Numerical results obtained through RSM-SED model , for the averaged Nusselt number in a rectangular duct with aspect ratio (1:2).
Rectangular ducts are widely used in heat transfer devices, for instance, in compact heat exchangers, gas turbine cooling systems, cooling channels in combustion chambers and nuclear reactors. Forced turbulent heat convection in a square or rectangular duct is one of the fundamental problems in the thermal science and fluid mechanics. Recently, Qin and Fletcher  showed that Prandtl's secondary flow of the second kind has a significant effect in the transport of heat and momentum, as revealed by the recent Large Eddy Simulation (LES) technique. Several experimental and numerical studies have been conducted on turbulent flow though of non-circular ducts: Nikuradse ; Gessner and Emery ; Gessner and Po ; Melling and Whitelaw ; Nakayama et al. ; Myon and Kobayashi ; Assato ; Assato and De Lemos ; Home et al. ; Luo et al. ; Ergin et al.  Launder and Ying ; Emery et al. ; Hirota et al.; Rokni ; Hongxing ; Yang and Hwang ; Park ;Zhang et al. ; Zheng et al. ; Su and Da Silva Neto ; Saidi and Sundén ; Rokni ; Valencia ; Sharatchandra and Rhode ; Campo et al. ; Rokni and Sundén ; Yang and Ebadian  and others. The Melling and Whitelaw´s  experimental work shows characteristics of turbulent flow in a rectangular duct where they have been used a laser-Doppler anemometer in which report the axial development mean velocity, secondary mean velocity, etc. Nakayama et al.  show the analysis of the fully developed flow field in rectangular and trapezoidal cross-section ducts; finite difference method was implemented and the model of Launder and Ying  has been used. On the other hand, Hirota et al.  present an experimental work in turbulent heat transfer in square ducts; they show details of turbulent flows and temperature fields. Likewise, Rokni  carried out a comparison of four different turbulence models for predicting the turbulent Reynolds stresses, and three turbulent heat flux models for square ducts. The literature presents various turbulence modeling in which confirm Linear Eddy Viscosity Models (LEVM) can give inaccurate predictions for Reynolds normal stresses: it does not have ability to predict secondary flows in non-circular ducts due to its isotropic treatment. In spite of that, they is one of the most popular model in the engineering due to its simplicity, good numerical stability in which can be applied in a wide variety of flows. Thus, the Nonlinear Eddy Viscosity Model (NLEVM) represents a progress of the classical LEVM in which this last one gives inequality treatment of the Reynolds normal stresses, needs of conditions for calculating turbulence-driven secondary flow in non-circular ducts and it has relatively high cost for solving the necessary two-equation formulation. The Reynolds Stress Model (RSM), also called second order or second moment closure model, is very accurate in the calculation of mean flow properties and Reynolds stresses, for simple to more complex flows including wall jets, asymmetric channel, non-circular duct and curved flows. However, RSM has some disadvantages, such as, very large computing costs. For calculating the turbulent heat fluxes, the Simple Eddy Diffusivity (SED) and Generalized Gradient Diffusion Hypothesis (GGDH) models have been adopted and investigated. Most of the works presented in the literature show results assuming constant temperature on the wall. However, in many engineering applications the heat fluxes and surface temperatures are non-constants around the duct, therefore becoming important the knowledge of the variation of the conductance around the duct, according to Kays and Crawford . According to Garcia´s developments , it is possible to carry out analysis with non-constant wall temperature boundary conditions. In this case, it is necessary to define a value that represents the mean wall temperatures in a given cross section, in which he has named T
2. Mathematical formulation
2.1. Governing equations
For analyses of fully developed turbulent flow and heat transfer, the following hypothesis has been adopted: steady state, condition of non-slip on the wall and fluid with constant properties. The turbulent Reynolds stress and the turbulent heat flux were modeled and solved by algebraic and/or differential expressions.
2.2. Turbulence models for reynolds stresses
2.2.1. Nonlinear Eddy Viscosity Model (NLEVM)
The NLEVM Model to reproduce the tensions of Reynolds, it is necessary to include non-linear terms in the basic constitutive equations. This is done by attempting to capture the sensitivity of the curvatures of the stream lines. This model is based on the initial proposal of Speziale . The Reynolds average equations, Equations (1) to (3), are applied for the device presents in the Figure 1(a) and (b).
The velocity components
The symbols and, represent the rate of the turbulent kinetic energy production and the turbulent viscosity, respectively, are expressed by:
In the present work for NLEVM, the formulations of Low Reynolds Number will be assumed for wall treatment. The damping functions and observed in Equations (5) and (6) were proposed by Abe et al . These functions and the constant and are used in equations. The subscript
This expression shows that the second term of the right side in Equation (7) represents the nonlinear relation added the original constitutive relation. This quadratic term represents the degree of anisotropy between the normal tensions of Reynolds responsible for predicting the secondary flow in non circular ducts. The values of proposed by Speziale  is equal to 1.68. In this work, will be analyzed and will adopt values for the formulation of Low Reynolds Numbers. The normal and shear tensions of Reynolds are expressed as:
The following differences for the normal tensions of Reynolds are presented and used in order to predict the anisotropy in turbulent flow at non circular ducts,
2.2.2. Reynolds Stress Model (RSM)
The most complex turbulence model is the Reynolds Stress Model (RSM), also known as second order model. It involves calculations of Reynolds stresses to an individual form,. These Reynolds stresses are used for formulating the differential equations of turbulent flow transport. The individual Reynolds stresses are utilized to close the average Reynolds equations of the momentum conservation. This model has shown superiority in relation to the two equation models (for example,) in simulating of complex flows that involve swirl, rotation, etc. The exact transport equations for the Reynolds stresses, , can be written as:
Where the letters represent: (a) Local derivative of the time; (b) Convection; (c) Turbulent diffusion; (d) Molecular diffusion; (e) Production term of stresses; (f) Buoyancy production term; (g) Pressure-stress (redistribution); (h) Dissipation term; (i) Production term for the rotation system; (j) Source term. The terms of the exact equations presented previously, , , and do not require modeling. However, the terms, , and need to be modeled to close the equations. For the present analysis, the model LRR (Launder, et al ) is chosen, which assumes that the correlation of velocity- pressure is a linear function of anisotropy tensor in the phenomenology of the redistribution,. For the wall treatment, it is also assumed the Low Reynolds numbers and periodic conditions in according to Rokni . This model had been simulated in the commercial code Fluent 6.3.
2.3. Turbulence models for turbulent heat flux
2.3.1. Simple Eddy Diffusivity (SED)
This method is based on the Boussinesq viscosity model. The turbulent diffusivity for the energy equation can be expressed as:, where the turbulent Prandtl number, for the SED model assumes a value constant in the entire region. For the air, it assumes a value equal to 0.89. The turbulent heat flux is given by,
2.3.2. Generalized Gradient Diffusion Hypothesis (GGDH)
Daly and Harlow  present the following formulation to the turbulent heat flux:
The constant, assumes the value of 0.3. The main advantage of this model is in considering the anisotropic behavior of the fluid heat transport in ducts.
2.3.3. Dimensionless energy equation for SED and GGDH models
For a given cross section of area “
Kays and Crawford  developed a formulation to rectangular cross section ducts. They considered the boundary conditions with prescribed uniform wall temperatures at the cross section, and at the duct length. According to Garcia , it is possible to carry out an analysis with non-uniform wall temperature boundary conditions. In this case, it is necessary to define a value that represents the mean wall temperatures in a given cross section, “
It is possible to develop a formula similar to Kays and Crawford , and a new expression for the turbulent energy equation can be presented as:
The following considerations are applied to obtain the variables in dimensionless form:
The fluid temperature field “
It is possible to compute the heat transfer rate per unit length on the wall surface, “
Equation (30) can be integrated to two cross sections (inlet,
From Equation (31), the bulk temperature longitudinal (
When considering uniform wall temperature, the Equations (32) and (33) are equal to zero, and for these particular conditions, it is possible to notice that these boundary conditions are not functions of “”. That simplification becomes equal to the one studied by Patankar . Equations (24)-(26), as well as the boundary conditions from Equations (32) and (33), form a set of differential equations, in which “
2.3.4. Additional equations
Additional equations were utilized for the calculation of the factor of friction Moody,; coefficient of friction Fanning,; Prandtl law; local Nusselt number for the Low Reynolds formulation (Rokni ), and Correlation of Gnielinsky, respectively. Thus, the additional equations are given by the following equations:
4. Numerical implementation
After applying the method of finite differences to the algebraic equations, to obtain the temperature fields, the following five steps indicate the developed methodology in the numerical solution. (Garcia ):
Step 1. To define the function value of the non uniform temperatures in the walls of the duct, This function can be expressed by a Fourier expansion;
Step 2. To obtain velocity field and estimated values for “”
Step 4. Dimensionless energy equation (Temperature field,
Step 5. A value for “
For all steps, “tolerance of 10-7” is the value to be accomplished by the convergence criteria, which is applicable to
5. Results and discussion
5.1. Fluid flow and heat transfer field
The Figure 2(a) shows the utilized grid (120X120) in the numerical simulation for the formulations of Low Reynolds, the Figure 2(b) it represents the secondary flow contours and comparisons of the velocity profile (NLEVM, Assato ) with the experimental work of the Melling and Whitelaw  for fluid water and Re=42000.
The predicted distributions of the friction coefficient (NLEVM and RSM) and Nusselt number (SED and GGDH) dependence on Reynolds number for fully developed flow and heat transfer in a square duct is shown in Figure 3(a) and 3(b), respectively.
Figure 4(a): comparisons of the Results (RSM-SED) numerical with the experimental for temperature profile (wall constant temperature) with fluid air and
Already the Figure 5(a) shows: The variation of the temperature profile with non-uniform wall temperature, represented by means of functions sine (Case II), south=(350-20Sin(ζ))K, north=(400-50Sin(ζ))K, east=(330+20Sin(ζ))K, west=(350+50Sin(ζ))K. where ζ is function of the radians (0-π/2) and
The Figures 6 (a) and (b), shown the temperature distribution for a rectangular duct aspect ratio (1:2) represented by means of a function sine (Case II). A third case denominated Case III is represented by: south=
In the doctoral thesis Garcia (1996) was analyzed the laminar flow coupled to the conduction and radiation in rectangular ducts and concluded that as increases the aspect ratio, the Nusselt number found in the coupling, differs from that found for ducts with constant temperature imposed around the perimeter of the section. which shows that could be making a mistake to consider the literature results without calculating the energy equation.
In the present study, the variations of the average Nusselt number for a square duct and different cases analyzed (uniform and non uniform temperature in the perimeter) are minimal. Already in the case of rectangular duct with an aspect ratio (1:2), the variations should be taken into account as shown in Table 1.
|Cases Analyzed||Reynolds number|
|Case II||65000||139, 682||-|
|Case III||65000||145, 059||-|
|Case III||28853||76, 769||-|
The results shown what for the friction factor and Nusselt number in a wide range of the Reynolds number with uniform wall temperature have a reasonable approach with the experimental works and correlation of the literature, (Figure 3(a), (b)). The Figures 4(b) and 5(a) shown new results investigated in present study, which is observed a distortion of the temperatures field and as consequence a variation of the Nusselt number caused mainly by the distribution of the non-uniform wall temperature (Case I and II, with fluid air and
Term production for the rotation system
buoyancy production Term
production Term of tensions
y+ dimensionless wall distance
Term of dissipation
distance normal to the wall
dimensionless temperature distribution
Term of pressure-tension (redistribution)
Μ dynamic viscosity
μt turbulent viscosity
turbulent Prandtl number
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