Influence of Mainstream Turbulence on Heat Transfer Coefficients From a Gas Turbine Blade

1994 ◽  
Vol 116 (4) ◽  
pp. 896-903 ◽  
Author(s):  
L. Zhang ◽  
J.-C. Han
Keyword(s):  
Heat Transfer ◽  
Boundary Layer ◽  
Reynolds Number ◽  
Heat Transfer Coefficient ◽  
Gas Turbine ◽  
Surface Heat ◽  
Transfer Coefficients ◽  
Suction Surface ◽  
Heat Transfer Coefficients ◽  
The Heat Transfer Coefficient

The influence of mainstream turbulence on surface heat transfer coefficients of a gas turbine blade was studied. A five-blade linear cascade in a low-speed wind tunnel facility was used in the experiments. The mainstream Reynolds numbers were 100,000, 200,000, and 300,000 based on the cascade inlet velocity and blade chord length. The grid-generated turbulence intensities at the cascade inlet were varied between 2.8 and 17 percent. A hot-wire anemometer system measured turbulence intensities, mean and time-dependent velocities at the cascade inlet, outlet, and several locations in the middle of the flow passage. A thin-foil thermocouple instrumented blade determined the surface heat transfer coefficients. The results show that the mainstream turbulence promotes earlier and broader boundary layer transition, causes higher heat transfer coefficients on the suction surface, and significantly enhances the heat transfer coefficient on the pressure surface. The onset of transition on the suction surface boundary layer moves forward with increased mainstream turbulence intensity and Reynolds number. The heat transfer coefficient augmentations and peak values on the suction and pressure surfaces are affected by the mainstream turbulence and Reynolds number.

Influence of Unsteady Wake on Heat Transfer Coefficient From a Gas Turbine Blade

1993 ◽  
Vol 115 (4) ◽  
pp. 904-911 ◽  
Author(s):  
J.-C. Han ◽  
L. Zhang ◽  
S. Ou
Keyword(s):  
Heat Transfer ◽  
Reynolds Number ◽  
Gas Turbine ◽  
Strouhal Number ◽  
Turbine Blade ◽  
Surface Heat ◽  
Transfer Coefficients ◽  
Heat Transfer Coefficients ◽  
Unsteady Wake ◽  
Wake Passing

The effect of unsteady wake on surface heat transfer coefficients of a gas turbine blade was experimentally determined using a spoked wheel type wake generator. The experiments were performed with a five-airfoil linear cascade in a low-speed wind tunnel facility. The cascade inlet Reynolds number based on the blade chord was varied from 1 to 3 × 105. The wake Strouhal number was varied between 0 and 1.6 by changing the rotating wake passing frequency (rod speed and rod number), rod diameter, and cascade inlet velocity. A hot-wire anemometer system was located at the cascade inlet to detect the instantaneous velocity, phase-averaged mean velocity, and turbulence intensity induced by the passing wake. A thin foil thermocouple instrumented blade was used to determine the surface heat transfer coefficients. The results show that the unsteady passing wake promotes earlier and broader boundary layer transition and causes much higher heat transfer coefficients on the suction surface, whereas the passing wake also significantly enhances heat transfer coefficients on the pressure surface. The blade heat transfer coefficients for a given Reynolds number flow increase with the wake Strouhal number by increasing the rod speed, rod number, or rod diameter. For a given wake passing frequency and rod diameter, the blade heat transfer coefficients decrease with decreasing Reynolds number, although the corresponding wake Strouhal number is increased. The results suggest that both the Reynolds and Strouhal numbers are important parameters in determining the blade heat transfer coefficients in unsteady wake flow conditions.

Flow and Heat Transfer of Micro-Tube Bank

2010 ◽  
Author(s):  
Yasuo Koizumi ◽  
Atsushi Katsuta ◽  
Hiroyasu Ohtake
Keyword(s):  
Heat Transfer ◽  
Reynolds Number ◽  
Heat Transfer Coefficient ◽  
Transfer Coefficient ◽  
Transfer Coefficients ◽  
Tube Bank ◽  
Heat Transfer Coefficients ◽  
Line Array ◽  
The Difference ◽  
The Heat Transfer Coefficient

Heat transfer and flow behavior in a mini-tube bank was examined. The tube bank was simulated with wires of 1 mm diameter. The wires were arranged in the 5×5 in-line array and the 5×5 staggered array with the arranging pitch = 3. Experiments were performed in the range of the tube Reynolds number Re = 4 ∼ 3,500. Numerical analyses were also performed with the commercial CFD code of STAR-CD. The heat transfer coefficient of the tube of the first row was well expressed with the existing heat transfer correlations. In the case of the in-line array, unlike usual sized tube banks, the measured heat transfer coefficients of the tubes after the second row were lower than those of the first row and the difference between those increased as the Reynolds number was increased. At approximately Reynolds number ≃ 50, the difference turned to decrease; the heat transfer coefficients initiate to recover to the first row value. Then, the heat transfer coefficient in the rear row became larger at approximately Re ≃ 1,000 than that of the first row. In the case of the staggered array, the decrease in the heat transfer coefficient in the rear row was smaller than that in the case of the in-line array. The recovery of the heat transfer coefficient to the first row value started at a little bit lower Reynolds number and it exceeded the first row value at approximately Re ≃ 700. The flow visualization results and also the STAR-CD analytical results indicated that when the Reynolds number was low, the wake region of the preceding tube was stagnant. This flow stagnation caused the heat transfer deterioration in the front part of the rear tube, which resulted in the lower heat transfer coefficient of the rear tube than that of the first row. As the Reynolds number was increased, the flow state in the wake region changed from the stagnant condition to the more disturbed condition by periodical shedding of the Karman vortex. This change caused the recovery of the heat transfer in the front region of the rear tube, which resulted in the recovery of the heat transfer coefficient of the rear tube.

scholarly journals The Heat Transfer Coefficient Predictions in Engineering Applications

2021 ◽  
Vol 2108 (1) ◽  
pp. 012022
Author(s):  
Junchi Wan
Keyword(s):  
Heat Transfer ◽  
Boundary Layer ◽  
Heat Transfer Coefficient ◽  
Transfer Coefficient ◽  
Temperature Ratio ◽  
Transfer Coefficients ◽  
Engineering Applications ◽  
Reynolds Analogy ◽  
Heat Transfer Coefficients ◽  
The Heat Transfer Coefficient

Abstract Most engineering applications have boundary layers; the convective transport of mass, momentum and heat normally occurs through a thin boundary layer close to the wall. It is essential to predict the boundary layer heat transfer phenomenon on the surface of various engineering machines through calculations. The experimental, analogy and numerical methods are the three main methods used to obtain convective heat transfer coefficient. The Reynolds analogy provides a useful method to estimate the heat transfer rate with known surface friction. In the Reynolds analogy, the heat transfer coefficient is independent of the temperature ratio between the wall and the fluid. Other methods also ignore the effect of the temperature ratio. This paper summarizes the methods of predicting heat transfer coefficients in engineering applications. The effects of the temperature ratio between the wall and the fluid on the heat transfer coefficient predictions are studied by summarizing the researches. Through the summary, it can be found that the heat transfer coefficients do show a dependence on the temperature ratio. And these effects are more obvious in turbulent flow and pointing out that the inaccuracy in the determination of the heat transfer coefficient and proposing that the conjugate heat transfer analysis is the future direction of development.

Study on Heat Transfer and Flow Analysis of 5×5 and 5×1 Mini-Tube Bank of Micro Heat Exchanger

2007 ◽  
Author(s):  
M. Arai ◽  
Y. Koizumi ◽  
H. Ohtake
Keyword(s):  
Heat Transfer ◽  
Reynolds Number ◽  
Heat Transfer Coefficient ◽  
Transfer Coefficient ◽  
Flow State ◽  
Wake Region ◽  
Transfer Coefficients ◽  
Tube Bank ◽  
Heat Transfer Coefficients ◽  
The Heat Transfer Coefficient

Heat transfer and flow behavior in the mini rod bank were examined. The tube bank was simulated with 5 wires of 1 mm diameter. The wires were arranged on the center line of the flow channel of 30 mm wide, 15 mm high and 300 mm long. The pitch between wires were varied from 1.5 mm to 9 mm. Experiments were performed in the range of the rod Re = 1 ∼ 400, i.e. the flow velocity in the channel was in the range of 0.0036 m/s ∼ 0.34 m/s. The measured heat transfer coefficients of the first row were a little bit higher than, rather close to, the predicted values by the correlations. The heat transfer coefficients after the second row were lower than those of the first row. The difference between those increased as the Reynolds number was increased. Around Reynolds number = 100, the difference turned to decrease. After the occurrence of the heat transfer coefficient recovery in the rows after the second row, the deeper the row was, the larger the heat transfer coefficient was. The flow visualization results and the analytical results by the STAR-CD code indicated that when the Reynolds number was low, the wake region of the preceding rod was stagnant. This flow stagnation caused the heat transfer coefficient deterioration around the stagnation point of the rear rod. As the Reynolds number was increased, the flow state in the wake region changed from the stagnant condition to the more disturbed condition by periodical shedding of the Karman vortex from the preceding rod. This agitation of the wake region by the vortices caused the recovery of the deteriorated heat transfer coefficients. The deeper the row was, the more disturbed the wake flow state was. The measured average heat transfer coefficients of the tube bank agreed well with the analytical results by the STAR-CD code. The measured and the analyzed results were close to the predicted values by correlations.

The Effect of a Spherical Protuberance on the Local Heat Transfer to a Turbulent Boundary Layer

1968 ◽  
Vol 90 (4) ◽  
pp. 408-412 ◽  
Author(s):  
R. A. Seban ◽  
G. L. Caldwell
Keyword(s):  
Heat Transfer ◽  
Boundary Layer ◽  
Heat Transfer Coefficient ◽  
Boundary Layer Thickness ◽  
Local Heat Transfer ◽  
Local Heat ◽  
Sphere Diameter ◽  
Transfer Coefficients ◽  
Heat Transfer Coefficients ◽  
The Heat Transfer Coefficient

Local heat transfer coefficients are presented for a single spherical protuberance on a plate, along which the boundary layer was turbulent, for air speeds from 50 to 150 fps. Two spheres were used to produce ratios of sphere diameter to boundary-layer thickness of the order of 2 and 0.7. The heat transfer coefficient behind the sphere depends approximately on the eight-tenths power of the velocity, its maximum is located about 2 dia downstream of the sphere, and the downstream effect is limited spanwise to a region about 4 dia in width.

scholarly journals Systematic heat transfer measurements in highly viscous binary fluids

2021 ◽  
Author(s):  
Ann-Christin Fleer ◽  
Markus Richter ◽  
Roland Span
Keyword(s):  
Heat Transfer ◽  
Heat Transfer Coefficient ◽  
Transfer Coefficient ◽  
Volatile Component ◽  
Test Tube ◽  
Heat Fluxes ◽  
Transfer Coefficients ◽  
Heat Transfer Coefficients ◽  
Low Viscosity ◽  
The Heat Transfer Coefficient

AbstractInvestigations of flow boiling in highly viscous fluids show that heat transfer mechanisms in such fluids are different from those in fluids of low viscosity like refrigerants or water. To gain a better understanding, a modified standard apparatus was developed; it was specifically designed for fluids of high viscosity up to 1000 Pa∙s and enables heat transfer measurements with a single horizontal test tube over a wide range of heat fluxes. Here, we present measurements of the heat transfer coefficient at pool boiling conditions in highly viscous binary mixtures of three different polydimethylsiloxanes (PDMS) and n-pentane, which is the volatile component in the mixture. Systematic measurements were carried out to investigate pool boiling in mixtures with a focus on the temperature, the viscosity of the non-volatile component and the fraction of the volatile component on the heat transfer coefficient. Furthermore, copper test tubes with polished and sanded surfaces were used to evaluate the influence of the surface structure on the heat transfer coefficient. The results show that viscosity and composition of the mixture have the strongest effect on the heat transfer coefficient in highly viscous mixtures, whereby the viscosity of the mixture depends on the base viscosity of the used PDMS, on the concentration of n-pentane in the mixture, and on the temperature. For nucleate boiling, the influence of the surface structure of the test tube is less pronounced than observed in boiling experiments with pure fluids of low viscosity, but the relative enhancement of the heat transfer coefficient is still significant. In particular for mixtures with high concentrations of the volatile component and at high pool temperature, heat transfer coefficients increase with heat flux until they reach a maximum. At further increased heat fluxes the heat transfer coefficients decrease again. Observed temperature differences between heating surface and pool are much larger than for boiling fluids with low viscosity. Temperature differences up to 137 K (for a mixture containing 5% n-pentane by mass at a heat flux of 13.6 kW/m2) were measured.

An Experimental Study of Heat Transfer on an Outlet Guide Vane

2014 ◽  
Author(s):  
Chenglong Wang ◽  
Lei Wang ◽  
Bengt Sundén ◽  
Valery Chernoray ◽  
Hans Abrahamsson
Keyword(s):  
Heat Transfer ◽  
Boundary Layer ◽  
Reynolds Number ◽  
Turbulent Boundary Layer ◽  
Guide Vane ◽  
Incidence Angle ◽  
Boundary Layer Transition ◽  
Transfer Coefficients ◽  
Suction Surface ◽  
Layer Transition

In the present study, the heat transfer characteristics on the suction and pressure sides of an outlet guide vane (OGV) are investigated by using liquid crystal thermography (LCT) method in a linear cascade. Because the OGV has a complex curved surface, it is necessary to calibrate the LCT by taking into account the effect of viewing angles of the camera. Based on the calibration results, heat transfer measurements of the OGV were conducted. Both on- and off-design conditions were tested, where the incidence angles of the OGV were 25 degrees and −25 degrees, respectively. The Reynolds numbers, based on the axial flow velocity and the chord length, were 300,000 and 450,000. In addition, heat transfer on suction side of the OGV with +40 degrees incidence angle was measured. The results indicate that the Reynolds number and incidence angle have considerable influences upon the heat transfer on both pressure and suction surfaces. For on-design conditions, laminar-turbulent boundary layer transitions are on both sides, but no flow separation occurs; on the contrary, for off-design conditions, the position of laminar-turbulent boundary layer transition is significantly displaced downstream on the suction surface, and a separation occurs from the leading edge on the pressure surface. As expected, larger Reynolds number gives higher heat transfer coefficients on both sides of the OGV.

The Simulation Study of Turbulence Models for Conjugate Heat Transfer Analysis of a High Pressure Air-Cooled Gas Turbine

2010 ◽  
Author(s):  
Zhenfeng Wang ◽  
Peigang Yan ◽  
Hongfei Tang ◽  
Hongyan Huang ◽  
Wanjin Han
Keyword(s):  
Heat Transfer ◽  
Reynolds Number ◽  
Heat Transfer Coefficient ◽  
Transfer Coefficient ◽  
Gas Turbine ◽  
Turbulence Model ◽  
Conjugate Heat Transfer ◽  
Turbulence Models ◽  
Turbulence Characteristic ◽  
The Heat Transfer Coefficient

The different turbulence models are adopted to simulate NASA-MarkII high pressure air-cooled gas turbine. The experimental work condition is Run 5411. The paper researches that the effect of different turbulence models for the flow and heat transfer characteristics of turbine. The turbulence models include: the laminar turbulence model, high Reynolds number k-ε turbulence model, low Reynolds number turbulence model (k-ω standard format, k-ω-SST and k-ω-SST-γ-θ) and B-L algebra turbulence model which is adopted by the compiled code. The results show that the different turbulence models can give good flow characteristics results of turbine, but the heat transfer characteristics results are different. Comparing to the experimental results, k-ω-SST-θ-γ turbulence model results are more accurate and can simulate accurately the flow and heat transfer characteristics of turbine with transition flow characteristics. But k-ω-SST-γ-θ turbulence model overestimates the turbulence kinetic energy of blade local region and makes the heat transfer coefficient higher. It causes that local region temperature is higher. The results of B-L algebra turbulence model show that the results of B-L model are accurate besides it has 4% temperature error in the transition region. As to the other turbulence models, the results show that all turbulence models can simulate the temperature distribution on the blade pressure surface except the laminar turbulence model underestimates the heat transfer coefficient of turbulence flow region. On the blade suction surface with transition flow characteristics, high Reynolds number k-ε turbulence model overestimates the heat transfer coefficient and causes the blade surface temperature is high about 90K than the experimental result. Low Reynolds number k-ω standard format and k-ω-SST turbulence models also overestimate the blade surface temperature value. So it can draw a conclusion that the unreasonable choice of turbulence models can cause biggish errors for conjugate heat transfer problem of turbine. The combination of k-ω-SST-γ-θ model and B-L algebra model can get more accurate turbine thermal environment results. In addition, in order to obtain the affect of different turbulence models for gas turbine conjugate heat transfer problem. The different turbulence models are adopted to simulate the different computation mesh domains (First case and Second case). As to each cooling passages, the first case gives the wall heat transfer coefficient of each cooling passages and the second case considers the conjugate heat transfer course between the cooling passages and blade. It can draw a conclusion that the application of heat transfer coefficient on the wall of each cooling passages avoids the accumulative error. So, for the turbine vane geometry models with complex cooling passages or holes, the choice of turbulence models and the analysis of different mesh domains are important. At last, different turbulence characteristic boundary conditions of turbine inner-cooling passages are given and K-ω-SST-γ-θ turbulence model is adopted in order to obtain the effect of turbulence characteristic boundary conditions for the conjugate heat transfer computation results. The results show that the turbulence characteristic boundary conditions of turbine inner-cooling passages have a great effect on the conjugate heat transfer results of high pressure gas turbine.

Darryl E. Metzger Memorial Session Paper: Comparison of Calculated and Measured Heat Transfer Coefficients for Transonic and Supersonic Boundary-Layer Flows

1995 ◽  
Vol 117 (2) ◽  
pp. 248-254 ◽  
Author(s):  
C. Hu¨rst ◽  
A. Schulz ◽  
S. Wittig
Keyword(s):  
Heat Transfer ◽  
Boundary Layer ◽  
Reynolds Number ◽  
High Speed ◽  
Three Dimensional ◽  
Supersonic Boundary Layer ◽  
Nozzle Wall ◽  
Transfer Coefficients ◽  
Heat Transfer Coefficients ◽  
Three Dimensional Finite Element

The present study compares measured and computed heat transfer coefficients for high-speed boundary layer nozzle flows under engine Reynolds number conditions (U∞=230 ÷ 880 m/s, Re* = 0.37 ÷ 1.07 × 106). Experimental data have been obtained by heat transfer measurements in a two-dimensional, nonsymmetric, convergent–divergent nozzle. The nozzle wall is convectively cooled using water passages. The coolant heat transfer data and nozzle surface temperatures are used as boundary conditions for a three-dimensional finite-element code, which is employed to calculate the temperature distribution inside the nozzle wall. Heat transfer coefficients along the hot gas nozzle wall are derived from the temperature gradients normal to the surface. The results are compared with numerical heat transfer predictions using the low-Reynolds-number k–ε turbulence model by Lam and Bremhorst. Influence of compressibility in the transport equations for the turbulence properties is taken into account by using the local averaged density. The results confirm that this simplification leads to good results for transonic and low supersonic flows.

An Experimental Investigation of the Rib Surface-Averaged Heat Transfer Coefficient in a Rib-Roughened Square Passage

1997 ◽  
Vol 119 (2) ◽  
pp. 381-389 ◽  
Author(s):  
M. E. Taslim ◽  
C. M. Wadsworth
Keyword(s):  
Heat Transfer ◽  
Heat Transfer Coefficient ◽  
Transfer Coefficient ◽  
Thermal Performance ◽  
Blockage Ratio ◽  
Transfer Coefficients ◽  
Height Ratio ◽  
Average Heat ◽  
Heat Transfer Coefficients ◽  
The Heat Transfer Coefficient

Turbine blade cooling, a common practice in modern aircraft engines, is accomplished, among other methods, by passing the cooling air through an often serpentine passage in the core of the blade. Furthermore, to enhance the heat transfer coefficient, these passages are roughened with rib-shaped turbulence promoters (turbulators). Considerable data are available on the heat transfer coefficient on the passage surface between the ribs. However, the heat transfer coefficients on the surface of the ribs themselves have not been investigated to the same extent. In small aircraft engines with small cooling passages and relatively large ribs, the rib surfaces comprise a large portion of the passage heat transfer area. Therefore, an accurate account of the heat transfer coefficient on the rib surfaces is critical in the overall design of the blade cooling system. The objective of this experimental investigation was to conduct a series of 13 tests to measure the rib surface-averaged heat transfer coefficient, hrib, in a square duct roughened with staggered 90 deg ribs. To investigate the effects that blockage ratio, e/Dh and pitch-to-height ratio, S/e, have on hrib and passage friction factor, three rib geometries corresponding to blockage ratios of 0.133, 0.167, and 0.25 were tested for pitch-to-height ratios of 5, 7, 8.5, and 10. Comparisons were made between the rib average heat transfer coefficient and that on the wall surface between two ribs, hfloor, reported previously. Heat transfer coefficients of the upstream-most rib and that of a typical rib located in the middle of the rib-roughened region of the passage wall were also compared. It is concluded that: 1 The rib average heat transfer coefficient is much higher than that for the area between the ribs; 2 similar to the heat transfer coefficient on the surface between the ribs, the average rib heat transfer coefficient increases with the blockage ratio; 3 a pitch-to-height ratios of 8.5 consistently produced the highest rib average heat transfer coefficients amongst all tested; 4 under otherwise identical conditions, ribs in upstream-most position produced lower heat transfer coefficients than the midchannel positions, 5 the upstream-most rib average heat transfer coefficients decreased with the blockage ratio; and 6 thermal performance decreased with increased blockage ratio. While a pitch-to-height ratio of 8.5 and 10 had the highest thermal performance for the smallest rib geometry, thermal performance of high blockage ribs did not change significantly with the pitch-to-height ratio.

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