat tube [m] - b length of flat tube [m] - B _pT lateral pitch of flat tube: B _pT = S ₁/T _p - d _h hydraulic diameter of flow channel [m] - D _naph diffusion of naphthalene [m²/s] - f friction factor: f = pd _h/(Lu ² _max/2) - h mass transfer coefficient [m/s] - H height of winglet type vortex generators [m] - j Colburn factor [–] - JF a dimensionless ratio, defined in Eq. (23) [–] - L streamwise length of fin [m] - L _PVG longitudinal pitch of vortex generators divided by fin spacing: L _pVG = l _VG/T _p - l _VG pitch of in-line vortex generators [m] - m mass [kg] - m mass sublimation rate of naphthalene [kg/m²·s] - Nu Nusselt number: Nu = d _h/ - P pressure of naphthalene vapor [Pa] - p non-dimensional pitch of in-line vortex generators: p = l _VG/S ₂ - Pr Prandtl number [–] - Q heat transfer rate [W] - R universal gas constant [m²/s²·K] - Re Reynolds number: Re = ·u _max·d _h/ - S ₁ transversal pitch between flat tubes [m] - S ₂ longitudinal pitch between flat tubes [m] - Sc Schmidt number [–] - Sh Sherwood number [–]: Sh = hd _h/D _naph - Sr staggered ratio [–]: Sr = (2Hsin – C)/(2Hsin) - T _p fin spacing [m] - T temperature [K] - u _max maximum velocity [m/s] - u average velocity of air [m/s] - V volume flow rate of air [m³/s] - x,y,z coordinates [m] - z sublimation depth[m] - heat transfer coefficient [W/m²·K] - heat conductivity [W/m·K] - viscosity [kg/m²·s] - density [kg/m³] - attack angle of vortex generator [°] - time interval for naphthalene sublimation [s] - fin thickness, distance between two VGs around the tube [m] - small interval - C distance between the stream direction centerlines of VGs - p pressure drop [Pa] - 0 without VG enhancement - 1, 2, I, II fin surface I, fin surface II, respectively - atm atmosphere - f fluid - fin fin - local local value - m average - naph naphthalene - n,b naphthalene at bulk flow - n,w naphthalene at wall - VG with VG enhancement - w wall or fin surface