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Three Dimensional CFD Analysis of

Backward Curved Radial Tipped Blade

Centrifugal Fan Designed as per Unified

Methodology with Varying Number of

Blades

N. Vibhakar #1

, S. D. Masutage#2

, S. A. Channiwala#3

#1 Mechanical Engineering Department, Sarvajanik College of Engineering & Technology, Surat – 395 001,

Gujarat, India, [email protected], Phone No. 0919428868858. #2

Mechanical Engineering Department, Sarvajanik College of Engineering & Technology, Surat – 395 001,

Gujarat, India, [email protected], Phone No. 0917567261003. #3

Mechanical Engineering Department, S. V. National Institute of Technology, Ichhchhanath, Surat - 395 007,

Gujarat, India, [email protected], Phone No.0919924166200.

Abstract The centrifugal fan designed by unified method is simulated using computational fluid

dynamics (CFD) approach. Fine mesh is generated for impeller blade zone to capture the

complex flow behaviour inside blades and mesh independency test is carried out for whole

computational domain. This three dimensional numerical analysis is steady and uses moving

reference frame (MRF) approach also known as frozen rotor method. Performance curves

are obtained under different variable inlet parameters like volume flow rate, rotational speed

and number of impeller blades. It is observed that number of blades increases, circulatory

flow reduces in blade passage and more energised flow develops. The recirculation and

separation phenomenon of fluid flow is observed inside centrifugal fan passages specifically

near tongue region. The results of this numerical analysis shows similar trend to standard

performance curves and validates unified design methodology. However, quantitative

agreement remains rather poor at off design condition.

Keywords: CFD, Centrifugal fan, Numerical simulation

#1

Corresponding Author

Nomenclature- As specified within text.

1.1 Introduction

The performance curves of a turbo machine can be obtained by theory, computation, and

by series of experiments. The experimental analysis is difficult, costly and time

consuming. To evaluate predicted performance of theoretical design, various

computational methods are available. They offer optimum design solutions without actual

fabrication or making prototypes which save time and expenditure.

Fluid passing through turbo machine or centrifugal fan in this case involves complex fluid

dynamics like flow separation, flow reversal, secondary flow, turbulent flow, boundary

layer effects, compressibility effects, heat transfer etc. This physics of flow can be

International Journal of emerging trends in engineering and development Issue 2, Vol.1 (Jan-2012) ISSN 2249-6149

Page 246

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mailto:[email protected]

mailto:[email protected]

simulated using computational fluid dynamics (CFD) softwares in discrete form of

calculations.

The most important and initial step in numerical simulation is geometry definition and grid

generation of computational domain. This process includes selection of grid types, grid

refinements and defining correct boundary conditions. Centrifugal fan involves flow

through rotating impeller and vaned or vaneless stationary diffuser. Meakhail and Park [1]

have studied the impeller-diffuser-volute casing interaction in centrifugal fan

experimentally and validated it numerically. They have used steady analysis results as an

initial parameter for unsteady analysis later on. Karnath and Sharma [2, 3] suggested

performance enhancement methods in blade passage of centrifugal fan using CFD

analysis. Trammel and Taulbee [4] had substituted blades by equivalent forces to avoid

time consumption and transient 3D simulation for their analysis of squirrel-cage blower.

Hassan, Sardar and Ghias [5] found large air flow deviation in test and CFD results at high

pressure and low air flow condition, using steady realizable k- model and concluded CFD

analysis to be more accurate for smaller mass flow rate. Steady-state interaction between

volute and impeller is studied by Fahua Gu, Engeda [6], using frozen rotor model and

explained the role of volute in reduced efficiency at off-design conditions. Power loss

occurs due to fluid drag on the reverse surface of the impeller back plate as rightly pointed

out by W.C. Osborne [7].

In this course of work, fan geometry is obtained as per unified design methodology

developed by Shah, Vibhakar and Channiwala [8]. CFD analysis carried out in this work is

to understand the volute-impeller interaction at design and off-design conditions under

varying mass flow rates, rotational speeds and number of blades. Steady, realizable k-

model with MRF approach is used to evaluate the flow behaviour inside centrifugal fan by

using ANSYS software.

2.1 Centrifugal Fan

The geometry of radial tipped backward curved centrifugal fan consisting inlet nozzle,

impeller (with 16 numbers of blades) and vaneless volute casing is shown in figure 1. The

blades are shrouded for mechanical strength and to reduce leakage losses between blades

and casing. Figure 2 shows wire meshing of centrifugal fan geometry.

Fig. 1 Wire Frame Diagram of Centrifugal Fan


47

Fig.2 Wire Frame and Volume Meshed Diagram of Centrifugal Fan

Table 1 show the specifications of the centrifugal fan and Table 2 gives the input design

point parameters for the fan under consideration.

Table 1

Specification of the simulated Centrifugal Fan

Table 2

Input Design Point Parameters for centrifugal fan

Parameters Values

Discharge Q 0.5

m3/s

Differential pressure p 981.2

Pa

Impeller speed N 2800

rpm

Suction pressure at nozzle inlet -196.4

Parameters Dimensions

Inlet nozzle Inlet diameter dn in 174 mm

Inlet nozzle Outlet diameter dn

out

155 mm

Impeller Inlet diameter d1 155 mm

Impeller Outlet diameter d2 303 mm

Impeller Width b2 56 mm

Outlet section area h*w 346 *111

mm2

Blade Inlet blade angle 1 42.270

Blade Outlet blade angle 2 900

Blade Thickness tblade 2 mm

Volute Casing Inlet diameter d3 303 mm

Volute Casing Outlet diameter

d4

591 mm


Page 248

psuction Pa

Fan inlet STP

3.1 Numerical Analysis

Numerical analysis software Gambit-FLUENT is used for this course of work. The Gambit

is a pre-processor, used for geometry and grid generation purpose. Tetrahedral elements

are used in all three zones of impeller, inlet nozzle and volute casing. After successful grid

generation, its independency test is carried out before proceeding for final simulation. Test

grid independency results are shown in fig. 3 graphically and numerically in Table 3.

Fig. 3 Mesh Independency Test

Table 3

mesh independency values

Mesh No. of mesh

elements

Pressure

ratio

a 219210 1.00788

b 419633 1.00825

c 869068 1.00824

d 1934222 1.008237

Figure 3 clearly indicates steady trend of pressure ratio after mesh b which is having

419633 numbers of tetrahedral elements and at pressure ratio 1.00825. Hence mesh b is

selected for further simulation work to reduce computational time, without affecting

optimum performance.

4.1 Simulation Parameters

Three dimensional simulation is carried out using ‘Reynolds-averaged Navier-Stokes’

equations (RANS) and ‘Realizable k- model’. This selected model gives superior

performance for flows involving rotation, separation and recirculation. The ‘standard’ wall

function is used to resolve the wall flows. The ‘SIMPLE’ algorithm is used for coupling

pressure and velocity.


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The discretization of pressure is done by using PRESTO scheme to obtain good results in

vortex flow field. The momentum, turbulent kinetic energy, turbulent dissipation rate and

energy are discretized using second order upwind method. The solution convergence

criterion is kept at maximum residuals of less than 10-5

.

The inlet nozzle and volute casing are stationary zones and impeller is a moving zone.

Being steady flow and very low pressure ratio in this case, moving reference frame (MRF)

approach is used to impose rotational field to impeller zone of centrifugal fan. The blades

are moving with a same rotational velocity as impeller zone by giving moving wall

condition with zero relative rotational speed to adjacent cell zone. Shadow wall method is

used as interior surface to create continuous flow path between moving and stationary

zones.

The mass flow rate is used as the inlet boundary condition at nozzle inlet and zero gradient

outflow condition is used at the casing outlet, assuming fully developed flow conditions.

At the inlet boundary condition, 5% turbulent intensity and 0.5 turbulent length scale is

applied, which is the cube root of domain volume and is used for turbulent specification

method. No-slip boundary condition is used for all the walls. The discharge is varied by

varying the input boundary condition i.e. mass flow rate at the nozzle inlet at the constant

rotational speed of impeller.

5.1 Results

The post processing work of this fan is done using FLUENT software. Inlet quantities are

calculated at the nozzle inlet plane and outlet quantities are calculated at the casing outlet

plane. Performance graphs are plotted in figure 3 by keeping rotational speed of impeller

constant and varying;

1. Number of impeller blades as 12, 16 and 24

2. Volume flow rate from 0.1 to 0.5 m3/sec

It is revealed from figure 3 (a) that static pressure decreases when discharge increases and

progresses to higher values with respect to increase in number of blades. This is because

pressure loss occurs within the blade passage due to flow separation. As number of blades

increases the flow within blades get more kinetic energy compared to fewer numbers of

blades. This extra kinetic energy increases the attachment of flow to the blade surface and

reduces flow separation effect. Design point change in static pressure is 981 Pa for 16

numbers of blades [8] and simulated result shows 834.7 Pa for similar conditions. This is

15% deviation between experiment and numerical results. Figure 3 (b) and 3 (c) shows

that when flow coefficient increases, pressure rise coefficient decreases and power

coefficient increases correspondingly. For 16 numbers of blades at design point discharge

0.5 m3/s, pressure rise coefficient reaches to 2.92 and power coefficient reaches at

maximum to 3.46 and subsequently decreases. Figure 3 (d) shows maximum total

efficiency 45.89%, which occurs at 0.25 flow coefficient.


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(a)

(b)

(c)

0

200

400

600

800

1000

1200

1400

1600

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7St

atic

Pre

ssu

re (

Pa)

Volume Flow Rate (m3/s)

z=12

z=16

z=24

Design point Z=16

0

1

2

3

4

5

6

7

0.0 0.1 0.2 0.3 0.4 0.5

Pre

ssu

re R

ise

Co

eff

icie

nt

Flow Coefficient

z=12z=16z=24Design point Z=16

0.00.51.01.52.02.53.03.54.04.55.0

0.0 0.1 0.2 0.3 0.4 0.5

Po

we

r C

oe

ffic

ien

t

Flow Coefficient

z=12

z=16

z=24

Design point Z=16


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(d)

Fig 4. (a) Static pressure Vs. volume flow rate (b) Pressure rise coefficient Vs. flow

coefficient (c) Power coefficient Vs. flow coefficient (d) Efficiency Vs. flow coefficient

for constant rotational velocity N=2800 rpm with 12, 16 and 24 numbers of blades

with design point.

The figure 5 shows the effect of varying rotational speed while 16 numbers of blades kept

constant. As per figure 4(a), the static pressure increases till the rotational speed

approaches design point speed. This is for the reason that, impeller blade profile is

designed for specific set of input parameters. Its performance deviates at off design

conditions and turbulence losses increases due to rotation of flow within blade passage.

Volume flow rate and rise in static pressure difference across the fan directly proportionate

to rotational speed till design conditions. Afterwards they remain constant for same

geometry. Dimensionless coefficients such as pressure rise coefficient, power coefficient

and efficiency are almost unaffected near design flow coefficient and has different values

at lower and higher flow coefficients (fig. 4 (b), (c) and (d)).

(a)

0.0

0.1

0.2

0.3

0.4

0.5

0.6

0.0 0.1 0.2 0.3 0.4 0.5Ef

fici

en

cy

Flow Coefficient

z=12z=16z=24Design point Z=16

0

250

500

750

1000

1250

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7

Stat

ic P

ress

ure

(P

a)

Volume Flow Rate (m 3/s)

N = 2500

N = 2650

N = 2800


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(b)

(c)

(d)

Fig 5. (a) Static pressure Vs. volume flow rate (b) Pressure rise coefficient Vs. flow

coefficient (c) Power coefficient Vs. flow coefficient (d) Efficiency Vs. flow coefficient

for 2500,2650 and 2800 rpm keeping no. of blades Z=16 constant

Efficient energy transfer in a centrifugal fan depends upon proper blade profile, gradual

change in area of volute casing and smooth surface finish. For such energy transfer Flow

lines must be parallel to each other and should generate streamlined flow within guided

three dimensional passages.

Figure 6 shows the streamlines in simulated centrifugal fan at mid-plane surface. The

streamlines remains parallel except near tongue region. Tongue obstruction creates flow

0

1

2

3

4

5

6

7

0.0 0.2 0.4 0.6P

ress

ure

Ris

e C

oe

ffic

ien

tFlow Coefficient

N = 2500

N = 2650

N = 2800

0

1

2

3

4

5

0.0 0.2 0.4 0.6

Po

we

r C

oe

ffic

ien

t

Flow Coefficient

N=2500

N=2650

N=2800

0.0

0.1

0.2

0.3

0.4

0.5

0.0 0.2 0.4 0.6

Effi

cie

ncy

Flow Coefficient

N=2500

N=2650

N=2800


Page 253

deviation and change in local velocity and also generates recirculation of flow. Flow

rotation within blade passage is also observed.

Fig 6 Streamlines at mid-plane iso-surface in simulated centrifugal fan for Z=16

The figure 7 shows velocity contours at mid-plane of an impeller varying blades in step of

12, 16 and 24 keeping rotation speed constant as 2800 rpm and at constant discharge

Q=0.5 m3/s.

As numbers of blade increases, fluid is efficiently guided within blade passage and vortex

region seen near tongue region also decreased. The large velocity region seen near

impeller zone suggests higher energy conversion in impeller and as per Austin Church [9],

it is advantageous to have large part of the total head developed from impeller.

Z=12


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Z=16

Z=24

Fig 7 Velocity contours at mid plane for Q=0.5 m3/s, N=2800 rpm and Z=12, Z=16

and Z=24

Here casing is designed by four arcs method which generates large exposure to flow at

exit. This tends to permit reversal of flow at fan outlet. As a result, large velocity

difference regions are observed near volute casing exit in velocity contours. Basic

intention of large exit area is to permit free vortex flow within entire casing volume. This

happens only at design point condition but at off design conditions it creates tendency of

retardation of flow and results in eddy formation within flow.

6.1 Conclusions

1.The results of numerical analysis, clearly follows the standard performance curves of a

centrifugal fan. Present numerical analysis is closer to design point parameters, for

centrifugal fan under study.

2.The mean pressure distribution around the volute casing is not uniform and jet and

wakes are observed in the vicinity of tongue region. Since the flow is considered

incompressible, pressure fluctuations at impeller outlet region is affecting flow around

impeller zone.


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3.The flow phenomenon of recirculation near tongue region is confirmed by numerical

analysis.

4.The exact amount of circulatory flow depends upon the shape of the blade passage. For a

given impeller more vanes make the passage narrower giving greater guidance to the

fluid and reducing the circulatory flow effect. Hence all the quantities varying with flow

coefficient are increased as the number of blades increases.

5.The streamlines clearly shows the rotating effect of blades on flow, within and outside of

impeller zone, which is imposed using MRF approach.

6.The numerical analysis shows that, for efficient energy transfer i.e. to achieve optimum

performance in centrifugal fan designed by unified methodology number of blades

should increase.

Acknowledgments

We are sincerely thankful to our institutes and management for permitting us to use

infrastructure, library and computational resources.

References

[1] Meakhail T. and Park S. O., A study of impeller-diffuser-volute casing interaction in a

centrifugal fan, Journal of Turbomachinery, Vol. 127, pp. 84-90, January 2005.

[2] Karanth K. V. and Sharma N. Y., Numerical analysis on the effect of varying number of

diffuser vanes on impeller-diffuser flow interaction in a centrifugal fan, World Journal

of Modeling and Simulation, Vol. 5, No. 1, pp. 63-71, 2009.

[3] Karanth K. V. and Sharma N. Y., Numerical Analysis of a Centrifugal Fan for

Improved Performance using Splitter Vanes, World Academy of Science, Engineering

and Technology, pp. 453-459, 2009.

[4] Tremmel M. and Taublee D. B., Calculation of the time averaged flow in squirrel-cage

blowers by substituting blades with equivalent forces, Journal of Turbomachinery, Vol.

130, pp. 031001-1-- 031001-11, July 2008.

[5] Hassan M. B., Sardar A. and Ghias R., CFD simulations of an automotive HVAC

blower operating under stable and unstable flow conditions, SAE International 2010.

[6] Fahua Gu and Abraham Engeda, A numerical investigation on the volute/impeller

steady-state interaction due to circumferential distortion, Proceedings of ASME

TURBO EXPO 2001, New Orleans, Louisiana, USA. 2001-GT-0328, June 4-7, 2001.

[7] W. C. Osborne, Fans, 2nd

edition, PERGAMON press, 1977.

[8] Shah K. H., Vibhakar N. N. and Channiwala S. A. Unified design and comparative

performance evaluation of forward and backward curved radial tipped centrifugal fan,

Proceedings of the International Conference on Mechanical Engineering 2003

(ICME2003), Dhaka, Bangladesh, 26- 28 December 2003.

[9] A. H. Church, Centrifugal Pumps & Blowers, Robert E. Krieger publishing company,

1972.

[10] FLUENT user guide 6.3, ANSYS Inc.


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