Experimental Analysis Of Heat Transfer Over Compound Dimpled-Ribbed Surface

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Volume 2, Special Issue 1 MEPCON 2015

131 Available online at www.ijiere.com

International Journal of Innovative and Emerging

Research in Engineering

e-ISSN: 2394 - 3343 p-ISSN: 2394 - 5494

Experimental Analysis of Heat Transfer Over Compound

Dimpled-Ribbed Surface

Pavan Jadhao

a

_{, S.H.Mankar,}

b

a,b_{Trinity college of Engg. And research, Pisoli, Pune, India.}

ABSTRACT:

The variation of the heat transfer coefficients in the channel with both angled ribs and dimples is measured using the experimental setup. To make the comparison, heat transfer coefficients for dimpled plate and ribbed plates were also presented. The channel aspect ratio was kept 2. The rib pitch is 50 mm, rib angles are 900,600 and 400, dimple diameter, and dimple center-to-center distance was 5 mm, and 8 mm, respectively. The Reynolds number based on the channel hydraulic diameter ranged between 12,000 and 30,000. Experimental Results show that the distribution of heat transfer coefficient was asymmetric due to the secondary flow induced by the ribs provided with different angles. Also, dimples produced between the ribs increased the heat coefficient with a considerable increase in pressure drop. Thus, the compound cooling technique with angled rib sand dimples should be considered as a one of the solution for improving the heat transfer performance of a gas turbine blade cooling technique.

Keywords: Dimple, Vortex, VG’s, Rib, friction factor.

I. INTRODUCTION

Number of researchers has taken the efforts for reducing the consumption of nonrenewable energy. Improving the efficiency of heat transfer is useful in applications such as macro and micro scale heat exchangers, gas turbine internal airfoil cooling, fuel elements of nuclear power plants, electronic devices cooling, combustion chamber liners, biomedical devices, powerful semiconductor devices etc. The heat transfer enhancement using dimples recently proving good option due to its relatively low pressure-loss characteristics. Dimples promote turbulent mixing in the flow and enhance the heat transfer, as they behave as a vortex generator. This paper is concerned with experimental set up for enhancement of the forced convection heat transfer over the compound dimpled and ribbed surface. The objective of the present work is to find out the heat transfer rate and all the results obtained will be compared with a flat surface result. The parametric study is carried for i) mass flow rate of air ii) Reynolds number iii) Dimple rib arrangement on the plate iv) heat input v) Nusselt number VI) Friction factor vii) Heat transfer coefficient.

II. STUDIESONHEATTRANSFERENHANCEMENT

IftikarAhemad H. Patel et al. [1] investigated heat transfer enhancement over the dimpled surface. The main objective of his experiment ware to find out the heat transfer and air flow distribution on dimpled surfaces and all the results obtained are compared with those from a flat surface. For obtaining the results, the spherical type dimples were fabricated, and the diameter and the depth of dimple were 6 mm and 3 mm respectively. Channel height was 25.4mm, two dimple configurations were tested. The Reynolds number based on the channel hydraulic diameter was varied from 5000 to 15000. From experimentation it was observed that thermal performance is increasing with increase in Reynolds number. But the thermal performance of inline dimples arrangement is poor as compared to the plate with staggered dimple arrangement.

Moon et al. [2] studied the channel height effecton heat transfer over the dimpled surfaces. Heat transfer coefficient and friction factors were experimentally investigated in rectangular channels which had dimples on one wall. The heat transfer coefficients were measured for relative channel heights (H/D ratio of 0.37, 0.74, 1.11 and 1.49) in a Reynolds number range from 12,000 to 60,000.The heat transfer enhancement was reported mostly outside of the dimples. The heat transfer enhancement was lowest on the upstream dimpled wall and highest in the vicinity of the downstream edge of the dimple. The heat transfer coefficient distribution exhibited a similar pattern throughout the studied H/D range (0.37<H/D<1.49). The heat transfer coefficient on a dimpled wall was approximately constant at a value of 2.1 times that of a smooth channel over the entire H/D range in the thermally developed region. The heat transfer enhancement ratio reported was independent with Reynolds number the friction factors in the aerodynamically fully developed region were consistently measured to be around 0.0412(only 1.6 to 2.0 times that of a smooth channel) and found relatively independent of the Reynolds number and channel height.

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132 for Reynolds number varying from 1250 to 61500. These include flow visualizations, stream wise velocity and local Nusselt numbers. The H/D ratio was kept constant as 0.5.. For Reynolds number from 10,200 to 13,800 ratios of dimple surface friction factor to smooth surface friction factors reported were 1.5 – 1.55. Such values along with heat transfer augmentations at the same Reynolds numbers provide further evidence of the feasibility of dimpled passages for internal turbine airfoil cooling.

Mahmood and Ligrani [4], analyzedexperimentally the influence of dimple aspect ratio, temperature ratio, Reynolds number and flow structures in a dimpled channel at Reynolds number varying from600 to 11,000 and H/D ratio varying as0.20, 0.25, 0.5 and 1.00. The results showed that the vortex pairs which were periodically shed from the dimples become stronger as channel height decreases with respect to the imprint diameter.

M. A. Saleh, H.E.Abdel-Hameed [5], studiedthe flow and heat transfer performance of a parallel/ counter flow heat exchanger, when the heat transfer surface is provided with dimples on one or both sides i.e. on cold fluid side and hot fluid side. The experimental set up consists of two parallel identically and geometrically passages: one for the hot fluid and the other for the cold fluid. The average duct height is 10 mm and duct width is 110 mm The Results consist of flow characteristics (mainly pressure distributions) and heat transfer characteristics (Nusselt number distributions) comparison against the non-dimpled case (smooth surface) was held. Authors also studied that the cases with various dimples depths (d/D= 0.2, 0.3 and 0.4) and arrangements (in line and staggered) were tested over a range of Reynolds number (50 to 3000). It was found that the overall heat transfer rates was 2.5 times greater for the dimpled surface compared to a smooth surface and the pressure drop penalties in the range of 1.5-2.0 over smooth surfaces.

Burgess and Ligrani [6] showed theexperimental results for the dimple depth to dimpleprint diameter (δ/D) ratios varying as 0.1, 0.2, and 0.3 to provide information on the influences of dimple depth. They reported that at all Reynolds numbers considered, Nusselt number augmentations increases as dimple depth increases (and all other experimental and geometric parameters held approximately constant). The data presented includes friction factors, local Nusselt numbers and averaged Nusselt numbers. For all Reynolds numbers considered results showed increasing globally averaged Nusselt number ratios as δ/D ratio increases from 0.1 to 0.3.

III.FLOWMECHANISMOVERADIMPLEDSURFACE

Dimpled surfaces are commonly known for their drag reduction characteristics in external flows over bodies. This is because dimples cause a change in the critical Reynolds number [9]. Figure 1 shows a figure for flow velocity profile on the vehicle’s centerline plane near the roof end. This leads to downstream pressure rise, which in turn generates reverse force acting against the main flow and generates reverse flow at downstream point C. No reverse flow occurs at point A located further upstream of point C because the momentum of the boundary layer is overcoming over the pressure gradient between points A and C, there is separation point B, where the pressure gradient and the momentum of the boundary layer are eqibalanced. In the lower zone close to the vehicle’s surface within the boundary layer, the airflow quickly loses its momentum as it moves downstream due to the viscosity of air which results in reversal of air flow. The purpose of adding VGs is to supply the momentum from higher region where airflow has large momentum compared with lower region where it has small momentum value. It is possible due to the stream wise vortices generated from vortex generators located just before the separation point, as shown in Figure.2. This allows the separation point to shift further downstream. Shifting the separation point downstream makes the expanded airflow to persist proportionately longer, the flow velocity at the separation point to become slower, and consequently the static pressure to become higher. The static pressure at the separation point governs over all pressures in the entire flow separation region. It works to reduce drag by increasing the back pressure. Shifting the separation point downstream, therefore, provides dual advantages in drag reduction: one is to narrow the separation region in which low pressure constitutes the cause of drag; another is to raise the pressure of the flow separation region. A combination of these two effects reduces the drag acting on the vehicle.

Figure1: Schematics of Figure2: Schematics of flowvelocity profile around

vortex Generator at Rear end

Delaying the flow separation and the drag by itself. The effect of delaying the flow separation point, however, saturates at a certain level, which suggests that there must be an optimum size for VGs. Thus although the purpose of using VGs is to control flow separation at the roof end of a sedan, it is so similar to the purpose of using VGs on aircraft [1].

A. Vortex Heat Transfer Enhancement Technique.

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133 in Figure3. Vortex Heat Transfer Enhancement commonly known as VHTE is the Enhancement of heat transfer by a system of 3-Dimensional surface cavities called as dimples having specific geometry, dimensions and mutual orientation. Each dimple acts as a “vortex generator” which provides an intensive and stable heat and mass transfer between the dimpled surface and gaseous heating/cooling media. [1]

Figure 3: VHTE Mechanism [1].

B. Objectives of the work.

From the previous studies it is observed that the extensive work has been carried out for the heat transfer enhancement on dimple surface, rough surface, and ribbed surface. All these geometries are studied separately but the combination of any two geometries can also be used for better enhancement. So, the combination of dimple plates along with the ribs is used for the experimentation. The work is concerned with experimental set up for enhancement of the forced convection heat transfer over the dimple and ribbed surface.

1. The objectives of the present work are to find out the heat transfer rate and thermal performance for the flat plate for selected range of Reynolds number.

2. To find out the heat transfer rate and thermal performance for plate with ribs only and the plate with compound

geometry of dimples and ribs with different angle of attachment.

3. Comparison of heat transfer coefficient and thermal performance of the plates with ribs only, plates with dimples

only and the plates with compound geometry of dimples and ribs, with that of flat plate. 4.

IV.METHODOLOGYANDEXPERIMENTALPROCEDURE

Test plate is of aluminium sheet of thickness 5 mm and having dimensions as 600 mm × 100 mm. The dimples produced on the test plate are of 5 mm diameter and 2 mm depth. For rectangular pattern arrangement total number of rows are employed in the stream wise direction similarly, ribs are in span wise direction with some offset. Figure 4.6 shows different types of combinations of test plates.

• Dimensions. 600 mm X 100 mm X 5 mm

• Material Aluminum • Dimple diameter 5 mm • Dimple depth 2 mm • Rib width 6 mm • Rib height 4 mm • Rib attachment angles 400_{, 60}0_{, 90}0

A schematic diagram of the experimental setup is shown in figure 4.14. It consists of an open loop flow circuit. The main components of the test apparatus are sequentially a blower (1), adapter (2), pressure measurement with manometer (3), orifice Flow Meter (4), open-close valve (5), insulation of asbestos rope or silica wool (6), Test plate (7), electric heater (8), air inlet (9), and air outlet (10). The channel inner cross section dimensions are 50 mm (wide) 50 mm height. The channel was constructed with 3 mm thick epoxy resin material with which has the temperature range of 150 0_{C to 200}0_{C. The actual photographs of experimental setup are shown in following figures 4.16 and 4.17.}

Figure 4: Layout for experimental setup. [2]

A. Experimental procedure.

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134 2. Strip plate heater was fabricated to provide heat input to the test surface. The capacity of the heater was to vary

heat input from 1000 Watts. The provision was made to fix the heater at the base of the each test plate.

3. The arrangement in the test section is sequentially uppermost is test plate bellow that heater plate and the bottom most plate is Bakelite plate.

4. The pest plates are provided with the slots for the thermocouple wires to fix into it in order to get the temperature readings of the plate during experimentation.

5. Water manometer was connected across the orifice meter to indicate the pressure differential in terms of centimetres of water column difference.

6. Pressure drop across the test section was measured using micro manometer. In the air flow bench the pipe was

used to connect the blower outlet to main test section. Next to the blower outlet, flow regulating valve is connected to the pipe to regulate the air flow.

7. Orifice meter is introduced next to the regulating valve to measure the regulated air flow rate in the pipe. Air

flows parallel to the test surface.

8. The strip plate heater fixed at the bottom of the test plate, was connected to power socket through dimmer stat.

B. Data analysis.

The mass flow rate of air is determined from the pressure drop across the orifice meter, using a following correlation.

𝑚𝑎 = 𝐶𝑑 × (

𝜋

4) × 𝑑𝑜 × 𝜌 × [

2 × 𝑔 × ℎ𝑎

1 − 𝛽4 ]

1 2

1

The useful heat gain of the air is calculated as,

𝑞 = 𝑚𝑎x 𝑐𝑝𝑎 x (𝑇𝑎𝑜 – 𝑇𝑎𝑖) 2

The heat transfer coefficient is found out by using Newton’s law of cooling, which states that the heat flux from the surface to fluid is proportional to the temperature difference between surface and fluid.

ℎ = 𝑞

(𝐴 x (𝑇𝑠 − 𝑇𝑏))

3

Nusselt number based on test plate length is defined as, 𝑁𝑢 =ℎ x 𝐿

𝐾 4

The Friction Factor (f) is evaluated by:

𝑓 = ∆𝑃 × 𝐷ℎ

2 × 𝜌𝑎 × 𝐿 × 𝑣2 5

Thermal performance Factor (n):

The thermal enhancement factor, n is defined as the ratio of heat transfer coefficient of an augmented surface, h to that of a smooth surface, ho at the same pumping power.

ɳ = (𝑁𝑢/𝑁𝑢0) 𝑓 𝑓0 1 3

6

V. RESULTSANDDISCUSSION

The test plates were fabricated such that every plates has different geometry for rib and dimple. Plates with dimples only, plate with 900_{ribs only followed by dimple plate with 90}0_{ribs, plate with 60}0_{ribs only followed by} plate with dimple and 600 _{ribs and the plate with 40}0_{ribs only followed by plate with dimples and 40}0_{ribs. All the} experimentation was done for the Reynolds number range of 18000 to 30000. Experimental results obtained are discussed below.

Figure 5: Variation of heat transfer coefficient for different plates with Reynolds number.

From the above figure 5, it is found that the heat transfer rate increases with increase in Reynolds number. The plate with only ribs shows more heat transfer rate as compared to base plate and the rate of heat transfer rate increase in further with the addition of dimples. The plate with 600_{ribs along with the dimples shows the highest heat}

0 10 20 30 40 50 60 70 80 90 100

12000 18000 24000 30000

Hea t tr a n sf er c o ef fi ci en t ( h )

Reynolds number (Re)

Dimple plate

Plate with 60 degree rib

Dimpled plate with 60 degree rib

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135 transfer coefficient. This is mainly due to the turbulence created in the dimples is more as compared to the rib only case. The more disturbance in the in the flow causes the Nusselt number to increase and increase in Nusselt number is followed by the increase in heat transfer coefficient.

Figure 6. Variation of thermal performance with Reynolds number.

Figure 6 shows the thermal performance factor versus the Reynolds number. The thermal performance factors increase as the Reynolds number increases. For the same aspect ratio case, the rib dimple compound case shows the highest thermal performance factor, followed by the rib only and dimple only cases. The thermal performance factor for the case with the same aspect ratio and for plate with 600_{rib and dimples is higher. This is due more turbulence} and strong vortex formation in this particular arrangement plate. The turbulence created leads to increase in Nusselt number which intern increases the thermal performance.

VI.CONCLUSION

The present work was conducted for experimental determination of effect of compound dimpled-ribbed surfaces on heat transfer over a flat surface under forced convection condition.

1) The Nusselt number ratio for the cases in which compound dimples and ribs are incorporated was higher than in

rib only or dimples only case.

2) The Nusselt number ratio is found higher for plate with 600_{ribs and dimples than other cases. This is possibly due} to the more suitable rib configuration for the former rather than the latter cases.

3) The value of maximum Nusselt number obtained for compound arrangement of dimples and rib is greater than

that for individual arrangement, keeping all other parameters constant. It shows that for heat transfer enhancement for compound arrangement is more effective than the rib only or dimple only arrangement.

Nomenclature-

d dimple depth

D dimple printed diameter

Dh channel hydraulic diameter

e rib height

f0 friction factor

h heat transfer coefficient

k thermal conductivity of the test section

l rib thickness

Nu Nusselt number, hDh/k

p rib pitch

Pr Prandtl number

ReDh Reynolds number based on hydraulic

Diameter.

Ma mass flow rate of air

REFERENCES

0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6

12000 18000 24000 30000

T

he

rm

al

Pe

rf

or

m

an

ce

(ɳ

)

Reynolds Number (Re)

Dimpled plate 40 degree rib For 60 degree rib

Dimpled plate with 60 degree rib Dimpled plate

Plate with 90 degree rib

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