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PRESSURE DROP AND HEAT TRANSFER CHARACTERISTICS OF

LOUVERED FIN HEAT EXCHANGERS

SUPERVISED BY:

MR. SHAHRIN HISHAM BIN AMIRNORDIN

PRESENTED BY:

DJAMAL HISSEIN DIDANE

Figure 1: Flat-sided tube and louvered plate fin heat transfersurface [1].

INTRODUCTION

Heat exchangers are devices that facilitate the exchange of heat between two fluids that are at different temperatures while keeping them from mixing with each other.

Louvered fin compact heat exchangers are used extensively in several automotive applications such as radiators, oil coolers, condensers, and charge air coolers [1].

Figure 2: Section through typical louvered-fin showing key geometrical parameters [15].

INTRODUCTION cont’

In order to improve the performance of the heat exchanger fins are added on the air side. These serve several purposes:

They increase the available surface for heat transfer and interrupt the growth of the boundary layer forming along the fin surface [1].

BACKGROUND OF STUDY

• More efficient in enhancing heat transfer.• Able to interrupt the growth of the boundary layer

forming along the fin surface.• Louvered fin appears to be the most suitable type of fin

for automotive applications. Advantages of louvered fin [1]:

• The associated pressure drop when using louvered fin is significant.

• Adding more fins will increase the material costDisadvantages [3]:

PROBLEM STATEMENT

Heat exchanger is an important device in automotive and air conditioning applications, therefore having an effective heat exchanger will enhance the performance of the whole system.

Past studies have shown that the flow in the heat exchanger is strongly dependent on geometrical parameters.

Hence, by manipulating the geometrical parameters of the fin, we will obtain a heat exchanger with maximum heat transfer coefficient and the pressure drop is within the allowable design limit[1].

OBJECTIVE

SCOPE OF STUDY

The objective of this study is to determine the pressure drop and heat transfer characteristics of a louvered fin heat exchanger.

Simulation will be performed using ANSYS Fluent.

Validation will be conducted using the experimental result from literature.

The Reynolds number (based on louver pitch) is 200-1000.

The air inlet temperature is 27 °C which is the room temperature.

LITERATURE REVIEW

Figure 3: Flow efficiency [15]

Flow efficiency

•Flow efficiency is used to describe the percentage of the fluid flowing along the louver direction.

•100 % efficiency represents ideal louver-directed flow while 0% represents complete duct-directed flow [14].

•As Reynolds number increases, flow undergoes a transition from duct directed flow (low efficiency) to louver directed flow (high efficiency) [11].

Figure 4: Section through louver array indicating possible flow directions [15].

LITERATURE REVIEW cont’

Flow behavior

•louvers act to realign the air flow in a direction parallel to their own planes.

•the degree of alignment with the louvers was a function of Reynolds number.

•At low Reynolds number values, realignment would be slight, but at high Reynolds number it was almost complete [15].

SUMMARY OF THE LITERATURE REVIEW

The flow efficiency is strongly dependent on the geometry, especially at low Reynolds numbers.

The flow efficiency increases with the Reynolds number and louver angle, but it decreases with the fin pitch and thickness ratio.

The heat transfer for louvered fins is more appropriately described by a Reynolds number based on the louver pitch.

A louvered fin heat exchanger produced a 25% increase in heat transfer and a 110% increase in pressure drop relative to a plain fin.

Louvered-fin flow behavior is generally laminar in The ReLP range tested (50-600) with vortex shedding occurring within the louver array for ReLP > 400, depending on the model.

METHODOLOGY

Figure 5: Methodology flow chart

GEOMETRICAL DETAILS

Configuration No.

Fin pitch, Fp

[mm]

Louver pitch,

Lp [mm]

Louver angle, α

[°]

Louver thickness, t [mm]

1 1.65 0.7

25.5 0.052 1.65 1.4

3 2.02 0.7

4 2.02 1.4

5 3.25 0.7

6 3.25 1.4

Table 1: Dimensional details of computational model

Figure 6: geometrical etails of the louver

LOUVERED FIN GEOMETRY

Figure 7: Isometric and side view of the louvered heat exchanger

PRE PROCESSINGDefine the model goal

Identity the model domainDesign and create the grid

PROCESSING(FLUENT)

Set up the numerical modelCompute and monitor the solution

POST PROCESSINGExamine the result

Consider revisions to the model

CFD ANALYSIS

FLOW CHART FOR CFD SIMULATION PROCESS

PARAMETERS

No

Louver pitch = 0.7 mm Louver pitch = 1.4 mm

Reynolds number Velocity (m/s) Reynolds number Velocity (m/s)

1 200 4.51 200 2.26

2 400 9.03 400 4.51

3 600 13.54 600 6.77

4 800 18.06 800 9.03

5 1000 22.57 1000 11.28

Table 2: Parameter for Numerical Study

Here is the geometrical parameters and velocity inputs been used throughout this study.

The velocity adopted in accordance with the Reynolds number and louver pitch.

BOUNDARY CONDITION

No Name Type of boundary condition

1 Inlet Velocity inlet

2 Outlet Pressure outlet

3 Side wall Wall

4 Wall Periodic

5 Fin Wall

RESULT & DISCUSSION

MESHING SCHEME STUDIED (VALIDATION)

Pressure Drop for each velocity Percentage Difference (%) Mean Difference

Velocity (m/s) 2.264.51 6.77 9.03 11.28 2.26 4.51 6.77 9.03 11.28

No Size of Elements

No of Elements

43.50143.40 255.89 383.34 523.40 43.50 143.40 255.89 383.34 523.40 Percentage

1 0.2 45548043.76

147.26 290.39 488.89 767.27 0.6 2.69 13.48 27.53 46.59 18.18

2 0.23 48523543.5

147.24 277.78 491.43 750.33 0 2.68 8.55 28.2 43.36 16.56

3 0.24 46436644.23

151.39 296.78 505.77 762.63 1.68 5.57 15.98 31.94 45.71 20.18

4 0.25 46330044.24

148.37 293.45 502.54 775.68 1.7 3.47 14.68 31.1 44.38 19.07

5 0.26 44368042.65

146.54 291.91 497.01 728.73 1.95 2.19 14.08 29.65 39.23 17.42

6 0.27 42148042.96

146.88 287.76 460.31 732.38 1.24 2.43 12.45 20.08 39.92 15.22

7 0.28 35560242

141.34 276.81 459 730.15 3.44 1.44 8.18 19.74 39.5 14.46

8 0.29 33579043.5

147.24 277.57 491.43 750.33 0 2.68 8.47 28.2 43.36 16.54

9 0.3 33855942.89

142.59 290.19 496.72 735.39 1.4 0.56 13.4 29.58 40.5 17.09

10 0.31 39480043.03

148.41 291.95 519.93 772.08 1.08 3.49 14.09 35.63 47.51 20.36

11 0.32 39536041.47

141.83 291.68 490.19 743.52 4.67 1.09 13.99 27.87 42.06 17.94

12 0.33 39604042.47

143.54 286.49 482.34 730.13 2.37 0.1 11.96 25.83 39.5 15.95

13 0.34 39112043.5

147.24 286.98 491.43 750.33 0 2.68 12.15 28.2 43.36 17.28

14 0.35 39220041.2

137.42 293.47 491 750.4 5.29 4.17 14.69 28.08 43.37 19.12

15 0.4 37292043.5

147.24 267.8 491.43 750.33 0 2.68 4.65 28.2 43.36 15.78

CONFIGURATION 1 AFTER MESHING

Difference of pressure drop between experiment and simulation

Air velocity

(m/s)

Experimental work [1, 2]

Current work Difference (%)

Pressure drop (Pa)

Pressure drop (Pa)

2.26 43.50 42.00 3.45

4.51 143.40 141.34 1.44

6.77 255.89 276.81 8.18

9.03 383.34 459.00 19.74

11.28 523.40 730.15 39.5

Average 14.46

2 4 6 8 10 12

0

100

200

300

400

500

600

700

800

Pres

sure

dro

p (P

a)

Reynolds Number (ReLp

)

Experimenatl Numerical

Figure 9: Numerical and experimental pressure drop against Reynolds number

200 400 600 800 1000

0

200

400

600

800

1000

1200

1400

1600

1800

2000

Pres

sure

dro

p (P

a)

Reynolds number (ReLp

)

Fp=1.65 Fp=2.02 Fp=3.25

2 4 6 8 10 12

0

200

400

600

800

1000

1200

Pres

sure

dro

p (P

a)


)

Fp=1.65 Fp=2.02 Fp=3.25

Figure 11: Pressure drop against Reynolds number at louver pitch 1.4 mm

Figure 10: Pressure drop against Reynolds number at louver pitch 0.7 mm

PRESSURE DROP, ∆P

200 400 600 800 10000.25

0.30

0.35

0.40

0.45

0.50

0.55

0.60

0.65

Hea

t tra

nsef

er c

oeffi

cien

t (W

/m2.

K)


)

Fp=1.65 Fp=2.02 Fp=3.25

200 400 600 800 1000

0.15

0.20

0.25

0.30

0.35

0.40

0.45

0.50

Hea

t tra

nsef

er c

oeffi

cien

t (W

/m2.

K)


)

Fp=1.65 Fp=2.02 Fp=3.25

Figure 12: Heat transfer coefficient versus Reynolds number at louver pitch 0.7mm

Figure 13: Heat transfer coefficient versus Reynolds number at louver pitch 1.4 mm

HEAT TRANSFER COEFFICIENT, h

200 400 600 800 10000

2

4

Eule

r Num

ber (

Eu)


)

Fp=1.65 Fp=2.02 Fp=3.25

200 400 600 800 1000

2

4

6

8

10

Eule

r Num

ber (

Eu)


)

Fp=1.65 Fp=2.02 Fp=3.25

Figure 15: Euler number versus Reynolds number at louver pitch 1.4 mm

Figure 14: Euler number versus Reynolds number at louver pitch 0.7 mm

EULER NUMBER, EuHigher Euler number means that higher pressure drop occurred.

200 400 600 800 10000.007

0.008

0.009

0.010

0.011

0.012

0.013

0.014

0.015

0.016

0.017

0.018

Nus

selt

Num

ber (

Nu)


)

Fp=1.65 Fp=2.02 Fp=3.25

200 400 600 800 1000

0.008

0.010

0.012

0.014

0.016

0.018

0.020

0.022

0.024

0.026

0.028

Nuss

elt N

umbe

r (Nu

)


)

Fp=1.65 Fp=2.02 Fp=3.25

Figure 17: Nusselt number versus Reynolds number at louver pitch 1.4 mm

Figure 16: Nusselt number versus Reynolds number at louver pitch 0.7 mm

NUSSELT NUMBER, Nu

Nusselt number is a ratio of convective to conductive heat transfer across the boundary. A larger Nusselt number corresponds to more active heat convection between two boundary.

200 400 600 800 10000.00

0.05

0.10

0.15

0.20

0.25

0.30

0.35

0.40

Fric

tion

Fact

or (f

)

Stan

ton

Num

ber (

St)


)

St- Fp=1.65 f- Fp=1.65

200 400 600 800 1000

0.02

0.04

0.06

0.08

0.10

0.12

0.14

0.16

Fric

tion

Fact

or (f

)

Stan

ton

Num

ber (

St)


)

St- Fp=2.02 f- Fp=2.02

Figure 19: Stanton number and friction factor against Reynolds number for configuration 3


STANTON NUMBER, St AND FRICTION FACTOR, f

200 400 600 800 1000

0.02

0.03

0.04

0.05

0.06

0.07

Fric

tion

Fact

or (f

)

Stan

ton

Num

ber (

St)


)

St- Fp=3.25 f- Fp=3.25

200 400 600 800 10000.0

0.2

0.4

0.6

0.8

1.0

Fric

tion

Fact

or (f

)

Stan

ton

Num

ber (

St)


)

St- Fp=1.65 f- Fp=1.65

Figure 20: Stanton number and friction factor against Reynolds number for

configuration 5

Figure 21: Stanton number and friction factor against Reynolds

number for configuration 2

STANTON NUMBER, St AND FRICTION FACTOR,f

200 400 600 800 10000.00

0.05

0.10

0.15

0.20

0.25

0.30

0.35

Fric

tion

Fact

or (f

)

Stan

ton

Num

ber (

St)


)

St- Fp=2.02 f- Fp=2.02

200 400 600 800 10000.00

0.05

0.10

0.15

0.20

0.25

0.30

0.35

Fric

tion

Fact

or (f

)

Stan

ton

Num

ber (

St)


)

St-Fp=3.25 f- Fp=3.25



STANTON NUMBER,St AND FRICTION FACTOR,f

TEMPERATURE, PRESSURE CONTOURS AND STREAMLINES FOR CONFIGURATION 2

CONCLUSIONS

The major findings are summarized as follows:

Heat transfer rate increases when the fin pitch is increased. While the opposite is true in the case of pressure drop. Pressure drop and heat transfer increases when the louver pitch is decreased. The friction factor decreases with the increase in fin pitch. While the opposite is true in the case of Stanton number. Greater heat transfer values are obtained as the fin pitch is increased, due to the increased heat transfer surface area. Greater heat transfer and pressure drop values are obtained as the Reynolds number is increased. That is due to the flow tends to be louver directed flow at high Reynolds number and duct directed flow at low Reynolds number, and this two behaviors have a huge impact on heat transfer and pressure drop respectively.

for paying attention.

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