Cfd analysis of mixed flow impeller

IJRET: International Journal of Research in Engineering and Technology eISSN: 2319-1163 | pISSN: 2321-7308
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Volume: 03 Special Issue: 03 | May-2014 | NCRIET-2014, Available @ http://www.ijret.org 601
CFD ANALYSIS OF MIXED FLOW IMPELLER
Neelambika1
, Veerbhadrappa2
1
Asst. Prof. Mechanical Engg Dept. Govt. Polytechnic Kalagi
2
Asst. Prof. Mech. Engg Dept. BKIT Bhalki-585328
Abstract
Now a day’s Computational Fluid dynamics (CFD) is playing a vital role in all the fields. To improve the efficiency of mixed flow
pump, Computational Fluid Dynamics analysis is one of the advanced tools used in the industry. A detailed CFD analysis was done to
predict the flow pattern inside the impeller which is an active pump component. From the results of CFD analysis, the velocity and
pressure in the outlet of the impeller is predicted. CFD analyses are done using Ansys software. These outlet flow conditions are used
to calculate the efficiency of the impeller. The calculated value of efficiency from the empirical relations is 55%. The optimum inlet
and outlet vane angles are calculated for the existing impeller by using the empirical relations. The CAD models of the mixed flow
impeller with optimum inlet and outlet angles are modeled using CAD modeling software Catia-V5R19. To find the relationship
between the vane angles and the impeller performance the optimum vane angle is achieved step by step. Three CAD models are
modeled with the vane angles between existing and optimum values. These models are analyzed individually to find the performance
of the impeller. In the first case, outlet angle is increased by 5°. From the outlet flow conditions, obtained from the CFD analysis, it is
evident that the reduced outlet recirculation and flow separation cause the improved efficiency. By changing the outlet angle the
efficiency of the impeller is improved to 59%. In the second case inlet angle is decreased by 10%. The efficiency of the impeller in
this case is 61%. From this analysis it is understood that the changes in the inlet vane angle did not change the efficiency of the
impeller as much as the changes in outlet angle. From the CFD analysis the efficiency of the impeller with optimum vane angles is
calculated as 65%. Thus, efficiency of the mixed flow impeller is improved by 18.18% by changing the inlet and outlet vane angles.
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1. INTRODUCTION
A wide variety of centrifugal pump types have been
constructed and used in many different applications in
industry and other technical sectors. However, their design
and performance prediction process is still a difficult task,
mainly due to the great number of free geometric parameters,
the effect of which cannot be directly evaluated. The
significant cost and time of the trial-and-error process by
constructing and testing physical prototypes reduces the profit
margins of the pump manufacturers. For this reason CFD
analysis is currently being used in the design and construction
stage of various pump types.
The experimental way of pump test can give the actual value
of head developed, power rating and efficiency. But the
internal flow conditions cannot be predicted by the
experimental results. From the CFD analysis software and
advanced post processing tools the complex flow inside the
impeller can be analyzed. The complex flow characteristics
like inlet pre-swirl, flow separation and outlet recirculation
cannot be visualized by the experimental way of pump test.
But in the case of CFD analysis the above flow characters can
be visualized clearly. Moreover design modification can be
done easily and thus CFD analysis reduces the product
development time and cost. Mixed flow pumps are widely
used for water transportation or as cooling water pumps in
power stations. Their operating range spans from full-load
down to close to the shut-off head. In order to develop a
reliable machine for this highly demanding operation, the
behavior of the flow in the entire pump has to be predicted by
a reliable computational method.
2. PUMP CHARACTERISTICS
2.1 Cavitation
Cavitation is a hydraulic phenomenon in which vapor bubbles
form and suddenly collapse (implode) as they move through a
pump impeller. Implosions occur on each of the vanes of the
impeller causing excessive noise. The hydraulic effect on the
pump is a significant reduction in performance. The
mechanical effects can include shock waves and vibration
which may result in damage to the impeller vanes, bearings,
and seals.
2.2 Vortexing
A vortex can occur at the impeller and may extend to the
surface of the liquid. If this occurs, air will be sucked into the
pump. The effects can be similar to cavitation: reduced
hydraulic efficiency and increased wear on the pump. Dicmas
identifies three stages of vortexing3:
2.3 Net Positive Suction Head Required
Net positive suction head required (NPSHR) is the head above
vapor pressure head required to ensure that cavitation does not
occur at the impeller. The value is specific to each pump inlet

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design. It is independent of the suction piping system. The
NPSHR is determined by bench scale tests of geometrically
similar pumps operating at a constant speed and discharge
with varying suction heads. The development of cavitation is
usually indicated by approximately a 3% drop in the head
developed at the suction inlet valve. The NPSHR is specified
by the manufacturer.
2.4 Specific Speed
Specific speed, Ns, is often quoted as a dimensionless design
index for pump impellers. However, strictly speaking, it has
dimensions that are in terms of:
[L0.75
T-0.5
] x revolutions per minute
Where L represents units of length, and T represents units of
time.
Specific speed is used to identify an upper limit on the shaft
speed for any particular combination of total head, flow, and
suction conditions. The specific speed is determined using the
following equation:
75.0
5.0
H
NQ
CNs  (2.1)
N = pump speed, rpm
Q = pump capacity at Best Efficiency Point
(BEP), m3
/hr (gpm)
H = pump head at BEP, m (ft)
C = unit factor = 0.861 * (1)
* For consistency with US manufacturers’ pump data, a unit
factor for metric units is given to result in a consistent value
that is really in English units. The designer should verify the
derivation and manufacturer’s use of specific speed.
By applying pump affinity laws, a pump designer can predict
the change in performance of a pump in response to changing
speed. Discussion of affinity laws is beyond the scope of this
manual. For a discussion, refer to a text such as Pump
Performance Characteristics and Applications.4
2.5 Suction Specific Speed
Suction Specific Speed, Nss, is a dimensionless index that is
both representative of the geometry of the suction side of the
impeller, and is used to determine the form and proportions of
the impeller. The index can be used to select a pump
maximum speed that will likely yield the smallest size of
pump for the design conditions. It is determined using the
following equation:
75.0
5.0
NPSHR
NQ
CNss  (8.2)
NPSHR=net positive suction head required, m (ft)
Nss = suction specific speed, rpm
N = pump speed, rpm
Q = pump capacity at Best Efficiency Point (BEP), m3
/hr
(gpm)
C = unit factor = 0.861 * (1)
Selection of Impeller Type
Table 1.1 indicates the ranges of specific speed over various
impellers that are recommended for use.
Table 1.1. Specific speed range for impellers
Impeller Type
Specific Speed
Upper Lower
Axial Flow 20,000 8,500
Mixed Flow (open) 10,000 5,000
Mixed Flow (closed) 6,200 4,000
Centrifugal 4,100 2,500
2.6 Pump Performance
Pump performance is measured as the variation in pumping
capacity with respect to total dynamic head. Pump
performance is a function of the following pump
characteristics: pump type, pump size, impeller size, and
speed.
Pump performance can be shown either as a single line curve
for one impeller diameter or as multiple curves for the
performance of several impeller diameters in one casing.
Figure 1.3 shows a typical single pump performance curve.
The point at which maximum efficiency is achieved is
represented by the best efficiency point, BEP, as shown in
Figure 1.4.
In this work, the mixed flow pump detailed geometric feature
of the impeller is studied and parameterization of impeller
geometry is done. Parameterization is done by reducing
number of controlling geometric variables, facilitating the
investigation of their individual or combined effects on the
flow and the impeller performance

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Fig.2.1: Typical pump performance curve
Fig.2.2: Simplified manufacturer’s pump performance curves
The specification of the pump is given in the Table 2.2.
Typical submersible pump consists of impeller, diffuser, riser,
discharge adapter, intermediate bowl and shaft is shown in
Fig.1.7The performance of a pump depends on active pump
components. The active pump components of a mixed flow
pump are mixed flow impeller and guide vanes. These parts
are important while analyzing a pump because mechanical
energy (rotation) is transferred into fluid energy (pressure,
velocity) by the active pump components. The geometry of a
mixed flow pump impeller is highly complex in nature. The
mixed flow geometry consists of several features of which the
following geometric features are important as they have direct
effect on overall pump performance.
Table 2.2: Specification of the pump
Bore size 152.4 mm
Capacity 40 m3/hr
Number of stages 6
Head 28 m per stage
Flow type Mixed flow
Fig.2.3: Mixed flow pump
3. CFD ANALYSIS OF MIXED FLOW PUMP
A thorough knowledge of all significant inviscid effects (blade
blockage, flow turning, finite hub/tip ratio, etc.) and viscous
effects (boundary-layer growth, energy dissipation, etc.) is
essential in the accurate prediction of the flow in all turbo
machinery. Relevant to this, the availability of computer with
large storage capacities and fast computation times greatly
enhance the possibility of numerically solving the complete
equations of motion. It is widely accepted that the
incompressible time-averaged Navier-Stokes equation
Pump Performance
0
5
10
15
20
25
0 500 1000 1500
Pump Rate (m3
/hr)
TotalDynamicHead
(m)
0
5
10
15
20
25
30
0 200 400 600 800 1000 1200 1400 1600
Flow (m3
/hr)
Head(m)
450
350
375
400
325
425
82
80
78
76
82
80
78
76
Best Efficiency Point
Impeller diameter (mm)
Performance curve for impeller size
Efficiency curve

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together with the continuity equation comprise a closed set of
equations, and which should be supplemented by appropriate
boundary and initial conditions to form a complete description
of the fluid velocity u(x,t) r and pressure field in flow pumps.
3.1 Geometric Modeling Pump
Table.3.1: Blade Impeller Parameter
Impeller inlet (Di) 75 mm
Impeller outlet (Do) 105 mm
Blades number 6
Inlet angle(α) 550
Outlet angle (β) 750
Blade thickness (t) 1.25 mm
Blade inlet height (Li) 21 mm
Blade outlet height (Lo) 16 m
Fig 3.1: Existing Impeller vane dimensions
In RANS simulations, the choice of mesh type is of critical
importance. In this study, the widely used structured body-
fitted curvilinear meshes are chosen rather than unstructured
tetrahedral meshes. Body-fitted structured meshes are well
suited for viscous flow because they can be easily compressed
near all solid surfaces. Using a multi block approach, they are
also convenient for discretizing the flow passages in
turbomachinery flows with rather straightforward geometries,
which includes blade tip clearances and relative motion.
The grid used for the present study is shown in Figure 3.6 and
3.7 with the 3D view and blade-to-blade view. The grid for the
used for this study is impeller 0.3 millions.
Fig.3.2: Blade to Blade View of Mixed Impeller Pump
Fig. 3.3: CFD Meshing of Mixed Impeller Pump
4. RESULTS AND DISCUSSIONS
4.1 Inlet angle 750
and outlet Angle 550
Flow distribution in impeller at different span
The pressure increases gradually along stream wise direction
within impeller passage and has higher pressure in pressure
side than suction side of the impeller blade. However, the
pressure developed inside the impeller is not so uniform. The
isobar lines are not all perpendicular to the pressure side of the
blade inside the impeller passage; this indicated that there
could be a flow separation because of the pressure gradient
effect. The fig 4.1 shows the pressure distribution within the
impeller at three different 50 percent spans

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Fig. 4.1: Pressure distribution at 50% span
Velocity also increases gradually along streamwise direction
within the impeller passage.As the flow enters the impeller
eye, it is diverted to the blade-to-blade passage. The flow at
the entrance is not shockless because of the unsteady flow
entering the impellerpassage. The separation of flow can be
seen at the blade leading edge. Since, the flow atthe inlet of
impeller is not tangential to the blade, the flow along the blade
is not uniformand hence the separation of flow takes place
along the surface of blade. Here it can beseen that flow
separation is taking place on both side of the blade, ie,
pressure and suction side as shown in fig 4.2
Fig. 4.2: Velocity distribution at 50% span
4.1.2 Pressure and Velocity between Impeller Blades
The distribution of total pressure between the blades of the
impeller is shown in the fig [4.3]. The lowest total pressure
appears at the inlet of the impeller suction side. This is the
position where cavitation often appears in the centrifugal
pump. The highest total pressure occurs at the outlet of
impeller, where the kinetic energy of flow reaches maximum.
Fig. 4.3: Contour of Pt (Total Pressure) at 50% Span
The fig [4.4] shows the static pressure distribution at the span
of 50 between the blades of the impeller. It is observed that the
static pressure inside the impeller blades is asymmetry
distributed. The minimum static pressure area appears at the
back of the impeller blade ie suction side, at the inlet.
Fig. 4.4: Contour of Ps (Static Pressure) at 50% Span
The fig [4.5] shows the relative velocity distribution at 50 span
between two blades of the impeller. The rising flow speed of
the fluid from inlet to outlet reaches maximum at the impeller
blade outlet. From the pressure distribution plot, pressure
gradient changes gradually from inlet radius to outlet radius
for each impeller. In order to transmit power to the liquid,
pressure on the leading or pressure side of the vane should be
higher than pressure on the suction side of the vane.

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Fig. 4.5: Contour of Width of the element at 50% Span
4.1. Fig. 4.6: Contour of Width element at Blade trailing edge
Fig [4.13] shows the velocity streamlines at the trailing edge
which shows that how the flow leaves the blades of the
impeller.
Fig.4.7: Velocity Streamlines at Blade trailing edge
4.1.3 Blade Loading at Pressure and Suction Side
The blade loading of pressure and suction side are drawn at
three different locations on the blade at the span of 20, 50 and
80 from hub towards the shroud. The pressure loading on the
impeller blade is shown in fig [4.14]. Pressure load on the
impeller blade is plot along the stream wise direction. The
pressure difference on the pressure and suction sides of the
blade suggests that the flow inside the impeller experiences
the shearing effects due to the pressure difference on blade-to-
blade passage wall.
Fig.4.8: Blade loading at 20, 50 and 80 span
4.1.4 Stream Wise Plots
The effects of flow rate on the pressure inside the impeller
passage along the stream wise direction are shown in Fig
[4.15], which shows the flow rate effects on the pressure
distribution on the impeller blade. The pressure for the
minimum flow rate is found maximum at the outlet of the
impeller and vice versa. Similarly relative velocity is shown in
4.16.
Fig. 4.9: Stream Wise plots Pressure

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Fig. 4.10: Stream Wise plots Relative velocity
5. CONCLUSIONS
Based on the CFD analysis of the mixed flow impeller, the
following conclusions are made. The best efficiency point of
the pump is found to be 11 lps. The existing impeller, the head
and efficiency are found out to be 19.24 m and 55%
respectively. The impeller 1, the percentage increase in the
head and efficiency are 3.22% and 7.27% respectively. The
impeller 2, the percentage increase in the head and efficiency
are 10.29% and 10.91% respectively. The impeller 3, the
percentage increase in the head and efficiency are 13.66% and
18.18% respectively. Based on the above, it is concluded that
impeller 3 gives better performance. Thus CFD analysis is an
effective tool to analyze quickly and inexpensively the effect
of design and operating parameter of pump. By properly
designing pump impeller, the efficiency of pump can be
improved.
REFERENCES
[1] Lazarkiewicz, S. and Troskolanski, A. T. (1965)
Impeller Pumps, Pergamon Press Ltd., Oxford.
[2] Goto, A. (1995) “Numerical and Experimental Study of
3-D Flow Fields within a Diffuser Pump Stage at Off-
Design Condition,” ASME Fluids 95, FED-Vol.227,
pp. 1-9.
[3] 11. CFX User Manual Version 5.6 (2003) CFX
Ltd., United Kingdom. 12. Muggli, F. A., Holbein, P.
and Dupont, P., (2002) “CFD Calculation of aMixed
Flow Pump Characteristics from Shutoff to Maximum
Flow” ASME Journal of Fluids Engineering., Vol. 124,
pp. 798-802. 94
[4] 13. CFdesign User’s Guide Version 7.0 (2004)
Blue Ridge Numerics, Inc., Charlottesville.
[5] 14. Hirsch, Ch. (1994) “ CFD Methodology and
Validation for Turbomachinery Flows”,
Turbomachinery Design Using CFD-AGARD Lecture
Series 195,Specialised Printing Services, Essex, pp 4-7.
[6] 15. Majidi, K., (2005) “ Numerical Study of
Unsteady Flow in a Centrifugal Pump” ASME Journal
of Turbomachinery., Vol. 127, pp. 363-371.

Cfd analysis of mixed flow impeller

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Cfd analysis of mixed flow impeller