gives basic info on cfd topics

1. THE STUDY OF A CENTRIFUGAL PUMP IMPELLER BY
VARYING THE OUTLET BLADE ANGLE
C.SYAMSUNDAR 0109-1654
Supervisor: Dr. P.USHASRI
Department of Mechanical Engineering,
UCEOU.

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OUTLINE:
1. Introduction to Centrifugal Pumps
2. Literature Review
2.1. Parametric Study of a Centrifugal Pump Impeller by
Varying the Outlet Blade Angle
2.2. Performance Analysis of a Centrifugal Pump
2.3. Work done by the Impeller on Water
2.4. Effect Of Outlet Blade Angle (β2)
3. CFD Setup
4. Results
5. References

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1. INTRODUCTION

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The complexity of the flow in a turbomachine is mainly due to
the three dimensional developed structures involving;
Turbulence,
Secondary flows and
Unsteadiness etc,.
Nowadays, the design demands a detailed understanding of the
internal flows during, the design and off-design operating
conditions.
Computational fluid dynamics (CFD) have successfully
contributed for the prediction of the flow through the pump
impeller.

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The pump performance and energy consumption is affected by
various parameters.
The most critical parameters are;
1) The impeller outlet diameter,
2) The outlet blade angle and
3) The blade number.
Till now; several algorithms have been proposed and
developed, for the numerical simulation of the flow field of a
centrifugal pump impeller but the study of critical parameters
that influence the performance of a pump is not available in the
open literature.

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In this project three shrouded impellers with outlet blade angle
20 deg, 30 deg and 50 deg respectively were designed and with
the aid of CFD, the flow patterns through the pump as well as its
performance in design and off-design operation are predicted.
For each outlet blade angle of the impeller;
The flow pattern
The pressure distribution in the blade passages and
The head-capacity curves are compared.
The simulation is done by using Autocad 2007, and ANSYS
WORKBENCH.
The analysis is carried out with ANSYS CFX11.

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2. LITERATURE REVIEW
2.1) Parametric Study of a Centrifugal Pump Impeller by
Varying the Outlet Blade Angle
(Source: E.C. Bacharoudis, A.E. Filios, M.D. Mentzos and D.P. Margaris
Mechanical Engineering and Aeronautics Department,
University of Patras,
The Open Mechanical Engineering Journal, 2008, 2, 75-83)
THE IMPELLER GEOMETRY:

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Three shrouded impellers of constant width (b=20mm) with six
untwisted blades backward facing have been designed.
The blade length in the three impellers is equal.
All impellers have the same diameters in suction and pressure
side as well as the same blade’s leading edge angle (β1=140) .
There is a variation in the blade’s trailing edge angle which is
β2=200, 300 and 500.
The diameters of the impellers at the suction and pressure side
are D1=150mm and D2=280mm,
The rotational speed ,N= 925rpm,
Flow rate, Q=0.0125m3/sec
Estimated pump’s total head, H= 10m

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As the outlet blade angle increases the
performance curve becomes smoother and flatter
for the whole range of the flow rates.
There is a gain in the head which is more than
6% when the outlet blade angle increases from
20 deg to 50 deg.
The increase of the outlet blade angle causes a
significant improvement of the Hydraulic
Efficiency.

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2.2 Design and Performance Analysis of Centrifugal Pump
Source: Khin Cho Thin, Mya Mya Khaing, and Khin Maung Aye
World Academy of Science, Engineering and Technology 46 2008
CENTRIFUGAL PUMP IMPELLER DESIGN:
1) Head, H = 10 m
2) Pump speed, N = 925 r.p.m
3) Acceleration due to gravitational, g = 9.81 m/s2
4) Density of water, ρ= 1000 kg/m3
5) Specific speed, Ns = =18.39
6) Flow rate, Q = 0.0125 m3/s
4
3
H
Q
N

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7) The Inlet diameter of impeller =D1=(1.1~1.15)K0 =150 mm.
Where K0 is the constant= 4.5
8) The Outlet diameter of impeller =D2=19.2( ) =280mm.
9) The diameter of impeller eye=D0=K0 =40mm.
10) The shaft diameter at the hub section =dsh=
Torsional moment =T=9.65
11) The hub diameter is usually 1.5 to 2 times of the shaft
diameter.
Dbt = (1.5 ~ 2 ) dsh
12) Hydraulic efficiency= =83%
13) Number of Blades=Z= =6
N
Q
100
H
N
gH
2
3
N
Q
3
2
.
0 
T
N
Nmax
2
0 )
172
.
0
(log
42
.
0
1



D
r

2
)
sin(
5
.
6 2
1
1
2
1
2

 










D
D
D
D

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2.3 WORK DONE BY THE IMPELLER ON WATER

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Work done by the impeller on the water per second per unit
weight= (Vw2U2- Vw1U1)
With the assumption of no whirl component of velocity at inlet, the
work done by the impeller on the water per second per unit weight
= (Vw2U2 )
Work done by the impeller on the water per second = (Vw2U2)
g
1
g
1
g
W
From the outlet velocity triangle

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2.4 EFFECT OF OUTLET BLADE OUTLET ANGLE (β2):
From the outlet velocity triangle;
•In case of forward facing blades, and hence cot is negative and
therefore is more than .
Not desirable for pumps due to instability.
Becomes a pump for –ve Q.
•In case of radial blades, and
•In case of backward facing blades, and
Typical arrangement for pumps.
Becomes a pump for large +ve Q.

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For both radial and forward facing blades, the power is rising
monotonically as the flow rate is increased. , if the pump motor is
rated for maximum power, then it will be under utilized most of the
time, resulting in an increased cost for the extra rating.
In the case of backward facing blades, the maximum
efficiency occurs in the region of maximum power.

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BLADE ANGLE 200
3. CFD SETUP

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BLADE ANGLE 500

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MESH GENERATION:
BLADE ANGLE 200

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BLADE ANGLE 500

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BLADE ANGLE 200
BLADE ANGLE 500
Total number of nodes = 746141 Total number of nodes = 737765
Total number of tetrahedra = 3318311 Total number of tetrahedra = 3314841
Total number of pyramids = 879 Total number of pyramids = 776
Total number of prisms = 251078 Total number of prisms = 237481
Total number of elements = 3570268 Total number of elements = 3553098
MESH STATISTICS

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Boundary Conditions:
Fluid=Water
Heat Transfer Model = Isothermal
Turbulence Model = k epsilon
Turbulent Wall Functions = Scalable
Buoyancy Model = Non Buoyant
Domain Motion = Rotating

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4. RESULTS
PRESSURE CONTOURS ON MID SPAN
OUTLET BLADE ANGLE 200 OUTLET BLADE ANGLE 500

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The static pressure values increases with the increase of the
blade outlet angle. For β2=200 the maximum value is
1.614*105Pa and for β2=500 the maximum value is 1.640*105
Pa.
According to the Potential Theory, the flow at the inlet section is
displaced towards the suction side.
At constant radial position, it is observed that there a static
pressure drop form the pressure side to the suction side of the
impeller blade.
The static pressure patterns are not same in the planes between
the hub to shroud.
The minimum value of the static pressure is located at the
leading edge of the blades at the suction side.
As β2 increases further, a recirculation zone will established at
the trailing edge of the blade.

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VELOCITY CONTOURS ON MID SPAN
OUTLET BLADE ANGLE 200
OUTLET BLADE ANGLE 500

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The Maximum value of velocity decreases with the increase in β2
angle. For β2=200 the maximum value is 8.673 m/s and for
β2=500 the maximum value is 7.915m/s.
Jet wake structure is observed at the suction side of the impeller
is located at the leading edge of the blades. Here; we get high
velocity of flows and also flow separation will occur.
Velocity distribution is non-uniform in the blade passage across
the impeller width.
These non-uniformities are mainly present in the region of the
leading edges and these are minimized at the outlet of the
impeller.
The non-uniformities at the leading edges are due to the
different wall shape of the hub and shroud and also due to
inability of the fluid to adjust its path in the imposed entrance
geometry.
The absolute velocity slightly increased as the fluid moves
towards to the discharge of the impeller
Nearly uniform velocity profiles are developed at the outlet of the
impeller.

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5. NEXT PLANS:
I have planned to design and develop the third
model with β2=300 and do the
Flow pattern
Pressure distributions in the blade passage
Velocity vectors analysis.
The head-capacity curves to be drawn for all the
three models.
I planned to complete this work through the
validation with the journal and simultaneously to
investigate other crucial design parameters.

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6. REFEENCES:
1. E.C. Bcharoudis, A.E. ilios ,M.D. Mentzos and D.P. Margaris,
”Parametric Study of a Centrifugal Pump Impeller by Varying the
Outlet Blade Angle”, The open Mechanical Engineering Journal, pp75-
832008,.
2. Khin Cho Thin, Mya Mya Khaing, and Khin Maung Aye,
”Design and Performance Analysis of Centrifugal Pump”, World
Academy of Science, Engineering and Technology pp46,2008.
3. Vasilios A. Grapsas, John S. Anagnostopoulos and Dimitrios E. Papantonis,
“Hydrodynamic Design of Radial Flow Pump Impeller by Surface
Parameterization”, International Conference on Experiments / Process /
System Modeling / Simulation / Optimization, Athens, 6-9 July,2005.
4. Douglas R. Adkins and Christopher E. Brennen,“Origins of
Hydrodynamic Forces on Centrifugal Pump Impellers”, California
Institute of Technology, Pasadena, California.
5. M.H.Shojaee Fard and F.A.Boyaghchi, ”Studies on the Influence of
Various Blade Outlet Angles in a Centrifugal Pump when Handling
Viscous Fluids”, American Journal of Applied Sciences pp718-724,2007

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6. J.D. Denton, “The calculation of three-dimensional viscous flow
through multistage turbomachinery”, ASME Journal of
Turbomachinery, vol. 114, pp. 18-26, 1992.
7. M.J. Zhang, C.G. Gu, and Y.M. Miao, “Numerical study of the internal
flow field of a centrifugal impeller”, ASME Paper 94-GT-357, 1994.
8. J. Gonzalez, J. Parrondo, C. Santolaria, and E. Blanco, “Steady and
unsteady forces for a centrifugal pump with impeller to tongue
pump variation”, ASME Journal of Fluids Engineering, vol. 128, pp. 454-
462, 2006.
9. J. Gonzalez, J. Fernandez, E. Blanco, and C. Santolaria, “Numerical
simulation of the dynamic effects due to impellervolute
interaction in a centrifugal pump”, ASME Journal of Fluids Engineering,
vol. 124, pp. 348-355, 2002.
10. Centrifugal and Rotary Pumps- Fundamentals with Applications,
by Lev Nelik.
11. Centrifugal Pumps - T. Eckert, M. Hooker.
12. Know and Understand Centrifugal Pumps - Larry Bachus
and Angel Custodia
13. Centrifugal Pump Design and Application, Second
Edition,Val S. Lobaoff, Robert R.Ross.

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Thank You