Development of low flow high head pump

i
DEVELOPMENT OF LOW FLOW/HIGH HEAD
PUMP
A thesis submitted
in partial fulfilment of the requirements
for the degree of
Master of Technology
by
Don Poul Jose
(R123216003)
to the
Department of Mechanical Engineering
University of Petroleum and Energy Studies
Dehradun, India
December 2017

iii
Department of Mechanical Engineering
University of Petroleum and Energy Studies
Certificate
It is certified that the work contained in the thesis/project titled Development
of Low Flow/High Head Pump by Don Poul Jose has been carried out under
my supervision and that this work has not been submitted elsewhere for a
degree.
Dr. Ashish Karn
(Asst. Prof. Sel. Grade)
Department of Mechanical Engineering,
College of Engineering, U.P.E.S.
Dehradun, Uttarakhand
India-248007
Mr. Shyam Pandey
(Associate Professor – S.S)
Department of Mechanical Engineering,
College of Engineering, U.P.E.S.
Dehradun, Uttarakhand
India- 248007
December 2017

v
Dedicated to
My Parents,
Well-wishers,
and
My Faculty.

vii
Abstract
This paper deals with the development of Low Flow/High Head Pump. In this paper,
centrifugal pump is studied by using a single-stage end suction centrifugal pump. Two
main components of a centrifugal pump are the impeller and the casing. The impeller is
a rotating component and the casing is a stationary component. In centrifugal pump, water
enters axially through the impeller eyes and water exits radially. The pump casing is to
guide the liquid to the impeller, converts into pressure the high velocity kinetic energy of
the flow from the impeller discharge and leads liquid away of the energy having imparted
to the liquid comes from the volute casing. The design of Low Flow/High Head
centrifugal pump are chosen because of its need and importance in chemical industry.
Low flow/High Head pump demands was born when the demand of more efficient and
improved chemical reactions started requiring higher process pressure and temperatures.
It is the most useful mechanical rotor-dynamic machine in fluid works which widely used
in pump Seawater booster for critical services, Light hydrocarbon boosting,
Petrochemical processing, Light Vacuum Gas Oil(LVGO) and Heavy Vacuum Gas
Oil(HVGO), Heavy-duty chemical processing etc. A brief history of these pump types
are taken for study and comparing the data such as type of impellers, efficiency, head
coefficient, head-capacity curves and relative cost of the current designs and study the
current needs. Evaluating the needs and study of advantages and disadvantages from the
design point of view to develop an effective plan of execution. With summing up all data
collected and the current market requirement data, mathematically design a pump with
property of Low flow/High Head. A very advanced software pack including Creo and
CFturbo 10 is used for designing the impeller with the help of mathematically calculated
values and inputs.

ix
Acknowledgements
It’s my pleasure and duty to express my indebtedness to Mr. Ramakrishnan
Ramamoorthy, CEO & Managing Director at Egger Pumps India Pvt Ltd and Mr
Nithyanandhan Krishnaswamy, COO for allowing me to undergo my project work at
Egger Pumps India.
I take this opportunity to express my sincere thanks to my supervisor Dr. Ashish
Karn (Asst. Prof. Sel. Grade), Mechanical Engineering, UPES Dehradun, for his
guidance and support.
I also express my gratitude to external guide Mr Vimal R, Senior Design
Engineer at Egger Pumps India Pvt Ltd. It was a privilege working under him.
My sincere thanks to Mr. Ashok Kumar, Assistant Professor, Course
Coordinator of M-Tech Rotating Equipment Engineering, UPES Dehradun, for
providing me this Internship in this organisation, without him today was a distant dream.
I also thank Mr. Shyam Pandey, Head of the Department, M-Tech Rotating
Equipment, UPES Dehradun, for permitting me to do this project and for his constant
encouragement and support throughout the project.
I am very much thankful to the University of Petroleum and Energy Studies
(UPES), Dehradun, for providing me this opportunity of pursuing M-Tech - Rotating
Equipment Engineering in a peaceful environment with ample resources.
I also express my gratitude to my Faculties of UPES Dehradun, and Employees
of Egger Pumps India Pvt Ltd for their constant guidance.
In the end, I would like to acknowledge My Parents, Family Members. Without
their support, this work would not have been possible.
DON POUL JOSE

xi
Contents
Certificate
Abstract
Acknowledgement
Contents
List of Figures
Nomenclature
iii
vii
ix
xi
xiv
xviii
1. Introduction 1
1.1. Motivation 2
1.2. Solution 3
2. Literature Review 4
3. Pump 8
3.1. Specific Speed 9
3.1.1. Deriving the formula for Specific speed 11
3.1.2. Specific Speed Basics 12
3.1.3. Specific Speed and the Pump Curve 13
3.2. Impellers 14
3.2.1. Vane Design 15
3.2.2. Impeller Blade Types 16
3.2.3. Impeller Design 18
3.2.4. Velocity Triangles 21
3.3. Casing 23
3.3.1. Single Volute Casing for Process Pump 26
3.3.2. Results and Conclusions 28
4. Mathematical Design of Low Flow/High Head Pump – Calculations 30
4.1. Duty Condition
4.2. Calculations
4.2.1. Pump Rotational Speed – n (rpm) 30
4.2.2. Specific Speed-Optimum Geometry Versus Specific Speed 33

xii
4.2.3. Pump power output or Hydraulic power 36
4.2.4. Torque 37
4.2.5. Brake Horse Power(BHP) 38
4.2.6. Shaft Torque 39
4.2.7. Shaft Diameter 40
4.2.8. Inputs to CFturbo 10 to generate the inlet meridional velocity 41
4.2.9. The inlet meridional velocity 42
4.2.10. The inlet diameter (d0) 43
4.2.11. Impeller Width Calculation (Inlet) 43
4.2.12. Blade inlet angle (β1) 43
4.2.13. Fluid velocity relative to the impellers 44
4.2.14. Inlet velocity triangle 44
4.2.15. Inputs to CFturbo 10 to generate the outlet meridional velocity 45
4.2.16. Outlet meridional velocity 45
4.2.17. Blade outlet angle (β2) 46
4.2.18. Number of blades, Z 47
4.2.19. Impeller Width Calculation 47
4.2.20. Calculated Values 48
4.3. Designing of Impeller 51
4.4. Developed Impeller 55
4.5. Modelling using CFturbo 56
4.6. Predicted performance curve 58
5. Conclusion 60
6. References 61

xiv
List of Figures
S.No Figure Caption Page
3.1 The pump classification according to the motion of work 9
3.2 Comparison of pump profiles with the values of specific speed 10
3.3 Figure illustrates the general appearance of impellers of
various specific speed designs, with inlet being at the bottom.
12
3.4 HQ and BHP curves for various specific speeds 14
3.5 Open , Semi-Open and Closed Impeller designs 15
3.6 Forward , Radial and Backward facing impeller vanes 17
3.7 Water Vapour contour at 20 lps 17
3.8 Head-Discharge characteristics of different blades 18
3.9 Velocity diagram in an impeller stage 21
3.10 Velocity triangles 21
3.11 Single-volute casing & Double-volute casing 23
3.12 Radial force for single and double-volute casing 24
3.13 Casing with and without Diffuser 26
3.14 The volute with radial diffuser 27
3.15 The volute with tangential diffuser 27
3.16 Head vs Discharge 28
3.17 Head vs Efficiency 28
3.18 The pressure fluctuation intensity in the diffuser of volute 30
4.1 Representation of magnetic poles in a pump motor 31
4.2 Synchronous and approximate full load speed of AC electrical
motors
32

xv
4.3 Optimum geometry as a function of BEP specific speed (for
single stage rotors).
33
4.4 Approximate relative impeller shapes and efficiencies as
related to specific speed
39
4.5 Global Setup – Pump Inputs 41
4.6 Input of Hub Diameter to get Input Meridional velocity 42
4.7 Inlet Velocity triangle – With calculated Values 44
4.8 Input of Impeller Diameter and Impeller outlet width 45
4.9 Outlet Velocity triangle – With calculated Values 46
4.10 Velocity triangles generated by CFturbo 50
4.11 Input of Pump duty conditions 51
4.12 Input to generate Meridional Contour 52
4.13 Generated Meridional Contour 52
4.14 Developed Hub Curve 53
4.15 Representation of velocity triangles of centrifugal pump along
with the blade
54
4.16 Designed impeller curve 54
4.17 Completion of Impeller Design 55
4.18 Designed impeller for given duty conditions 55
4.19 Impeller design developed by CFturbo 56
4.20 Input of Volute main dimensions 57
4.21 Development of the pump diffuser 57
4.22 Volute design generated by CFturbo 58
4.23 Predicted Performance Curve 58

xviii
Nomenclature
A0 Inlet cross section area in m2
dsh Diameter of a shaft in meter
dh Diameter of a hub in meter
d1 Diameter of the inlet in m
d2 Diameter of the impeller in m
n Rotational speed, rpm
H Head, Piezo metric height in meter of water
Q Discharge (Flow rate), in m3
/hr
Ns Specific speed
HBEP Head of existing pump at best efficiency point
QBEP Discharge of existing pump at best efficiency point
P Power transmitted to the fluid by the pump in Watt
g Acceleration due to gravity
W Work against gravity
Psh Shaft power in Watt
T Shaft torque in Nm
Cm1 Inlet meridional velocity in m/s
Kcm1 Velocity coefficient by Stepanoff
Q0 Design flow rate in m3
/hr
c1 Absolute velocity at the tip of the impeller in m/s
u1 Peripheral velocity of the impeller at the inlet in m/s
w1 Fluid velocity relative to the impellers at the inlet in m/s
β1 Angle between relative velocity and peripheral velocity at inlet
α1 Angle between peripheral and absolute velocity at inlet
ρ Density of the liquid in kg/m3
ƞ Efficiency of the pump in percentage
b2 Width of the impeller
cm2 Outlet meridional velocity in m/s
c2 Absolute velocity at the end of the impeller in m/s
u2 Peripheral velocity of the impeller at the outlet in m/s
w2 Fluid velocity relative to the impellers at the outlet in m/s

xix
β2 Angle between relative velocity and peripheral velocity at
outlet
α2 Angle between peripheral and absolute velocity at outlet
Z Number of impeller blades

1
1. Introduction
A pump is a machine, which converts mechanical energy to useful hydraulic energy.
Pumps has a vital role in many household and industrial purposes. Various variety of
pumps are used all over the world for different range of applications.
Pumps can be divided or classified mainly into two types, dynamic pumps and positive
displacement pumps. Radial flow (Centrifugal pumps), Axial flow (Propeller pumps) and
mixed flow pumps are referred as dynamic pumps. In Positive displacement (PD) pumps
energy is transferred to the liquid which is in a fixed displacement volume, energy may
be transferred by means of rotating motion of gear or screws, by a casing or a cylinder,
or by reciprocating pistons or plungers. In a dynamic pump energy is transferred to the
pumped fluid with help of a disk with curved vane which rotating on a shaft and is called
as the impeller. The impeller transfer kinetic energy to the fluid by means of its shape and
high rotational velocity. This energy is transformed to pressure energy when the fluid
reaches the pump casing. The pressure head difference between the inlet and the outlet,
or Total Head produced by the pump, is proportional to the impeller speed and diameter.
Therefore, to obtain a higher head, the rotational speed or the impeller diameter can be
increased. Among the various types of pumps, mixed flow pumps are widely used in flood
supply, irrigation, urban water supply, cooling water system for various power plants,
firefighting systems and other numerous other fields.
In the beginning before 1930’s things were pretty basic. But as the demand of transferring
fluids from one area to other increased, new technologies of pumping came up. Low
flow/High Head pump demands was born when the demand of more efficient and
improved chemical reactions started requiring higher process pressure and temperatures.
Now a days Low Flow/High Head Pumps are widely needed in oil & gas, hydrocarbon
processing industry and power generation. They are mainly used to pump Seawater

2
booster for critical services, Light hydrocarbon boosting, Petrochemical processing, Light
Vacuum Gas Oil (LVGO) and Heavy Vacuum Gas Oil (HVGO), Heavy-duty chemical
processing etc. So in order to work on this project a brief history of these pump types are
taken for study and comparing the data such as type of impellers, efficiency, head
coefficient, head-capacity curves and relative cost of the current designs and study the
current needs. The design of the centrifugal pump impeller is not a universally
standardized one. So there is no particular method to follow the basic just start with
finding the specific speed and deciding the impeller type for the parameters given. Every
firm depends on its designer’s experience, expertise and technical intuition to design a
good impeller. The fact that the impeller flow physics has not been understood fully has
led the designers to fall back on tried and tested old design methodologies.
1.1Motivation
As per now there are mainly three pumps which serve this features. They are rotating
casing pump, high speed centrifugal pump and early regenerative turbine pumps. But all
the current design faces difficulties because a speed increasing gear is normally required
to obtain high head and when speed is increased, high internal relative velocities make
pump subject to erosion if abrasives are present. This high speed increases the chance of
wear ring leakage which in turn effect the efficiency of the pump. As due to this high
radial load in single stage design, the shaft may have large amount of deflections which
in turn affects the pumps badly. Many companies use an add on Inducer to produce High
discharge head but it has limitations of achieving the output at near Best Efficiency Point
and the also has limitations to the fluids which can be used to pump through. So an
improved design which eliminates the use of inducer and which doesn’t have higher
operating speed has to be designed.

3
1.2Solution
So it is necessary to have an improved design which can offer the Low flow/High Head
which produces less relative internal velocities and steady output. As per current needs in
industries evaluated, Pumps which serve Low flow/High head has certain difficulties
when it comes in producing high head. To achieve the same, a brief history of these pump
types are taken for study and comparing the data such as type of impellers, efficiency,
head coefficient, head-capacity curves, relative cost of the current designs and study the
current needs and also study of advantages and disadvantages from the design point are
evaluated to prepare an effective plan for execution. With summing up all data collected
and the current market requirement data, mathematically design a pump with property of
Low flow/High Head pumping. A newly advanced software from the turbo machinery
consulting network CFDnetwork ie CFturbo is used for designing purpose , to give a
flexibility to designing job and to evaluate the chances of parameters that we can vary for
advancement of properties.

4
2. Literature Review
Fan Meng conducted an experimental investigation on pressure fluctuations in centrifugal
pump volute casing. The situation was taken under consideration mainly under part load
condition. He took two diffuser volutes for testing, one with a radial diffuser and other
with tangential diffuser. Both volute casing was designed for same impeller parameters,
where 3 impeller diameter was taken for the detailed study. The pump parameters include
rotation speed n=2900rpm, design flow Qdes=330m3
/h and the design head Hdes=48m and
the specific speed Ns= 162.3.A computational grid was generated by ANSYS ICEM-
CFD. With the help of numerical simulations Head Vs Flow chart and Efficiency Vs Flow
chart is plotted for three different flow conditions. Here it was observed two diffusers
were showing almost same results and the volute with radial load had higher efficiency
value at part load and other conditions had almost same results in efficiency. This showed
the necessity of detailed look at part-load condition so the pressure fluctuations were
studied in detailed using statistical analysis method. The result was observed and
concluded that in radial volute the pressure fluctuation intensity was more average so
reduced load on impeller. As my design was planned for low flow / high head pumps
which are mainly required in chemical plants, pressure fluctuation is an important thing
to look into. So a more average pressure fluctuation intensity can be selected over the
other, so the study and development would be focusing on volute with radial diffuser
The study done by Shalin P Marthe and Rishi R Saxena on the performance characteristic
of centrifugal pump on centrifugal pump. They did their study on all the three types of
impeller blade that is the backward, radial and forward bladed impellers and performance
characteristic where evaluated. The design of impeller was done with creo parametrics
with outlet blade angles of 70o
,80o
,90o
and 100o
. In comparison of streamline it was

5
observed that turbulence was least at angle 70o
and high at angle 100o
.If turbulence is
more it can lead to cavitation so 70o
is a safe angle comparatively with other angles taken.
Another simulation done for water vapour contour it was observed that water vapour
formation increases with increase in outlet blade angle. It was also observed that low
pressure zone develops with backward to forward bladed pump. It was concluded that
lower the outlet blade angle the head achieved is low. As my study is for high head pumps
the higher the value of blade outlet angle would benefit me but the chances of high
turbulence may lead to low pressure zones which in turn can cause cavitation.
A study done by Conard C Mowrey eplains the history and theory behind high head /low
flow pumps. The study included comparison of 3 major types of pumps i.e. Rotating
casing pump, High speed centrifugal pump and Regenerative turbine pumps. Among the
all three the high speed centrifugal pumps are the one to be discussed as this is the one
which will help the study of my centrifugal pumps of low flow/high head category. High
speed centrifugal pump had the working principle same but the change is that they use a
different impeller named Barske which needs high speed to pump the fluid and uses a
gear drive to achieve this job. Conard compare the efficiencies and report the drawbacks
of all pumps. The common drawback he found was high internal velocity of these pumps
are a problem when it comes to pumping a fluid which has abrasives and in turn increases
wear ring leakage. They can be only used for low suction applications and as it is high
speed shaft deflection is major issue. So we needed conclude the design should be for
normal speed where fluids with abrasives can be pumped and eliminating the extra gear
drive which may increase maintenance issues.
Mr Ragoth Sing and M. Natraj performed a detailed analysis on a self-designed pump
impeller which they designed using Solid works. SolidWorks Flow Simulation is a
method to approach using computerized method to get detailed results of flow in

6
centrifugal pump. The performance of the centrifugal pump relays on the pump impeller
parameters ad CFD analysis helps the designer to get optimum parameters by simulations.
The methodology they followed contained 6 stages where initially the specification such
as flow and head to obtain in the final pump were selected, next step was to calculate the
design calculations which included calculation of specific speed, output power and torque
and efficiencies hence with the help of this base calculations find the shaft diameter,
velocities and blade angles. Next step was to develop vane profile development where the
vane was designed using Circular arc method and point to point approach with the
calculated values of velocities and angles. Then modelling of impeller using solid works
was done. Flow simulations using CFD was done and evaluated the pressure and velocity
contour of the impeller. Hence it was found that backward curved vanes had better flow
distribution than forward curved. From the results it was notices backward curved
impeller had better efficiency but the head pressure was more for forward curved
impeller. A comparison of both circular methods designed impeller and point to point
approach was done and it was seen the circular arc method designed impeller showed
more efficiency because the variation was minimum. So it can be concluded to follow the
circular arc method to design so as to achieve more efficient results.
E.C. Bacharoudis and A.E. Filios0 performed a study on centrifugal pump impellers. The
Study was done by varying outlet blade angles. Impeller diameter, blade angle and
number of blades are most important parameters which alters the pump properties. Here,
Bacharoudis do the study of 3 impellers where all 3 have same diameter but different
outlet blade angle. The pump impellers he studied was having outlet blade diameter of
20o
,30o
and 50o
and using Fluent an analysis was done for turbulent flow and for the study
of different impellers configuration when the parameters are varied. The analysis was
done on a pump designed on one dimensional flow theory and had characteristics

7
Q=43m3
/h and Head 10m and Speed is 925rpm. According to CFD predictions at nominal
flow rate the value of head was found H=9m.It was concluded that the 10 percentage of
the head loss may be because of not considering 3 dimensional flow structure and less
concentration on hydraulic losses. It was also observed that there is great increase in flow
rate this may also have affected the reduction in head. Considering the 3 impeller outlet
blade angle it was observed that when the outlet blade angle was increased from 20o
- 50o
there was a rise of 6% in head value but overall hydraulic efficiency went down by 5%.
But in when it was working in off-design it was observed that a good improvement was
observed in hydraulic efficiency with the increase of blade angle when it was working at
higher flow rates.
Ms. Shalin Marthe and Mr Rishi Saxena made a research on influence of outlet blade
angle on cavitation in radial pump. Cavitation can be roughly assumed by the water
vapour formation in the pump. The project was performed with different backward blade
impeller with varying in angle. They modelled 6 impeller in ptc creo with different degree
of blade angle for this study. They created models of 20, 30, 40,60,70,80 degree bladed
impeller. Numerical simulations were carried out using Ansys CFX at 2 different flow
rate all this 6 impellers. The first test was done for 1.5 Litres per second and the second
one was for 5 litres per second. It was observed that the blades with greater exit angle,
occupied more water vapour and for taking a single type of blade with the increase of
flow rate , the volume occupied by water vapour is more in greater flow rate. Comparing
the pressure during the analysis greater pressure was observed at the discharge section. It
was also observed that with increase in discharge pressure decreases which may lead to
cavitation phenomenon. While plotting discharge and cavitation number it was observed
much similar characteristics for all angles. As cavitation is not a fair condition for pumps

8
and it may lead to erosion of pump impeller materials, as per the study and analysis done
it was found that angle 20 degree to 30 degree is the fair angle ranges.

9
3. Pump
A pump is a machinery or device for raising, compressing or transferring fluid. A fluid
can be gasses or any liquid. Pumps are one of the most often sold and used mechanical
devices and can be found in almost every industry. Due to this there is a wide range of
different pumps available. In general, the family of pumps is separated into positive
displacement and kinetic pumps. A subcategory of kinetic pumps are centrifugal pumps
which are again separated into radial pumps, mixed flow pumps and axial pumps. But
even at the axial end of the spectrum there is still a part of the energy coming from
centrifugal force unless most of the energy is generated by vane action. On the other hand,
side in radial pumps almost all the energy comes from centrifugal force but there is still
a part coming from vane action. There are also several pumps combining both principles
placed somewhere in between the two extremes in the centrifugal pump spectrum known
as mixed flow impellers. Characteristic for radial pumps are low specific speeds. As
shown in the diagram below there are many options in pump design.
3.1 The pump classification according to the motion of work

10
3.1 Specific Speed
Specific speed is important factor and first step to be calculated while designing a pump.
Specific speed is a value to characterise the shape of a impeller. It is a concept developed
for water turbines in the year 1915, which was later applied to centrifugal pumps by
Stepanoff, 1948. Specific speed is a way to “normalize” the performance of these
hydraulic machines. Low specific speed characterise a radial impeller and is increasing
up to high specific speed at axial impellers. The between are known as Francis‐vane and
mixed‐flow impeller. Specific speed is only a designing engineering significance used to
predict pump characteristics.
Definition of Specific Speed
Specific speed is the speed in RPM at which a given impeller would operate if reduced
(or increased) proportionally in size so as to deliver a capacity of one GPM at a head of
one foot. By itself, this seems meaningless, but taken into the bigger picture, the specific
speed (NS) becomes a dimensionless number that describes the hydraulic features of a
pump, and more specifically of a pump’s impellers. Designers use specific speed, coupled
with modelling laws and other tools such as the affinity laws to fix curve shape, predict
theoretical efficiencies, HP’s, etc.
3.2 Comparison of pump profiles with the values of specific speed

11
Specific speed Ns, is used to characterize turbomachinery speed. Common commercial
and industrial practices use dimensioned versions which are of equal utility. Specific
speed is most commonly used in pump applications to define the suction specific speed -
a quasi-non-dimensional number that categorizes pump impellers as to their type and
proportions. In Imperial units it is defined as the speed in revolutions per minute at which
a geometrically similar impeller would operate if it were of such a size as to deliver one
gallon per minute against one foot of hydraulic head. In metric unit’s flow may be in l/s
or m³/s and head in m, and care must be taken to state the units used.
Specific speed is an index used to predict desired pump or turbine performance. i.e. it
predicts the general shape of a pumps impeller. It is this impeller's shape that predicts its
flow and head characteristics so that the designer can then select a pump or turbine most
appropriate for a particular application. Once the desired specific speed is known, basic
dimensions of the unit's components can be easily calculated.
Note on the Statement, and Calculation of Specific Speed:
• NS of a pump is calculated using BEP of the pump at full diameter.
• On multi-stage pumps, NS is computed for the first stage only.
• If asked specifically by a customer for the NS of a pump being proposed, Ns
should be calculated using the proposed pump’s BEP at the proposed diameter.
4/3
2/1
)(
)(*
H
Qn
Ns =
Where
NS is specific speed (dimensionless)
n is pump rotational speed (rpm)
Q is flowrate at BEP
H is total head (m) at BEP

12
3.1.1 Deriving the formula for Specific speed
Q1 = Capacity of the existing pump at the
(Best Efficiency Point) BEP
H1 = Head of the existing pump at BEP
N1 = rotating speed of the existing pump
Q2 = Capacity of the new pump at the
(Best Efficiency Point) BEP
H2 = Head of the new pump at BEP
N2 = rotating speed of the new pump
As per the model law






=
1
23
1
2
N
N
f
Q
Q
from which
3
1
2
1
3
1
1
2












=
N
N
Q
Q
f
2
1
22
1
2






=
N
N
f
H
H
from which
2
1
1
2
2
1












=
H
H
N
N
f
Equating the two f
2
1
1
2
2
1
3
1
2
1
3
1
1
2












=











H
H
N
N
N
N
Q
Q
Combining 21 NN
2
1
1
2
3
2
2
1
3
1
1
2












=





H
H
N
N
Q
Q
Taking all to the 3/2 power
4
3
1
2
2
1
2
1
1
2












=





H
H
N
N
Q
Q
Grouping them with subscripts
4
3
1
2
1
11
4
3
2
2
1
22
H
QN
H
QN
= So we can conclude
4
3
2
1
BEP
BEP
s
H
NQ
N =

13
3.1.2 Specific Speed Basics
The specific speed is largely related to the
impeller discharge angle, relative to the
inlet. Pumps in which the discharge of the
impeller is directly radial to the suction; that
is, where the flow transitions “rapidly” from
one plane to the other, have a low specific
speed and are called “radial vane” or “radial
flow impellers.” The NS will be in the
neighbourhood of from 500 to 1700. These
pumps will usually exhibit a “low flow to
head” ratio.
At the “other end of the scale,” the fluid will
be discharged from the impeller along the
same axis as it enters. These impellers (or
propellers) have high specific speeds,
generally above 10000, and are referred to as axial flow impellers and have a high flow
to head ratio.
Between these two extremes fall:
Mixed Flow Impellers, with an NS range of from around 4000 to 10000 begin to transition
away from the suction axis, but discharge between the axial and radial angles, and
generally exhibit a high flow to moderate head ratio.
• Francis vane Impellers, between mixed and radial flow, with NS values from
around 1700 to 4000. Francis vane impellers are frequently discussed in the industry, and
3.3 Figure illustrates the general appearance of
impellers of various specific speed designs, with inlet
being at the bottom.

14
are simply impellers which have vanes curvature such that the transition from the inlet
axis to radial axis is completed more gradually.
3.1.3 Specific Speed and The Pump Curve
The specific speed of the pump will determine
the general (theoretical) curve slope, from
shutoff to runout. Low specific speed impellers
will exhibit “flat head” curves and high specific
speed pumps will exhibit steeper curves.
(Remember that specific speed is taken for a
single stage only. Many multi-stage pumps
exhibit steep curves when all stages are shown
together, but have a low NS.
So we are only going to consider radial flow as
the purpose is Low Flow/High head to pump
chemicals so the impellers and casing would be discussed according to the view of
centrifugal pumps. The value of Ns is calculated in calculations to finalise and conclude
with a proof that the pump going to be designed should be of centrifugal range. From the
specific speed we can obtain head coefficient and hence we can calculate the diameter of
the impeller blade.
3.2 Impellers
Impeller is the device which rotates, and transfer energy to fluid. And is functioned by
transferring energy from the motor that drives the pump to the fluid being pumped by
accelerating the fluid outwards from the centre of rotation.
3.4 HQ and BHP curves for various specific
speeds

15
Impellers are usually classified in two ways:
Specific Speed (NS): The relationship between the amount of flow an impeller produces
and the amount of head or pressure generated is called specific speed.
Physical Design: Details such as whether an impeller is open or enclosed, whether it is
single or double suction, and the way the impeller vanes are designed can all be used to
describe and classify impellers. So as from specific speed previously described we
concluded to discuss only on radial impeller. So we will consider only centrifugal range
of impellers
Impellers can be split into 3 as Closed impeller, Semi-Open Impeller and Open impeller
A Closed impellers have a back and front wall around the vanes, to increase strength.
Closed impellers are used primarily in larger pumps and can be used in applications that
handle suspended-solid service. These types of impellers are commonly found in clear
liquid applications. They don't do well with solids and are difficult to clean if they become
clogged.
A Semi-Closed impellers have a back wall that adds strength to the impeller. Semi-closed
impellers are usually used with liquids or products that have solids. Reduced efficiency
is a common problem with semi-closed impellers, but the ability to pass solids is an
important trade-off.
3.5 Open , Semi-Open and Closed Impeller designs

16
An open impeller has vanes that are attached to a centre hub and mounted directly onto a
shaft. There is no wall surrounding the vanes which makes open impellers weaker than
closed or semi-closed valves. Open impellers are generally faster and easier to clean and
repair. Open impellers are usually used in smaller pumps and pumps that handle
suspended solids.
As the design is planned for a process pump it should have capability of handling solids
too. So the design is done as semi-open impeller.
3.2.1 Vane Design
The impeller vanes are the heart of the impeller. The rest of the impeller design is just
there to carry, protect and balance the impeller vanes during operation. Some impellers
have many vanes and tight internal clearances. These are typically intended for water
service and generally fall between the radial-vane and Francis-vane specific speed fields.
Other impellers have just one or two vanes and large internal clearances. These types are
often called solids-handling impellers and generally fall between the Francis-vane and
mixed-flow fields. Still others are designed with a single vane and no lower shroud, or
with vanes that do not extend very far down into product being pumped. These are called
screw and vortex impellers respectively, and are intended for applications with a high
concentration of solids. Finally, there are impellers will no shroud at all, top or bottom,
such as what you see in the axial-flow field. Now pumps with recessed impellers are also
available where vortex flow principles are taken, 20% fluid come in contact with impeller.
3.2.2 Impeller Blade Types
A centrifugal pump can be classified according to the impeller blade orientation
1. Forward Swept vanes

17
2. Radial exit vanes
3. Backward swept vanes
Forward swept impeller have outlet blade angle greater than 90o
, for Radial exit blade the
outlet blade angle is 90o
and for Backward swept impeller blade have outlet blade angle
less than 90o
.
As from studies we can discuss these three blading in view to achieve high head.
First thing to be consider is the Water Vapour contour. We can explain the same by giving
an example of simulation 3 types of impeller blades.
Here it was observed as the blade
increase angle from 70 to 100, the water
vapour formation increases. So the 70o
impeller is mostly chosen for normal
centrifugal pump because as it becomes
forward vane the chance of formation of
water vapour is high.
The performance of the centrifugal pump
depends upon the size and configuration of the vanes used. As we can see in the head –
3.6 Forward, Radial and Backward facing impeller vanes
3.7 Water Vapour contour at 20 lps

18
discharge curve plotted for the same
impeller blades previously simulated, it
was seen that for higher outlet blade
angle a high head was observed
compared to other blades. So for we
conclude while designing a pump for
high head forward and radial exit vanes
are preferred over the backward curved.
3.2.3 Impeller Design
The impeller dimensions are designed based on the head and discharge. The following
are the steps involved in designing a centrifugal impeller
• From the head (H) and discharge (Q), the kinematic specific speed (NS) is
calculated
4
3
2
1
H
NQ
Ns =
• Shaft power is determined with the help of Hydraulic power and assumed
efficiency
The hydraulic power may be calculated using one of the formulae below
81.9*** HQP ρ=
P = Power transmitted to the fluid by the pump in Watt.
Q = Flow in m3
/s i.e. 2/3600 m3
/s & 25/3600 m3
/s
ρ = Density of the liquid in kg/m3
. = 1000kg/m3
H = Piezo metric height in meter of water= 35m
3.8 Head-Discharge characteristics of different blades

19
Average Intensity of gravity, g =9.81m/s2
Well we've got a vertical outlet, moving the fluid upwards against gravity by a
certain height, H (the head). So we can say the pump is doing work against gravity
ie.
W = Force x distance
= weight of fluid x head = m g H
Power is the rate of doing work so.
P = m g H/t = (m/t) x g x H
= (mass flow rate) x g x H
= (density of fluid x volumetric flow rate) x g x H
81.9*** HQP ρ=
Now finding shaft power
η
P
Psh =
• Now we will find shaft diameter
3
16
s
sh
s
T
d
π
=
Where T is the torque to be transmitted through the shaft and Ss is Shear stress
Shaft torque T
Rpm
kWP
T sh )(
*9552=
• Inlet velocity (U1) is estimated using

20
11 95.0 mcU ×=
gHKc cmm 211
=
Cm1The inlet meridional velocity
• From the inlet velocity and the new discharge (Q°) calculated after accounting for
volumetric efficiency, the inlet cross section area (A0) is calculated
1
0
0 / mCQA =
• From the area, the inlet diameter (d1) is calculated
π
0
1
4A
d =
• Blade inlet angle (β1) is calculated
1
1
1tan
u
Cm
=β
60
1
1
nd
u
π
=
• Breadth of the impeller (b1) at the inlet is
1
1
1
d
A
b
π
=
• Blade outlet angle (β2)
• The outlet peripheral velocity (u2) can be calculated as follows
gHKu u 222 =
where Ku2 is the experimental velocity coefficient.
• The outlet diameter (d2) and breadth of impeller at outlet are

21
n
u
d
π
2
2
60
=
2
2
2
d
A
b
π
=
3.2.4 Velocity triangles
In pumps, motion of the fluid is needed to be specified according to the rotating motion
of the impeller. The absolute velocity c can be regarded as velocity relative to the casing
of the pump. So this can be sum of two velocities that is the peripheral velocity of the
impeller u and fluid velocity relative to the impeller w
wuc +=
When the velocities are plotted, it forms a velocity triangle or a velocity parallelogram.
The velocities are given subscripts to give an identity for inlet and outlet of impeller,
generally they are given as 1 and 2 respectively.  and β are angle which represent angle
of the absolute and relative velocities at inlet and outlet of impeller.
When it is in pumps with axial inlet we usually assume swirl to be zero, so when swirl is
zero that means angle of absolute at the inlet 1 will be 90o.
Now looking theoretically
angle at outlet β2 is aligned with the camber angle of impeller but in reality this angle
deviates due to slip and blockage in blade.
3.9 Velocity diagram in an impeller stage 3.10 Velocity triangles

22
The peripheral velocity u can be easily calculated by knowing rotational speed n of the
impeller and diameter where the velocity is evaluated, d.
60
n
du π=
The absolute velocity c , can be split into meridional velocity cm and peripheral velocity
cu . With swirl = zero at the inlet cu1 is negligible so cm1 = c1.
By taking conservation of mass the relation between both meridional velocities at inlet
and outlet can be found. Qla is the volume flow passed through the impeller A1 and A2 is
area at inlet and outlet of impeller respectively
2
1
2
2
2
A
A
c
A
Q
c m
la
m ==
Areas A1 and A2 can be calculated with the help of dh the hub diameter, d0 impeller eye
diameter, d2 diameter at the outlet and b2 the width at outlet of the impeller
22
01
4
hddA −=
π
222 bdA π=

23
3.3 Casing
A centrifugal pump consists of a set of rotating vanes enclosed within a housing which is
called casing, that is used to impart energy to a fluid through centrifugal force. Radial
forces are a result of the static pressure in the casing. Therefore, axial deflections may
occur and lead to interference between the impeller and the casing. The magnitude and
the direction of the radial force depend on the flow rate and the head.
When designing the casing for the pump, it is possible to control the hydraulic radial
forces. Two casing types are worth mentioning: the single-volute casing and the double-
volute casing.
As you can tell from figure, both casings are shaped as a volute. The difference between
them is that the double-volute has a guide vane. The primary purpose of a volute casing
is to convert the kinetic energy into pressure.
3.11 Single-volute casing & Double-volute casing

24
Single volute
Single-volute pumps have been in existence from Day One. Pump volute has single lip
so Pumps designed using single- volume casings of constant velocity design are more
efficient than those using more complicated volute designs. Single volute is usually used
in small low capacity pumps where a double volute design is impractical due to relatively
small size of volute passageway which make obtaining good quality commercial casting
difficult. Pumps with single volute design have higher radial loads. The single-volute
pump is characterised by a symmetric pressure in the volute at the optimum efficiency
point, which leads to zero radial load. At all
other points, the pressure around the impeller
is not regular and consequently a radial force
is present.
Double volute
Pump volute has dual lips located 180 degrees
apart resulting in balanced radial loads and
most centrifugal pumps are of double volute design. The main advantage of a double
volute over a single volute is the balancing of radial loads on the impeller, as the double
cutwater construction leads to a more equal pressure distribution in the volute, as from
figure we can see the double-volute casing develops a constant low radial reaction force
at any capacity. Minimizing the radial load on the bearings over the full operating range
can have a significant impact on the lifetime of a pump, since bearing failures are the
second most common reason for pump failures. However, a double volute adds additional
hydraulic resistance
3.12 Radial force for single and double-volute
casing

25
Diffuser casing
In addition to single and double volute casings, another type of pump casing exists: the
diffuser casing.
The stator section of a centrifugal pump, after flow exits the impeller, is usually either a
diffuser or a volute. The purpose of each of these two stator types is to efficiently diffuse
velocity energy into pressure. Diffusers are characterized by a plurality of radially
symmetric diffusing passageways surrounding the impeller. Either a volute-shaped or
annular collector is used in tandem with the diffuser.
For a single stage centrifugal pump, a diffuser type design is usually costlier to produce
because the diffuser ring is an extra part plus some incremental added machining for the
casing. The casing must still function as a collector to convey the flow from the diffuser
to the discharge nozzle. No matter how this is done, the diffuser can offer little
comparative advantage in the size of a single-stage pump.
Diffuser designs are often more efficient at the best efficiency rate of flow, compared to
that of a volute. Also, a custom diffuser can be made for each application in order to
maximize the efficiency for a specific duty point.
A volute proponent might argue that the diffuser is less efficient at off-peak flow rates
where the pump will operate a good portion of the time. The efficiency differences may
not be significant and unless large amounts of power are involved, these debates seldom
carry much weight in relation to the competing prices of the pumps offered, or user
preference for either volute or diffuser.
Radial thrust acting on the impeller develops from a non-uniform circumferential pressure
distribution. The stator design plays an important role in this. For some applications,
especially with a single-stage overhung impeller type pump that will operate continuously

26
at flows substantially away from its Best Efficiency Point flow, a diffuser/collector
arrangement can provide a lower magnitude of radial thrust.
Volute type casings are the norm for solids handling pumps that require wide open
passageways. A few specialty high pressure single casing pumps utilize the structural
support that the vane diffuser can provide for the collector scroll.
Vertical turbine and vertical bowl type casings are mostly of the vane diffuser type.
Manufacturers have generally rationalized the choice of pump stator based on market
needs, application requirements and production costs. Any evaluation regarding the
selection of a diffuser or a volute should be considered in the context of specific pump
types, specific applications and manufacturers’ product offerings.
3.3.1 Single Volute casing for Process Pump
We would be going for single volute design as the pump to be designed is a process pump
of Low Flow / High head and as per the head and discharge took, there load and size of
pump is comparatively low so the chance of use of double volute and diffuser is
eliminated. Now we can discuss mainly on Single volute casing types. Taking in
consideration of study from Fan Meng, regarding “Effect of two diffuser types of volute
on pressure fluctuation in centrifugal pump under part-load condition” a study and
3.13 Casing with and without Diffuser

27
conclusion is obtained to finalize the type of design of volute casing to be taken for the
pump to be designed. In this study done by Meng, two diffuser types of volute were
designed to study the effect on pressure fluctuation features of centrifugal pump under
part-load condition, with the same volute design parameters and impeller parameters.
The two type of volute taken for study are (A) The volute with radial diffuser (B) The
volute with tangential diffuser
The volute with radial diffuser have centre line discharge. Casing for the American
National Standards Institute(ANSI) chemical process pumps have centreline discharge
and suction both in the horizontal and vertical pump. This make it easier for laying out
the pipe in the system and to reduce the loading in nozzle. It is because centreline nozzle
eliminates the moment arm from the centreline of the casing to the centreline of the nozzle
that exists with tangential discharge.
3.14 The volute with radial diffuser
3.15 The volute with tangential diffuser

28
3.3.2 Results and Conclusions
Comparison of Performance
By using simulation under three different rate of flow, the pump performances is obtained
for the 2 different type of volute.
From the figure, the values of head in the volute with radial diffuser and tangential
diffuser decrease gradually with the increase of flow rate. In the volute with radial
diffuser, the maximum head value is 60meter, and the minimum head value is 47meter.
In the volute with tangential diffuser, the maximum head value is 60.2meter, and the
minimum head value is 46.9meter.
In the two diffuser types of volute, the efficiency has large difference at part load
condition and for other condition it is almost the same. So, it is necessary and important
to study the pressure fluctuation under part-load condition to finalize the type of volute
casing is to be selected for the design.
Pressure fluctuation intensity distribution
As study done by Meng , pressure fluctuation under part-load conditions are taken by
creating two cross section of fluid channel respectively in the diffuser and near tongue of
volute with two diffuser types
3.16 Head vs Discharge 3.17 Head vs Efficiency

29
From figure, in the diffuser of volute, both in radial diffuser or tangential diffuser, under
part load condition, the pressure fluctuation intensity of outer side is higher than that of
inner side. The large gradient distribution of pressure fluctuation intensity was shown in
the position close to the wall. Comparing the pressure fluctuation intensity in radial
diffuser with that of tangential diffuser, the distribution of pressure fluctuation intensity
is more average in radial diffuser. So the radial diffuser is more convenient in using for
process pumps where chemicals are used.
3.18 The pressure fluctuation intensity in the diffuser of volute

30
4. Mathematical Design of Low Flow/High Head
Pump – Calculations
Impeller is designed on the basic of design flow rate, pump head and pump specific speed.
So, the design data are required to design the centrifugal pump. The client will usually
specify the desired head, pump capacity and purpose. The rest of the pump parameters
are derived with standardized formulas.
4.1 Duty Conditions
For design calculation, the design parameters are taken as follows:
Flow-Q(m3
/h) Head-H (m) Min/Max
2 20 Min
25 35 Max
4.2 Calculations
4.2.1 Pump Rotational Speed – n (rpm)
Rotational speed (also called speed, or speed of rotation) can be quantified as the number
of revolutions a rotating system makes within a defined period of time. The unit used for
rotational speed is s–1
(rev/s); pump speed is generally given in min–1
(rpm).
The rotating frequency of the pump shaft therefore characterises a pump's rotational
speed(n). It should not be confused with specific speed (Ns) and is always defined as a
positive figure. The pump direction of rotation is specified as clockwise or anti-clockwise
and is separate to the defined direction of rotation of the impeller, which, when turning
to the right with respect to the direction of inflow, is clockwise.

31
The selection of pump rotational speed is closely related to the characteristics of the pump
hydraulic system (circumferential speed, impeller, specific speed), as the overall strength
and economic efficiency of the pump and drive system need to be taken into account.
Most pumps operate at rotational speeds between 1000 and 3000 rpm but frequently
reach in excess of 6,000 rpm with special gearing and turbine drives.
pfn /)120*(=
f - Frequency (Hz) = 50Hz
p - Number of Poles = 2
rpmn 30002/)120*50( ==
pfn /)120*(=
f - Frequency (Hz) = 50Hz
p - Number of Poles = 4
rpmn 15004/)120*50( ==
So we can see for 4 pole it is 1500rpm and 2 pole it is 3000rpm, but we have to consider
slip also. So a basic understanding is required.
4.1 Representation of magnetic poles in a pump motor

32
Basic Understanding - Difference Between 2 Pole and 4 Pole
Motors
A motor is an electric device that converts electrical energy into mechanical energy,
specifically in the form of torque delivered through a shaft. The motors operate on the
principle of electromagnetic induction as described by Michael Faraday.
2-Pole Motor
A motor that contain two poles or a single pair of magnetic poles north and south are said
to be a 2-pole motor. Often stator windings are the north and south poles. Number of
stator windings can give any reasonable number of poles ranging from 2 to 12. Motors
with more than 12 poles are available, but they are not in common use. The synchronous
speeds of the motors are directly dependent on the number of poles as given in the
following expression
Synchronous speed of the motor = (120×frequency)/(number of poles)
Therefore, the speed of a 2 pole motor connected to the main power has 3000rpm
synchronous speed. With the rated load, operating speeds may decrease to about 2900rpm
due to both slip and load. In two pole motors, the rotor turns 1800rpm in half the cycle.
4.2 Synchronous and approximate full load speed of AC electrical motors
A motor with 4 poles runs with frequency 50 Hz. The synchronous speed is 1500 rpm and the typical full load speed is
1450 rpm. The slip is the difference between synchronous and load speed - 50 rpm.

33
Therefore, on one cycle of the source, rotor makes one cycle. The amount of energy used
is relatively low in two pole motors, and the torque delivered is also low.
4-Pole Motor
A motor that contains four poles in the stator or two pairs of magnetic poles in alternating
order; N > S > N > S. The synchronous speed of a four pole motor connected to the mains
power is 1500rpm, which is half the speed of the 2-pole motor. With the rated load,
operating speeds can decrease to a value around 1450rpm.
In four pole motors, the rotor turns 900rpm for every half cycle. Therefore, the rotor
completes 1 cycle for every two cycles of the source. Hence the amount of energy
consumed is twice the amount of 2 pole motor and theoretically delivers twice the torque.
So We calculate for 2 pole and 4 pole i.e. rpm = 2900 and 1450 (respectively)
4.2.2 Specific Speed-Optimum Geometry versus Specific
Speed
Fundamental to any system of classifying pumps is the rotor geometry that is optimum
for each type, as illustrated in Figure in terms of the specific speed Ns. Here Q is the
volume flow rate or capacity, n is the rotating speed and Ω is the angular speed, and ∆H
or just H is the pump head—all at the best efficiency point (BEP). In this case the pump
4.3 Optimum geometry as a function of BEP specific speed (for single stage rotors).

34
performance in terms of the head coefficient ψ = g∆H/(Ω2
r2
)is influenced only by the
flow coefficient or specific flow Qs = Q/(Ωr3
).
4/3
2/1
)(
)(*
H
Qn
Ns =
667.1007
)35(
)25(*2900
4/3
2/1
==sN
The lower values (500 to 1500) on the left of the figure describe the geometry of the radial
vane impeller while the higher values (9000 and higher) on the right of the figure equate
to true axial flow impellers (propellers). A radial vane impeller discharges 100% of its
flow perpendicular to its suction, usually with a low flow-to-head ratio.
So from figure we get our Specific Speed Ns=1007.667, Comes in Centrifugal range
(Radial Impeller Ns) . And ψ = 0.45 (From Figure)
Approximate Relative Diameter of the impeller (d2)
ψπ
Hg
N
d ∆=
1
2
n is the rotating speed and = 2900rpm
∆H (or just H) is the pump head = 35m
The head coefficient (ψ) = 0.45
md 182.0
45.0
)35*81.9(
2900*
60
2 ==
π
d2= 182mm
Or by deriving Impeller diameter, d2
We Know,
2
ωdmrdF =

35
AdFdP =
drbrdddm φρρ =∀=
Substitute for dF and dm . Write dP
)(
2
2
1
2
2
22
1
2
2
1
2
1
2
rrrdr
brd
rdrbrd
dP −==
⋅
= ∫∫ ∫
ρω
ρω
φ
ωφρ
)(
2
2
1
2
2
2
12 rrPP −=−
ρω
divide by ρg
g
rr
g
PP
ρ
ρω
ρ






−
=
−
)(
2
2
1
2
2
2
12
( )
g
rr
g
P
g
P
2
2
1
22
2
2
12 ωω
ρρ
−
=−
So we Can write
( )
g
r
g
P
2
2
1
2
2
2
2
ω
ρ
=
We know
g
P
H
ρ
=
ωrU =
So ,
g
U
H
2
2
2
2 =
Substituting d2 /2(ω) for U2 and solving for D2

36
N
H
N
HgH
d 222
2
6.84
2
81.9260222
=
×××
==
πω
So impeller diameter d2
Multiplying the right side by an experimentally determined coefficient Φ
N
H
d 2
2
6.84 ×Φ×
=
Most of the plotted points fall within a range of 0.9 to 1.1. for Φ . We take 1.05 as it is
calculated for high head
mmmd 2.1811812.0
2900
3505.16.84
2 ==
××
=
d2= 181.2mm
4.2.3 Pump power output or Hydraulic power
The Pump power output or Hydraulic power which is also known as absorbed power,
represents the energy imparted on the fluid being pumped to increase its velocity and
pressure. The hydraulic power may be calculated using one of the formulae below,
depending on the available data.
81.9*** HQP ρ=
P = Power transmitted to the fluid by the pump in Watt.
Q = Flow in m3/s i.e. 2/3600 m3
/s & 25/3600 m3
/s
ρ = Density of the liquid in kg/m3
= 1000kg/m3
H = Piezo metric height in meter of water= 35m
Average Intensity of gravity, g =9.81m/s2

37
So all factors are known, to find the Pump Power values are being substituted
Well we've got a vertical outlet, moving the fluid upwards against gravity by a certain
height, H (the head). So we can say the pump is doing work against gravity ie.
W = Force x distance
= weight of fluid x head = m g H
Power is the rate of doing work so.
P= m g H/t = (m/t) x g x H
= (mass flow rate) x g x H
= (density of fluid x volumetric flow rate) x g x H
81.9*** HQP ρ=
wP 75.19081.9*35*1000*
3600
2 ==
wP 2384.37581.9*35*1000*
3600
25 ==
P= 0.191kW=0.256hp (For Initial 2 m3
/h)
P = 2.384kW=3.196hp (For Final 25 m3
/h)
4.2.4 Torque
The torque, or turning moment, for a pump may be estimated from the power curve in SI
units by
For 1450rpm
Rpm
kWP )(
*9552=τ
Nm258.1
1450
191.0
*9552 ==

38
Nm704.15
1450
384.2
*9552 ==
T= 1.258Nm (For Initial 2 m3
/h)
T = 15.704Nm (For Final 25 m3
/h)
For 2900rpm
Rpm
kWP
T
)(
*9552`=
Nm629.0
2900
191.0
*9552 ==
Nm852.7
2900
384.2
*9552 ==
T= 0.629Nm (For Initial 2 m3
/h)
T = 7.852Nm (For Final 25 m3
/h)
4.2.5 Brake Horse Power(BHP)
Before brake horse power we should look into efficiency of a pump
Efficiency is equal to water horse power divided by brake horse power
Hence,
η
WHP
BHP =
From the figure shown below, with the given flow of 25m3
/h and calculated value of
specific speed is 1007.667rpm, we select a tentative value of efficiency of 70%. The
radial range of specific speed was calculated previously.

39
So we know from previous figure we take
%70=η
And from previous calculations
hpWHP 196.3=
So then BHP is equal to
7.0
196.3 hp
BHP =
kwhpBHP 404.3565.4 ==
4.2.6 Shaft Torque
At 2900 rpm
Rpm
kWP
T
)(
*9552=
4.4 Approximate relative impeller shapes and efficiencies as related to specific speed

40
Nm212.11
2900
3.404
*9552 ==
NmT 212.11=
4.2.7 Shaft diameter
Calculate shaft diameter based on torque. Increase the calculated value somewhat to
allow for bending moment which is unknown at this point and to ensure that the critical
speed exceeds the operational speed by a reasonable margin. The bending moment will
depend on the weight distribution of the shaft and any unbalanced radial thrust acting on
the impeller.
Diameter of shaft is equal to
3
16
s
sh
s
T
d
π
=
Where T is the torque to be transmitted through the shaft and Ss is Shear stress in psi and
is taken as 8Mpa ie 1160.3psi (Assuming)
Shaft Diameter dsh
3
16
s
sh
s
T
d
π
=
3
757.68943.1160
212.1116
××
×
=
π
shd mmmdsh 19.205030.019205 ==
We need to consider a Factor of Safety of 1.15 (General Consideration in design of shaft)
So our required minimum shaft diameter is 19.20505 x SF
22.0857815.120505.1920505.19 =×=×= SFdsh
dsh=22mm

41
Hub diameter dh = (1.5)dsh
shh dd ×= 5.1
So ,
mmdh 33225.1 =×= dh = 3.3cm
Here dh and d1 is took same
4.2.8 Inputs to CFturbo 10 to generate the inlet meridional
velocity
The inputs are given. The type of machine configuration will automatically change to
Radial (High Pressure) after the input of flow, head and speed.
4.5 Global Setup – Pump Inputs
The design is done for clean water as mentioned earlier, so fluid selected is Water.
In impeller main dimension section the hydraulic efficiency is given as 70% (which is
obtained from Figure 4.4 previously. Input the allowable shear stress for shaft which was
taken 8 Mpa ie 1160.3 psi for shaft diameter calculations and Factor of safety
22mm=dsh

42
1.15.Looking the compatibility, the calculated shaft diameter and diameter obtained by
software is greatly similar. The hub diameter of 33mm is also entered (Calculated value)
Now click calculate to obtain rest of values.
4.6 Input of Hub Diameter to get Input Meridional velocity
4.2.9 The inlet meridional velocity
gHKc cmm 211
=
where Kcm1 is the velocity coefficient by Stepanoff (1957) modified form in figure
Cm1 is Inlet meridional velocity , But we take the value of Cm1 with the assistance of
CFturbo 10 by giving the input values known.
smcm /3.21
= - From figure 4.6 The average inlet velocity is the inlet meridional velocity
The inlet cross section area (A0)
1
0
0 / mCQA =

43
A common volumetric efficiency for centrifugal pumps is 96%. Therefore, the design Q
becomes
96.0/0
QQ =
hmQ /042.2696.0/25 30
==
2
1
0
0 0.003145)36003.2/(042.26/ mcQA m =×==
2
45.310 cmA = A0 = 31.45cm2
4.2.10 The inlet diameter (d0)
π
0
0
4A
d =
cmd 33.6
45.314
0 =
×
=
π
d0 = 6.33cm
4.2.11 Impeller Width Calculation (Inlet)
1
1
1 d
Ab
π=
A1 - The inlet cross section area
( )
cmb 033.3
3.3
45.31
1 =
×
=
π
b1 = 30.33mm
4.2.12 Blade inlet angle (β1)
When dealing with an axial inlet we usually assume zero swirl, meaning 1 = 90o
The absolute velocity, c, can be decomposed into meridional and peripheral components
with subscripts m and u
With zero swirl at the inlet cu1 is negligible, and cm1 = c1

44
1
1
1tan
u
Cm
=β
sm
nd
u /011.5
60100
29003.3
60
1
1 =
×
××
==
ππ
smCC m /3.211 ==
°=





=





= −−
66.24
011.5
3.2
tantan 1
1
11
1
u
Cm
β β1 = 24.66o
4.2.13 Fluid velocity relative to the impellers
Cos β1=U1/W1
W1=U1/Sin β1 = 5.011/Cos (24.66o
) = 5.513m/s
4.2.14 Inlet velocity triangle
Plotting the inlet conditions
4.7 Inlet Velocity triangle – With calculated Values

45
4.2.15 Inputs to CFturbo 10 to generate the outlet meridional
velocity
The earlier option shown in figure 4.6 , which represents the main dimensions . Now we
would update the calculated impeller diameter and outlet width because this both
calculated values influences the outlet conditions.
4.8 Input of Impeller Diameter and Impeller outlet width
4.2.16 Outlet meridional velocity
Cm2 is Outlet meridional velocity
We take the value of Cm2 with the assistance of CFturbo 10 by giving values calculated
till now. From Figure 4.8 we can get Cm . The outlet conditions such as the outlet width
and diameter of the impeller influences the outlet meridional velocity so we input the
calculated values.
smcm /4.02 = cm2 = 0.4m/s

46
4.2.17 Blade outlet angle (β2)
To find u2 (m/s)
60
2
2
nd
u
π
=
d2 is Approximate relative Diameter of impeller which is equal to 182mm
n is the speed ie 2900rpm
sm
nd
u /64.27
601000
2900182
60
2
2 =
×
××
==
ππ
From figure 4.8 we see the generated values
• The outlet fluid velocity relative to the impeller w2=4.5 m/s
• Outlet meridional velocity cm2=0.4 m/s
o
m
Sin
w
c
Sin
1.5
5.4
4.01
2
2
2
2
==
=
−
β
β
Drawing the velocity triangle, with values u2 , Cm2 , w2 and β2
The outlet absolute velocity, c2 = 23.16m/s
c2 = 23.16m/s
4.9 Outlet Velocity triangle – With calculated Values

47
4.2.18 Number of blades, Z






=
3
2β
z - Stepanoff’s (1957) number of blades equation
But now a days, commonly used equation is Pfleiderer’s (1961)
( ) ( )[ ] 




 +
−+=
2
5.6 21
1212
ββ
Sinddddz
• d2 Approximate Relative Diameter = 182mm
• d1 impeller inlet diameter = 33mm
• Blade outlet angle β2 = 5.1o
• Blade inlet angle β1 = 24.64o
( ) ( )[ ] 




 +
−+=
2
1.566.24
33182331825.6 Sinz
z = 7
4.2.19 Impeller Width Calculation
Area at the outlet of the impeller A2
22 mCAQ ×=
22 / mCQA =
sm
hmA
4.0
25 3
2 =
36004.0
25 3
2 ×
=
sm
hmA
( ) ( )[ ] ( ) 9.68.141492155.6 == Sinz
7=z

48
2
2 017361.0 mA =
We know 222 bdA π=
Known values are d2 and A2
By re arranging we get
2
2
2 d
Ab
π=
We need to consider number of blades to find, so we divide the circumference by number
of blades
( )
cmmb 03.3030379.0
182.0
017361.0
2 ==
×
=
π
Now when we compare the impeller width obtained at inlet and outlet give same values
So we can say,
Impeller Width b2=3.03cm b2=3.03cm
4.2.20 Calculated Values
• n is pump rotational speed (rpm) = 2900rpm
• Ns Specific speed = 1007.667
• d2 Approximate Relative Diameter = 182mm
For Flow 2 m3/h
o P Pump Power = 0.191kW=0.256hp
o Torque = 1.258Nm(1450 rpm)
For Flow 25 m3/h

49
o P Pump Power = 4.6325kW=3.196hp
• Brake Horse Power(BHP) = 4.565hp = 3.404kw
• Shaft Torque at 1450rpm T = 22.424Nm
• Shaft Torque at 2900rpm T = 11.212Nm
• Diameter of shaft dsh=22mm
• Inlet meridional velocity cm = 2.3m/s
• The inlet cross section area A0 = 31.45cm2
• The inlet diameter d0 = 7cm
• Hub diameter dh=3.3cm
• Blade inlet angle β1 = 24.66o
• The inlet absolute velocity, c1 = 2.3m/s
• The inlet peripheral velocity of the impeller u1 = 5.011m/s
• The inlet fluid velocity relative to the impeller w1=5.513m/s
• Outlet meridional velocity Cm2=0.4m/s
• The Outlet peripheral velocity of the impeller u2 = 27.64m/s
• The outlet absolute velocity, c2 = 23.16m/s
• The outlet fluid velocity relative to the impeller w2=4.5m/s
• Blade outlet angle β2 = 5.1o

50
• Number of blades, Z = 7
• Impeller Width b2=3.03cm
4.10 Velocity triangles generated by CFturbo

51
4.3 Designing of Impeller
The design of impeller has different steps. The impeller blade curve modelling and design
of total impeller is completed with the help of Creo 4.0 and the Impeller hub flow pattern
and designs are obtained by using CFturbo 10. The designing included mainly 2 steps,
the designing of Impeller curve and design of Meridional Contour.
CFturbo is a software package that offers design of turbo machineries such as Pump,
Ventilator, Compressor and Turbine. This software is an advanced thinking of a
consulting office named CFDnetwork. In the year 2008 they developed this advance
software with completely focusing on to reduce the complications in turbomachinery
designing and the flexibility of the software pave way for advanced tasks.
The designing of Meridional Contour is done with the help of CFturbo 10. The designing
includes manual methods where we can create it flexibly with the calculated values.
Selecting a new work environment in CFturbo we select the pump interface where we can
input the Flow, Head and Speed which gives real-time outputs of the machine range in
right side of the window.
4.11 Input of Pump duty conditions

52
The hub designing involves the modelling of meridional contour. The modelling is done
by giving inputs to main dimension windows. The calculated values above are used to
generate the same.
4.12 Input to generate Meridional Contour
The Meridional Contour is generated and below showed is the 2D representation
4.13 Generated Meridional Contour

53
Generating the Hub curve from 3D representation
Now the curve is exported to Creo 4.0 for further development
4.14 Developed Hub Curve

54
Now the development of Impeller curve by creating
a plane at the base of the hub curve. Sketch and
divide the hub to impeller circles into 10 parts. The
construction is done with the point of view of Inlet
and outlet angles. So as we design a single curvature
model , the design is from angle 24.66 to 5.1. Select
the start point and draw a line with angle same as
inlet angle β1 from the tangent. Here the point of
tangency and the start point is taken as same and continue the same for all the 10 equally
divided points. Hence we get point and we join all with help of spline.
4.16 Designed impeller curve
Now we thicken and extrude the same to 30mm, which is calculated previously (Width
of the impeller b2) To develop the hub we would revolve the curve which is exported
from CFturbo. Now as we have 7 impeller blades and we need to develop the same equally
in the model so we use the pattern tool and develop the same. And hence we get the final
model of the impeller.
4.15 Representation of velocity
triangles of centrifugal pump along
with the blade

55
4.17 Completion of Impeller Design
4.4 Developed Impeller
4.18 Designed impeller for given duty conditions

56
4.5 Modelling using CFturbo
As in above the hub have already developed with the help of CFturbo. The calculated
values of impeller angle and number of impeller is given manually to generate the
impeller blade. A linear blade profile is taken for designing, the same concept is taken in
above design works. So hence the impeller is generated.
4.19 Impeller design developed by CFturbo
Now modelling volute, it includes the initial setup. Where the type of volute is to be
mentioned ie single or double volute. Here as we are working on single volute, we select
single volute. The earlier mentioned volumetric efficiency comes in role now. The spiral
inlet values are default mentioned as per the given impeller diameter and impeller width.
As we have given impeller diameter as 182mm and impeller width as 30mm and we have
given our flow conditions Cfturbo automatically calculate the best value to satisfy the
curve and generate the values of diameter and width of the spiral.

57
4.20 Input of Volute main dimensions
As per earlier studies we concluded to go with the radial kind of impeller over the
tangential ones because of its better efficiency, so we would select the radial option in
Diffuser.
4.21 Development of the pump diffuser
Generating the volute, CFturbo works with comparing of different input values which we
have already calculated. The inputs are mathematically calculated or in companies they
have specific charts and graphs for their references and standard. As the parameters have

58
been mathematically calculated after the study, the most inputs to the software is already
in hand. Now after generating the volute the software also give a predicted graph of
working.
4.22 Volute design generated by CFturbo
4.6 Predicted performance curve
4.23 Predicted Performance Curve

59
CFturbo enables it total eye by generating a Head vs Discharge performance curve for the
values which we input so as to compare our results of theoretical results and actual ones.
‘The curves generated in CFturbo is based on simple empirical estimations. Deviation
from design point data is possible. In reality experimental performance data of complete
pump stages or CFD results may differ from the value shown in the diagram.’ Here the
final Characteristics are in red colour, which our reference curve for 2900rpm.

60
5. Conclusion
In this project till now a brief study of history of Low flow/ High head pumps are done.
A broad study on the design of pumps was done in which the design of casing and impeller
was separately taken. From studies we saw that a radial volute casing will be appropriate
to be designed to have a more average pressure fluctuation intensity. The effect of the
forward swept vane, radial and backward swept vane were studied and found a radial and
a forward swept blading can achieve a higher head but from it is practically impossible
as the suction cannot be created in the same and the efficiency is very low so we choose
to stick on with Backward facing curve. The design includes points from two design
method berman method and stepanoff way of mathematical designing. As it is advanced
now the methods also includes input from CFturbo. To design a centrifugal pump impeller
a procedure is proposed. The design procedure leads to good results in a lesser time. . The
methodology includes 3 steps mathematically designing (From the calculations we got
the impeller dimensions which include impeller angles, diameters, blade width etc.) and
deriving the impeller dimensions , modelling the hub design using CFturbo and
Completion of model design in Creo. After Modelling, the prediction curve generated by
CFturbo indicates that the pump impeller can work under the given duty conditions.

61
6. References
1) Khin Cho Thin, Design and Performance Analysis of Centrifugal Pump
2) Val S Lobanoff , Robert R Ross , Centrifugal Pumps - Design & Application
3) George Frederick Round, Incompressible Flow Turbomachines: Design,
Selection, Applications, and Theory
4) C.V.S Rajesh , Design of Impeller Blade by Varying Blades and Type of Blades
Using Analytical , ISSN No: 2348-4845
5) Andrzej Wilk, (2010), Hydraulic efficiencies of impeller and pump obtained by
means of theoretical calculations and laboratory measurements for high speed
impeller pump with open-flow impeller with radial blades international journal of
mechanics, Issue 2, Volume 4.
6) Sverre Stefanussen Foslie,2013, Design of Centrifugal Pump for Produced
Water, Norwegian University of Science and Technology
7) Divya Zindania,2016, Design of blade of mixed flow pump impeller using mean
stream line method, 3rd International Conference on Innovations in Automation
and Mechatronics Engineering,ICIAME 2016
8) Shalin P Marathe, Mr.Rishi R Saxena ,Numerical Analysis On The
Performance Characteristics Of The Centrifugal Pump
9) E.C. Bacharoudis , Parametric Study of a Centrifugal Pump Impeller by Varying
the Outlet Blade Angle , The Open Mechanical Engineering Journal, 2008, 2, 75-
83
10) Fan Meng, Effect of two diffuser types of volute on pressure fluctuation in
centrifugal pump under part-load condition, International Symposium on
Transport Phenomena and Dynamics of Rotating MachinerycHawaii, Honolulu
April 10-15, 2016

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