Correlations of flow maldistribution parameters in an air cooled heat exchanger

The present paper provides correlations of flow maldistribution parameters in air‐cooled heat exchangers. The flow field in the inlet header was obtained through the numerical solution of the governing partial differential equations including the conservation equations of mass and momentum in addition to the equations of the turbulence model. The results were obtained for different number of nozzles of 2–4, different inlet flow velocities of 1–2.5m/s and different nozzle geometries in addition to incorporation of a second header. The results are presented in terms of mass flow rate distributions in the tubes of the heat exchanger and their standard deviations. The results indicate that the inlet flow velocity has insignificant influence on maldistribution while the nozzle geometry shape has a slight effect. Also, the results indicate that reducing the nozzle diameter results in an increase in the flow maldistribution. A 25% increase is obtained in the standard deviation as a result of decreasing the diameter by 25%. Increasing the number of nozzles has a significant influence on the maldistribution. A reduction of 62.5% in the standard deviation of the mass flow rate inside the tubes is achieved by increasing the number of nozzles from 2 to 4. The results indicate that incorporating a second header results in a significant reduction in the flow maldistribution. A 50% decrease in the standard deviation is achieved as a result of incorporation of a second header of seven holes. It is also found that the hole‐diameter distribution at the exit of the second header has a slight influence on the flow maldistribution. Correlations of the flow maldistribution in terms of the investigated parameters are presented. Copyright © 2007 John Wiley & Sons, Ltd.     

Refrigerant distribution in a parallel flow heat exchanger having vertical headers and heated horizontal tubes

Refrigerant R-410a flow distribution is experimentally studied in a test section simulating a parallel flow heat exchanger having vertical headers with two pass configuration. Tubes are heated to yield a test section outlet superheat of 5 °C with inlet quality of 0.3. Mass flux is varied from 50 kg/m² s to 70 kg/m² s. Effects of inlet and outlet locations are investigated in a search for an optimum configuration. Results show that, significant liquid flows through bottom channels, and less liquid is supplied to top channels. As for the inlet location, better flow distribution (pressure drop as well) is obtained for top inlet as compared with middle inlet. As for the outlet location, top or bottom outlet is better than middle outlet. Correlations are developed for the fraction of liquid or gas taken off by downstream channel as a function of header gas Reynolds number at immediate upstream. The correlations may be used to predict the liquid or gas distribution in a parallel flow heat exchanger having vertical headers. A novel thermal performance evaluation method, which accounts for tube-side flow mal-distribution is proposed.     

Further understanding of twisted tape effects as tube insert for heat transfer enhancement

Tube inserts are used as heat transfer enhancement tool for both retrofit and new design of shell and tube heat exchangers. This paper discusses and reviews the characteristics and performance of twisted tapes. The theory and application are also addressed. Industrial case study was selected to illustrate the behaviour effect that the twisted tapes impose at various laminar, transition and turbulent flow regions. This effect was demonstrated by changing the inside tube diameter and twist ratio through evaluating selected exchanger design parameters such as: local heat transfer coefficient, friction factor and pressure drop. Testing the exponent powers for Re and Pr at both laminar and turbulent regions were carried out. General design considerations are outlined for the use of twisted tapes in shell and tube heat exchangers.     

Quantitative infrared investigation of local heat transfer in a circular finned tube heat exchanger assembly

This work deals with the local heat transfer coefficient evaluation over the fin of the second row of a staggered circular finned tube heat exchanger assembly. The coefficient distribution is determined by using a transient technique and by calculating the energy balance during the fin cooling. The calculation model takes into account radiation with the surrounding and lateral heat conduction into the material. The method uses infrared measurements and integration between time bounds that depend on space. It is proposed to choose the integration bounds with an original criterion based on local heat transfer. Validation is performed on the reference case consisting in a thin plate in an aerodynamically and thermally developing channel flow. Then, distributions of Nusselt number on the circular fin are presented for several Reynolds numbers. The high resolution of the whole method and set-up allow detecting thermal imprints of developing horseshoe vortices. These imprints are analyzed by following their angular evolution around the tube.     

Solid‐particle erosion in the tube end of the tube sheet of a shell‐and‐tube heat exchanger

Erosion is one of the major problems in many industrial processes, and in particular, in heat exchangers. The effects of flow velocity and sand particle size on the rate of erosion in a typical shell‐and‐tube heat exchanger were investigated numerically using the Lagrangian particle‐tracking method. Erosion and penetration rates were obtained for sand particles of diameters ranging from 10 to 500 µm and for inlet flow velocities ranging from 0.197 to 2.95 m/s. A flow visualization experiment was conducted with the objective of verifying the accuracy of the continuous phase calculation procedure. Comparison with available experimental data of penetration rates was also conducted. These comparisons resulted in a good agreement. The results show that the location and number of eroded tubes depend mainly on the particle size and velocity magnitude at the header inlet. The rate of erosion depends exponentially on the velocity. The particle size shows negligible effect on the erosion rate at high velocity values and the large‐size particles show less erosion rates compared to the small‐size particles at low values of inlet flow velocities. The results indicated that the erosion and penetration rates are insignificant at the lower end of the velocity range. However, these rates were found to increase continuously with the increase of the inlet flow velocity for all particle sizes. The particle size creating the highest erosion rate was found to depend on the flow velocity range. Copyright © 2005 John Wiley & Sons, Ltd.     

The effect of inlet conditions on the air side hydraulic resistance and flow maldistribution in industrial air heaters

Experimental system hydraulic resistance measurements on a scale air heater unit have highlighted the excessive hydraulic resistance of typical industry configurations. Both poor header inlet conditions and large header expansion angles are shown to contribute to system hydraulic resistance magnitudes 20-100% higher than suitable benchmark cases. Typical centrifugal fan system efficiencies well under 80% multiply the system resistance effects resulting in larger fan power penalties. Velocity profile measurements taken upstream and downstream of the test heat exchanger under flow maldistribution conditions provide insight into the flow maldistribution spreading caused by the heat exchanger resistance. The anisotropic resistance of the plate fin-and-tube heat exchanger is shown to result in resistance induced flow dispersion being concentrated in the axis parallel to the plate fins.     

Two-phase flow distribution of air–water annular flow in a parallel flow heat exchanger

The air and water flow distribution are experimentally studied for a round header – flat tube geometry simulating a parallel flow heat exchanger. The number of branch flat tube is 30. The effects of tube outlet direction, tube protrusion depth as well as mass flux, and quality are investigated. The flow at the header inlet is identified as annular. For the downward flow configuration, the water flow distribution is significantly affected by the tube protrusion depth. For flush-mounted configuration, most of the water flows through frontal part of the header. As the protrusion depth increases, more water is forced to the rear part of the header. The effect of mass flux or quality is qualitatively the same as that of the protrusion depth. Increase of the mass flux or quality forces the water to rear part of the header. For the upward flow configuration, however, most of the water flows through rear part of the header. The protrusion depth, mass flux, or quality does not significantly alter the flow pattern. Possible explanations are provided based on the flow visualization results. Negligible difference on the water flow distribution was observed between the parallel and the reverse flow configuration.     

A study on the correlation between the thermal contact conductance and effective factors in fin–tube heat exchangers with 9.52 mm tube

The thermal contact resistance is a principal parameter interfering with heat transfer in a fin–tube heat exchanger. However, the thermal contact resistance in the interface between tubes and fins has not been clearly investigated. The objective of the present study is to examine the thermal contact conductance for various fin–tube heat exchangers with tube diameter of 9.52 mm and to find a correlation between the thermal contact conductance and effective factors such as expansion ratio, fin type, fin spacing and hydrophilic coating. In this study, experiments have been conducted only to measure heat transfer rate between hot and cold water. To minimize heat loss to the ambient air by the natural convection fin–tube heat exchangers have been placed in an insulated vacuum chamber. Also, a numerical scheme has been employed to calculate the thermal contact conductance with the experimental data. As a result, a new correlation including the influences of expansion ratio, slit of fin and fin coating has been introduced, and the portion of each thermal resistance has been estimated in the fin–tube heat exchangers with 9.52 mm tube.     

Enhancement of heat transfer in compact heat exchanger by different type of rib with holographic interferometry

An experimental study was carried out to investigate enhancement of heat transfer in compact heat exchanger by keeping pressure drop constant in a given range. Two different test matrices, cylindrical and triangular, used to find the optimum ribs were compared with a smooth channel. The investigation was performed with both laminar and turbulent forced flow for Reynolds numbers from 250 to 7000. The geometric parameters, in order to satisfied manufacturer demands, were fixed at p/e=6.67 and the wall temperature was held constant at 50°C. The technique of holographic interferometry was used to determine the temperature distribution in the test duct. Velocity distribution was measured using laser doppler anemometer techniques. For comparison the technique of global measurement was also used. The results revealed that cylindrical ribs are optimum heat transfer for conversion of pressure drop. An 8% experimental error was found in global measurement compared to holographic interferometry.     

Experimental studies on heat transfer and friction factor characteristics of CuO/water nanofluid under turbulent flow in a helically dimpled tube

An experimental investigation on the convective heat transfer and friction factor characteristics in the plain and helically dimpled tube under turbulent flow with constant heat flux is presented in this work using CuO/water nanofluid as working fluid. The effects of the dimples and nanofluid on the Nusselt number and the friction factor are determined in a circular tube with a fully developed turbulent flow for the Reynolds number in the range between 2500 and 6000. The height of the dimple/protrusion was 0.6 mm. The effect of the inclusion of nanoparticles on heat transfer enhancement, thermal conductivity, viscosity, and pressure loss in the turbulent flow region were investigated. The experiments were performed using helically dimpled tube with CuO/water nanofluid having 0.1%, 0.2% and 0.3% volume concentrations of nanoparticles as working fluid. The experimental results reveal that the use of nanofluids in a helically dimpled tube increases the heat transfer rate with negligible increase in friction factor compared to plain tube. The experimental results showed that the Nusselt number with dimpled tube and nanofluids under turbulent flow is about 19%, 27% and 39% (for 0.1%, 0.2% and 0.3% volume concentrations respectively) higher than the Nusselt number obtained with plain tube and water. The experimental results of isothermal pressure drop for turbulent flow showed that the dimpled tube friction factors were about 2-10% higher than the plain tube. The empirical correlations developed for Nusselt number and friction factor in terms of Reynolds number, pitch ratio and volume concentration fits with the experimental data within ±15%.     

Numerical estimation of mixed convection heat transfer in vertical helically coiled tube heat exchangers

A numerical investigation of the mixed convection heat transfer from vertical helically coiled tubes in a cylindrical shell at various Reynolds and Rayleigh numbers, various coil‐to‐tube diameter ratios and non‐dimensional coil pitches was carried out. The particular difference in this study compared with other similar studies is the boundary conditions for the helical coil. Most studies focus on constant wall temperature or constant heat flux, whereas in this study it was a fluid‐to‐fluid heat exchanger. The purpose of this article is to assess the influence of the tube diameter, coil pitch and shell‐side mass flow rate on shell‐side heat transfer coefficient of the heat exchanger. Different characteristic lengths were used in the Nusselt number calculations to determine which length best fits the data and finally it has been shown that the normalized length of the shell‐side of the heat exchanger reasonably demonstrates the desired relation. Copyright © 2010 John Wiley & Sons, Ltd.     

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     

Thermo-hydraulic characterization of a louvered fin and flat tube heat exchanger

In the present study, a whole heat exchanger with a hydraulic diameter of 2.3 mm is tested, which is a minichannel heat exchanger according to the Kandlikar classification. This is a louvered fin and flat tube heat exchanger currently used in car cooling systems, also known as radiator. A glycol-water mixture (60/40 in volume) circulates through the tubes at flows ranging from 100 to 7800 l/h and at a supply temperature of 90 °C. This fluid is cooled with ambient air at a temperature of 20 °C and at frontal air velocities varying between 0.5 and 7 m/s. The thermohydraulic performance of the heat exchanger is compared with the classical correlations given in the literature for the heat transfer and the friction factor calculation. On the glycol-water side the heat exchanger is characterized for Reynolds numbers from 30 to 8000. A first comparison is carried out with the correlations available in the literature with a purely predictive model by obtaining a predictive value with a systematic under prediction lower than 10%. In a second step a semi-empirical model is considered to identify the experimental heat transfer coefficients for this application.     

Local heat transfer in the first baffle compartment of the shell-and-tube heat exchangers for staggered tube arrangement

A mass transfer measuring method based on absorption, chemical and coupled colour reaction is used to visualize and determine the shell side local heat transfer in the first baffle compartment of shell-and-tube heat exchangers with segmental baffles for staggered tube arrangement. Local mass transfer coefficients were transformed into heat transfer coefficients by using the analogy between heat and mass transfer. The local, per-tube and integral heat transfer coefficient distributions are presented.     

Non-isothermal flow of polymer melts in a curved tube

The results of a numerical study (using finite differences) of heat transfer in polymer melt flow is presented. The rheological behaviour of the melt is described by a temperature-dependent power-law model. The curved tube wall is assumed to be at constant temperature. Convective and viscous dissipation terms are included in the energy equation. Velocity, temperature and viscosity profiles, Nusselt numbers, bulk temperatures, etc. are presented for a variety of flow conditions. Br — Brinkman number - c specific heat, J/kg K - De — Dean number - E dimensionless apparent viscosity, eq. (14d) - G dimensionless shear rate, eq. (19) - k parameter of the power-law model, °C^–1, eq. (7) - mass flow rate, kg/s - m ₀ parameter of the power-law model, Pa · s ⁿ , eq. (7) - n parameter of the power-law model, eq. (7) - Nu 2r _p/ — Nusselt number, eqs. (28,31) - p pressure, Pa - Pe — Péclet number - P —(p/)/r _c — pressure gradient, Pa/m - dissipated energy, W, eq. (29) - total energy, W, eq. (30) - r radial coordinate, m - r _c radius of tube-curvature, m, fig. 1 - r _p radius of tube, m, fig. 1 - r _t variable, m, eq. (6) - R dimensionless radial coordinate, eq. (14a) - R _c dimensionlessr _c, eq. (14a) - R _t dimensionlessr _t, eq. (14a) - t temperature, °C - bulk temperature, °C, eq. (27) - t ₀ inlet temperature of the melt, °C - t _w tube wall temperature, °C - T dimensionless temperature, eq. (14c) - T _w dimensionless tube wall temperature - T dimensionless bulk temperature - u ₁ variable, s^–1, eq. (4) - u ₂ variable, s^–1, eq. (5) - U ₁ dimensionlessu ₁, eq. (18) - U ₂ dimensionlessu ₂, eq. (18) - v velocity in-direction, m/s - average velocity of the melt, m/s - V dimensionlessv, eq. (14b) - dimensionless , eq. (15) - z r _c — centre length of the tube, m - Z dimensionlessz, eq. (14e) - heat transfer coefficient, W/m² K - shear rate, s^–1, eq. (8) - — shear rate, s^–1 - apparent viscosity, Pa · s, eq. (7) - ₀ — apparent viscosity, Pa · s - angular coordinate, rad, fig. 1 - thermal conductivity, W/m K - melt density, kg/m³ - axial coordinate, rad, fig. 1 - rate of strain tensor, s^–1, eq. (8) - (—p) pressure drop, Pa     

Pressure drop and heat transfer comparison for both microfin tube and twisted-tape inserts in laminar flow

Experiments were carried out to compare pressure drop and heat transfer coefficients for a plain, microfin, and twisted-tape insert-tubes. The twisted-tape experiments include three different twist ratios each with two different widths. The data were taken at Reynolds numbers well in the laminar region. The heat transfer data were obtained in a single shell-and-tube heat exchanger where steam is used as a heat source to obtain a uniform wall temperature and the working fluid in the tube is oil. The twist ratio and the width of the tape seem to have a large effect on the performance of the twisted-tape insert. The results demonstrate that as the twist ratio decreases, the twisted-tape will give better heat transfer enhancement. The loose-fit (W=10.8 mm) is recommended to be used in the design of heat exchanger where low twist ratios (Y=5.4, and Y=3.6) and high pressure drop situations are expected since it is easier to install and remove for cleaning purposes. Other than these situations, the tight-fit tape gives a better performance over the loose-fit tape. For the microfin tube tested in this paper, the data shows a small increase in both heat transfer and pressure drop. This type of microfin tube is not recommended to be used in laminar flow conditions.     

Pressure drop, heat transfer and performance of single-phase turbulent flow in spirally corrugated tubes

The paper presents the results of an experimental study that was carried out to determine turbulent friction and heat transfer characteristics of four spirally corrugated tubes, which have various geometrical parameters, with water and oil as the working fluids. Experiments were performed under conditions of Reynolds number varying from 6000 to 93,000 for water, and from 3200 to 19,000 for oil, respectively. The results show that the thermal performance of these tubes was superior compared to a smooth tube, but the heat transfer enhancements were not as large as the friction factor increases. Friction factors and heat transfer coefficient in these rough tubes were analyzed on the basis of momentum and heat transfer analogy, and the correlations obtained were compared with the present data and also the results of previous investigators. A mathematical model to evaluate the performance of spirally corrugated tube, which takes account of the large variation of fluid Prandtl number with temperature, was developed by the extension of previous work of Bergles and Webb. The results reported enable practical designs with standard products and optimization of tube geometry for specific conditions.     

Heat transfer and pressure loss of a very shallow fluidized-bed heat exchanger Part 2. Experiment with multirow tube banks

Experiments were carried out on heat transfer and pressure-loss characteristics of a low-pressure-loss fluidized-bed heat exchanger with an extremely small ststic bed height of glass beads. This heat exchanger was composed of a tube bankd in a staggered arrangement and a multislit distributor designed especially to reduce the pressure loss. The effect of the cross-sectional shape of the distributor, the particle size of the glass beads, the static bed height, and the number of rows on heat transfer performance were examined. The heat transfer and pressure-loss characteristics are compared with experimental data in terms of fan power consumption and the compactness of the heat exchanger.     

A note on flow and heat transfer of a viscoelastic fluid over a stretching sheet

This paper presents a study of the flow and heat transfer of an incompressible homogeneous second grade fluid past a stretching sheet. The governing partial differential equations are converted into ordinary differential equations by a similarity transformation. The effects of viscous dissipation and work due to deformation are considered in the energy equation and the variations of dimensionless surface temperature and dimensionless surface temperature gradient with various parameters are graphed and tabulated. Two cases are studied, namely, (i) the sheet with constant surface temperature (CST case) and (ii) the sheet with prescribed surface temperature (PST case).