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Design of Centrifugal Pump Impeller/Final Year Project/B.Tech/Mechanical Engg./WBUT/GIMT/2013

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Chapter 1: Executive Summary

Design of centrifugal pump impeller had been conducted by lot of people as a project

in partial fulfilment of their bachelor degree in engineering or technology. But, every design

must have some perspective. Design of an element existing previously is always done to

solve some problems or demands of the former one.

In the 21st century the entire world is facing a situation of energy crisis. Therefore, it

is essential to use energy efficient machines. We should have the sense that energy saved is

energy gained. Keeping these facts in mind we have taken the project to design a centrifugal

pump impeller, which will deliver a rated discharge at the expense of lowest possible power

input while working under a given head. The following figure is a schematic diagram of

centrifugal pump operations and its different components.

Figure 1.1: Typical installation of a centrifugal pump

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The design considers all the factors which are involved in power loss. Disk friction

and impeller vane angles have an important effect on the pump efficiency. They are also

responsible for cavity formation in a centrifugal pump. Vibration at higher speeds is a very

critical problem to be solved. Sometimes this can happen due to improper design, but most of

the times it happens due to fault in fabrication. Thrust load on impeller shaft (for single

suction impeller) is another factor to be solved. The design will take over all these existing

problems regarding smooth pumping.

Figure 1.2: Cross sectional view of a centrifugal pump

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Chapter 2: Problem Analysis

Pumping is one of the most basic and challenging problem for the engineers who

constructed this civilised society, from the day of beginning. Our ancestors solved it by

intelligent application of science and technology and innovation of different pumping

devices. From the historic ages, these devices had seen continuous technological

developments of their own. As a definition we can say, Pump is a device which lifts or

transfers liquids at the expense of input power[1]. Therefore, they are power absorbing

machines. Among all these pumping devices, the centrifugal pump has the widest

application for its suitability to any kind of work and its economical advantages. But,

centrifugal pump needs power input to elevate or deliver liquids. At this point, the problem is

getting generated. At present times, all of our conventional energy sources are getting

collapsed. Entire world is in search of non-conventional energy sources. Our planet is

suffering from useable form of energy. In this situation, considering the case of centrifugal,

our goal should be to try to get our coveted output at the expense of lesser power input. To

solve this problem, we are designing a centrifugal pump with lowest possible energy input to

get the rated discharge.

2.1. Problem ScopeEnergy crisis is the most threatening problem of 21 st century. Every human is bound

to solve this for the sake of his own survival. The solution can be found in different ways.

Either we can search for alternative energy sources or we can make an effort to reduce the

loss of energy in use. The former is already being done by different people. But the latter is

not yet under the lime light. This is the procedure can be applied for any particular product.To find a solution to the energy lacking situation particularly in the zone of pumping, a

centrifugal pump impeller could be designed with lowest possible energy consumption for a

rated discharge. This is also essential for an efficient pump to conserve the kinetic energy of

the fluid and convert it into pressure head. A centrifugal pump, working under a given NPSH,

delivering the rated discharge with lesser energy consumption, is sustainable and

economically advantageous too. The pump can be used in all suitable sectors and wherever it

is used, it will save energy.

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2.2. Technical ReviewThe most modern centrifugal pump of present time is a product of continuous

improvement of more than last 100 years. But the demand of the present situation is energyefficient machines. Technically speaking, efficiency of a system is defined as the ratio of

useful output provided by the system to the amount of input supplied to it. Energy efficient

design gives a practical solution to the feeling of energy conservation.

2.2.1.History of DevelopmentLike other intelligent innovations, centrifugal pump had many giants in its

development so that it is difficult to justly assign the credit to any individual for certain

particular features.

It is said that Johann Jordan designed a crude centrifugal pump in 1680, while Papin

build one in 1703. Euler discussed their theory in 1754. But these early pumps were merely

regarded as curiosities. The first practical centrifugal pump, called the Massachusetts pump,

was built in the United States in 1818. In 1830 a pump having a fairly good efficiency was

built by McCarthy at the dock yards of New York. About 1846 centrifugal pumps began to be

manufactured in England by Appold, Thompson and Gwynne. Appold improved the pump by

the addition of carved vanes in 1849. The addition of diffusion vanes so as to produce the

turbine pump is credited by some to Osborne Reynolds who designed such a pump in 1875.

This pump was not built until 1887 and their commercial manufacture was taken up by

Mather and Platt in 1893. By other the first turbine pump of good design is said to have been

produced by Sulzer, the Swiss engineer, in 1896. About the same time turbine pumps were

built by Byron Jackson, of San Francisco, and others.

The placing of centrifugal pump impellers in series so as to produce the multi-stage

pump was first done by W. H. Johnson in America in 1846. He built a 3-stage pump, but it

appears to have been of little commercial importance. Sulzer is generally given the credit for

being the first to manufacture multi-stage pumps of any importance. In 1894 he built a 3-

stage pump without diffusion vanes and in 1896 he constructed a 4-stage turbine pump. The

latter had a capacity of 5000 G.P.M. under a head of 460 ft.

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Although the centrifugal pump has been in existence for a considerable period, it is

only within the last 100 years that it has been widely used or rapidly improved. The reason

for this is that the centrifugal pump is a relatively high-speed machine and until there was no

form of motive power well suited to it. In the days of slow-speed steam engine the

reciprocating pump was better adapted to the conditions. But with the introduction of the

steam turbine and electric motor the conditions were reversed. For such sources of motive

power the reciprocating pump is not as well adapted as the centrifugal pump.

Before 100 years, source of motive power was a problem, and today the primary

problem is to conserve this motive power by reducing its losses. Lot of experiments are being

done to reach this goal. Greg Case et al proposing design with reduced vibration, rotor

rubbing, overstress, cyclic fatigue etc. [6]. E. C. Bacharoudis et al experimented with varying

blade angles for the sake of increase in efficiency in 2008 [7]. Mike Swanbom et al had given

primary importance to energy conservation in design and manufacturing of centrifugal pump

in their project in 2008 [8]. K. W. Cheah et al suggested that the unsteady flow generates in

the impeller passage in off design flow rate by their numerical flow simulation experiments in

2007 [9]. Weidong Zhou et al concluded that twisted impeller blades are more efficient than

that of the straight impeller blades by CFD analysis in 2003 [10]. Miguel Asuaje et al

experimented on the radial thrust on impeller shaft for different speeds, impellers and volutesusing CFX code in 2005 [11]. This development is continuously going on with the target of

most energy efficient centrifugal pump.

2.2.2.Economical AspectsFor any design, there is a very basic and primary requirement that the design must be

practical and economically feasible. In case of design of a centrifugal pump, we must have to

consider the cost of pumping in mind. The total cost of pumping is the sum of the fixed

charges and operating expenses. The former consists of interest on the capital cost, insurance,

taxes, depreciation and administration. The latter item includes labour, fuel or electric

current, supplies, repairs and other similar items.

The capital cost covers the cost of the pump, the motor or prime mover and possibly

the building, pipe lines and such other equipments that the pumping makes essential.

The total annual cost consists of fixed charges and operating costs for a period of 1

year. The cost of pumping per water horse-power or per 1000 gal per min or any similar unit

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is the total annual cost divided by the total capacity of the pump, meaning by total capacity

the water horse-power or the number of 1000 gal per min, or other units of which the pump is

capable. It will be a minimum when the pump is not operated at all as it will then consist of

the fixed charges only. It will be a maximum when the pump is operated continuously as that

will cause the operating expenses to be a maximum.

For a motor driven pumping unit the total annual cost of pumping is:

Where,

C = total annual cost

G = total number of gallons pumped per year

h = head in feet

S = cost per million B.t.u. supplied to the motor

D = duty in ft. ib. per million B.t.u.

L = cost of labour and similar items

F = total investment

i = interest rate on investment

d = rate of depreciation

t = taxes, insurance etc.

M = administration and similar items

Since 1000000 B.t.u. = 778000000 ft. lb., we may write Duty = 778000000 Pump

efficiency Motor efficiency. Therefore, the first part of the above equation becomes,

If, K = cost of power per k.w.hr.

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Now the above equation shows that, if the cost of power is high, it may be economical

to pay a high price for a high-duty pumping engine. But on the other hand a less expensive

centrifugal pump may often effect a saving even though its duty should be somewhat less.

The equation would also show that for intermittent service a cheap pump was desirable even

though it might be inefficient. But for constant service a high-duty pump is better even

though its first cost may be considerably higher.

2.3. Design Solution RequirementsWe are considering such factors which are responsible for power loss. In the design

we are trying to solve the problem of power loss in centrifugal pump. The designed pumpimpeller will have higher overall efficiency and will be energy efficient too. The impeller

shaft needed to experience minimum thrust under off-design conditions also. Improved blade

angles can increase the efficiency. Twisted blades are more preferable than that of the flat

blades. The shape of outlet flange should be relatively best suitable with the other parameters.

Vibration at higher working speeds is another factor to be solved in the design.

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Chapter 3: Review of Existing

Technology

It is very important to have a very broad and sharp knowledge about the theory and

existing technology of centrifugal pump impeller before going to design it. Design of a pump

impeller is a deep and detailed process to be followed. It requires a handful experience about

the technology and the experience grows from the study of the previously existing technology

of the element.

3.1. Categorization of Impeller according to ConstructionFrom basic theories there are three types of impellers according to their fabrication

and construction. Open impellers are intended to be used to pump fluids containing

particulates. Those impellers do not have shrouds on the front or the rare of the impeller,

allowing them to wear particulate that might clog other impellers along the stationary walls of

the front casing and rare cover. As the impeller rotates, the particulate is dragged along the

stationary walls causing it to wear down so it can pass through the impeller. An added benefit

of the open impellers is the lack of surface area on the impeller shrouds to allow the axial

loads to build on this drastically reduces the axial load on the pump, improving bearing life.

Some of the disadvantages of this style of impeller are the added leakage areas around the

front and rear of the impeller vanes. This often causes efficiencies of these impellers to be

lower than impellers of similar specific speed but different styles. Another drawback is the

thickness of the vanes. The vanes must generally be thicker due to the lack of shroud to help

support them. Because of the thicker vanes the impeller must necessarily have fewer vanes

and thus have less control over the fluid.

Closed impellers have shrouds on both sides of the impeller providing less leakage

and better control of the fluid flow. These impellers can be prone to clogging from particulate

in the pump. Generally the impeller efficiency is slightly better than open and semi-open

impellers due to the lower leakage rates around the impeller vanes. They are also generally

thought to have better control of the fluid direction because the added shrouds rotate with the

vanes, preventing the increased drag on the fluid imposed by the stationary walls. While the

support provided by the two shrouds on this style of impellers mechanically allows for

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thinner vanes, the vane number and thickness are somewhat restricted in cast impellers by the

need for wall sections of consistent thickness and the need to be able to remove the impeller

cores.

The design and variants of semi-open impellers are probably the most common type

of impeller in the world. As the name implies, the impeller is a hybrid of the open and closed

design. The design has one shroud commonly on the back of the impeller vanes. They share

many of the advantages of both the open and closed impellers, such as self-cleaning impeller

passages, excellent control of the fluid in the vane passages and the ability to use fairly thin

vanes. The absence of one of the shrouds often allows for more advanced vane profiles to be

used including the addition of splitter vanes that would be difficult to use on closed impeller

designs.

One of the most effective techniques used to balance hydrodynamic axial loads is the

use of double suction impellers. By using virtually the same impeller profile on both sides of

the impeller, the developed forces are nearly equal. These impellers are inherently balanced

and develop very low axial loads. One side of the impeller is generally designed to develop a

small axial force to load the thrust bearing slightly and prevent the impeller from hunting

back and forth during operation. A variant of the double impeller concept, used in multistage

designs, is the mounting of an even number of impellers on a single shaft so that they oppose

each other. The axial loads are then equalised on the rotor as a whole.

3.2. Impellers according to Specific Speed and FlowIf very large volume flow rates are needed, or if the velocity of flow is limited in the

entrance for reasons of the suction behaviour, radial flow pumps are frequently implemented

in a multi-flow way. Thereby two impellers with same dimensions deliver in a common

housing. With same delivery head the two flow rates are added together.

Since the maximum delivery head of an impeller is fixed by the pressure factor in

dependence of the design and upward the number of revolutions limited by firmness reasons,

for the achievement of large delivery heads several pump stages are connected in series. The

delivery heads of the single stages are added with same flow rate.

1. Low rapidity (specific speed = 10-30): Radial-flow impeller with simply curved blades.Pumps with low delivered flow and large delivery head.

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2. Medium rapidity (specific speed = 30-50): Impeller with radial discharge and doublecurved blades. Pumps with a middle delivered flow and middle delivery head.

3. Helicoidic impeller(specific speed = 50-80): Impeller with double curved blades. Pumpswith larger as middle delivered flow and smaller than middle delivery head.

4. Diagonal impeller with high rapidity (specific speed = 80-135) with double curvedblades. Pumps with high delivered flow and a low delivery head.

Figure 3.1: Impeller forms [7]

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5. Propeller impellerwith highest rapidity (specific speed = 135-330) and rotor blades inthe form of wings. Pumps with highest delivered flow and lowest delivery head.

3.3. A Case StudyA large double suction single stage pump, with an impeller diameter of 2.5 feet and a

running speed of 1500 rpm, was designed with close impeller vane/volute tongue clearance to

reach an aggressive efficiency level in a facility where energy was at a premium. During

installation, it was found that vibration levels got as high as the operating clearances in the

wearing rings (0.6 mm diametral), with the primary component at running speed. There was

no possibility of a resonance in this pump since both the shaft and the bearing housing natural

frequencies were above the 1X and 2X excitations and the 3X excitation due to suction flow

asymmetry, which is common in this style pump. The vane passes frequency of 4200 cpm

was far removed from the shaft first and second non-critically damped natural frequencies of

2850 and 19000 cpm respectively.

The reason for the high vibration was found to be 35 mils of misalignment at the

coupling due to the hydraulic loads on the pump discharge flange being far in excess of API

610 (1995) levels. The 30 inch discharge had a piping expansion joint at the flange, with no

tie-bars in place across the flange to carry the resulting thrust. After removal of the piping

forces through a grounded bulkhead bolted to the discharge flange, the pumps large 1X and

2X vibration levels were reduced to acceptable values per API 610 (1995).

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Chapter 4: Design Considerations

Owing to the inherent defects of the theory of the centrifugal pump, any method of

design must involve the use of empirical factors determined by experience. This is true of

practically all engineering work, but in some cases the factors are reasonably constant and are

known to vary as definite functions of other quantities. The determination of numerical

values for these factors is not so certain in the case of centrifugal pump.

The design of a centrifugal pump impeller is ultimately based upon the performances

of other impellers. The theory indicates what would be the general effect of altering certain

dimensions. Hence successful design consists of modifying or changing the design of

impellers which have been tested out rather than creation of entirely new patterns. After a

number of impellers of different types have been constructed and their performances properly

recorded, the designer will then be in a position to develop new designs and to predict results

with some assurance.

As illustration, suppose that it is required to design a centrifugal pump of a certain

capacity under a given head. The number of stages and the r.p.m. might be arbitrarily

assumed for non-technical reasons. But more scientifically they might be determined by

considering the factors affecting efficiency, due regard being shown commercial conditions at

the same time. But it is seen that the experimental data will be necessary before this much can

be done.

Having now the values of speed in r.p.m., discharge rate and head developed per

stage, the desired form of impeller characteristic may be selected. That is we may decide

whether a rising, a flat or a steep characteristic is more suitable for the particular work this

pump is to do. Having chosen this, it will be necessary to select the angle of the impeller vane

at exit. Again experience will be necessary for this to be done, since it cannot be determined

by the solution of any mathematical equation. The theory, however, indicates that the smaller

the outlet vane angle, the steeper the characteristic. Experience also points to the fact that the

fewer the number of vanes the steeper the characteristic. Also the theory will show that the

angle of the diffusion vanes or the area of the volute case, if vanes are lacking, has an effect

upon this. The larger the diffusion vane angle or the larger the case of a volute pump, the

higher the discharge and the lower the head at which the maximum efficiency will be found.Only by the study of the performances of other pumps for which these quantities are known

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can the proper values of outlet vane angle, diffusion vane angle and area of stream normal to

the direction of flow be chosen.

The next step is the selection of the factors speed ratio andflow ratio, whose values

may normally range from 0.95 to 1.25 and from 0.1 to 1.25 respectively. The steeper is the

characteristic; the larger is the value of speed ratio. Therefore speed ratio is some function of

the other relative quantities, as is flow ratio also. If the theory were capable of exact

application, we might compute values of speed ratio and flow ratio from the equation given

below, but even those equations involve the selection of a factor k which is a matter of

experience again. We shall, therefore, have to choose a value for speed ratio according to our

best judgement or according to values obtained by test upon a pump similar in design to the

one we are attempting. The value of flow ratio may be determined in the same manner as

speed ratio.

As a check upon the rationality of our values of outlet vane angle, speed ratio and

flow ratio we may substitute them in equation,

and see if the value of the expression is in accordance with the customary values for the line

of pumps whose data we may have. If our theory were exact the value would be the true

hydraulic efficiency, a value for which might reasonably be estimated. As our theory is

defective, that is, since the computed value of head imparted to the water by the impeller is

higher than the true value, this value will not be any definite physical quantity and is called

simply manometric coefficient. Or we might assume a value of the manometric

coefficient and compute flow ratio (Kf) from the above equation.

To design a centrifugal pump, some empirical formulations and considerations must

be followed. Following are some empirical consideration in the design of centrifugal pump:

i. Speed ratio (Ku) refers to the ratio of peripheral speed () at the impeller tip to thetheoretical jet velocity corresponding to manometric head.

The value ofKu varies from 0.95 to 1.25.

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ii. Flow ratio (Kf) refers to the ratio of flow velocity (Vf2) at exit to the theoretical jetvelocity corresponding to manometric head.

The value ofKfvaries from 0.1 to 0.25.

iii. Knowledge of peripheral speed () helps to compute diameterD2 at the outerperiphery of the impeller.

iv. Usually diameterD1 at the inner periphery is kept in the range:

v. The selection of outlet vane angle (2) depends on the type of head capacity

characteristics desired. For optimum efficiency, usually a value of about 25 is taken

for all specific speeds.

vi. The inlet vane angle is selected so that inlet absolute velocity may be radial. Theradius of curvature of vanes is selected depending on the inlet and outlet blade angles,

so that a smooth, separation free flow is obtained in the impeller passage.

vii. The number of vanes in an impeller depends on the pump size, the speed ratio, thevane load and the outlet blade angle. With low values of outlet blade angle, usually

six or eight vanes are adopted.

The main dimensions which affect the hydraulic features of the pump are thus

determined. The equation may be raised as to what assurance we have that the maximum

efficiency will be attained under the conditions of speed, head and discharge for which these

computations were made. The only explanation is that the values ofspeed ratio and flow

ratio, upon which the computation hinge, were selected according to values obtained with

previous pumps for their point of maximum efficiency. Furthermore all dimensions and

angles computed were determined upon the supposition that the flow specified would be the

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normal flow and provisions were made to maximize all the losses at the flow. However the

actual point of maximum gross efficiency is affected by the mechanical losses as well as the

hydraulic losses. It might be necessary to allow for this if it were not for the fact that it has

also entered into the previous pumps for which our values of speed ratio and flow ratio were

experimentally determined.

4.1. Losses and EfficiencyThe kinds of loss of centrifugal pumps can be differentiated in:

Internal losses:

Hydraulic losses or blade losses by friction, variations of the effective area or changesof direction.

Losses of quantity at the sealing places between impeller and housing, at the rotaryshaft seals and sometimes at the balance piston.

Wheel friction losses by friction at the external walls of the wheel.

External or mechanical losses:

Sliding surface losses by bearing friction or seal friction. Air friction at the clutches. Energy consumption of directly propelled auxiliary machines.

One can directly determine the overall efficiency and also the internal efficiency by

attempt, but as for the blade efficiency and the hydraulic efficiency this is not possible. It

must be computed from overall efficiency or internal efficiencyby excluding the losses,

which are not pressure losses.

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4.2. Outlet Blade AngleThe angle of outlet 2* can theoretically be selected freely within a wide range. An angle

2*>90 leads to backwards curved blades. 2*=90 means radially ending blades and2*90 for getting a lower c2. In addition, a large angle 2* has the disadvantages that it

requires with same delivery head a larger circumferential speed and so it causes larger wheel

friction losses.

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Because of the larger difference of pressure between entrance and exit of the impeller,

larger gap leakages are caused. However, these disadvantages cannot cover the crucially

better hydraulic efficiency. Therefore in centrifugal pumps only backwards curved blades

with angles of outlet 2*=140-160 are used.

Figure 4.1: Outlet blade angle and velocity triangle

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Chapter 5: Detailed Design

Design of a centrifugal pump impeller intends to balance the following

considerations:

Capital cost (initial) Performance Reliability Operational costs (life-cycle)

For many organizations only the first two items are considered in the purchase

decision, with initial cost being of primary importance. Focus on initial costs and

performance often leads to high life-cycle costs and reduce reliability.

The initial cost of a system can often be quite low when compared to the operational

costs of the equipment over time. Factors such as the cost of power, repair costs and lost

production are less commonly considered in most purchasing processes. Making sound,

informed purchasing decisions during the front end of the purchasing process can often

improve performance, increase reliability, reduce life-cycle costs and occasionally reduce

initial purchase costs.

The client will usually specify the desired head and pump capacity. The type and

speed of the driver may also be specified. Speed is governed by considerations of cost and

efficiency as well as drivers available to the client. Given these parameters, the task of the

engineer is to minimize cost. Here in the design of centrifugal pump impeller we have

considered the values of head, discharge and operating speed from experience of practical

applications.

Specified conditions:

Head (H) = 60 ft

Discharge (Q) = 2500 gpm

Speed (N) = 1500 rpm

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1. Specific Speed:First check the specific speed (Ns),

Here, Ks = 0.0174

We find the value of specific speed 43, which is permissible. Now, we have to calculate the

corresponding shape number.

2. Shape Number:

Shape number=

Figure 5.1: Relationship between shape number and gross efficiency

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According to graph (fig. 5.1) the value is within safe limit.

3. Quantity Flow rate:

Discharge in cubic feet per second (q)=

1 cubic feet per second = 448 gallon per minute4. Water Horsepower:

Output power (Pw) =

5. Shaft Power:From graph (fig. 5.2) we have taken the value of gross efficiency () is 80%.

Figure 5.2: Relationship between efficiency and specific speed

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Now, shaft power or power input =

6. Values from Experience:Now, calculated value from graph for double suction impeller,

Width ratio () = 0.175Speed ratio () = 1.085Diameter ratio (

) = 0.585Flow ratio (

) = 0.175

7. Impeller Diameter and Impeller Width:

v2 = outer rim velocity =

Figure 5.3: Relation between shape number, output and efficiency

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d2 = outer diameter of the impeller

Inner diameter (d1) = d2 diameter ratio = Outer width (b2) = Inner width (b1) =

8. Shaft Diameter:

Now, the required shaft torque is,

Assuming a shear stress of 4000 psi for the shaft (material - Steel SAE 1045), shaft diameter

(Ds) = To account for the unknown bending moment and critical speed increase the shaft diameter to

1.8 inch.

9. Hub Diameter and Length of Hub:The hub diameter Dh is generally taken in the range of

to inch and larger than Ds.Let, Dh = 1.8 + 0.5 = 2.3 inch and length of the hub lh =

10.Suction line velocity and the Diameter of Suction Flange:Now, assume a velocity of 10 ft/sec at the suction flange, thus,

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11.Diameter of the Impeller Eye:Assume the velocity at the eye of the impeller is 11 ft/sec.

For a double suction pump, assume that the leakage will not exceed 2%. Dividing the total,

12.Assumption of Blade Number and Thickness from Experience:Now, we are assuming impeller blade number 8 and thickness 5 mm. As the blade number

varies from 6 to 12 and the blade thickness varies from 4 mm to 8 mm.

13.Hydraulic Efficiency:Now,

Hydraulic efficiency () = 89%14.Outlet Velocity and Outlet Blade Angle:

Now, we have to calculate the actual outlet tangential velocity, .

Now, the theoretical outlet tangential velocity component is, .

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Where, is a factor can be taken as about 3 for low specific speed and as about 5 for highspecific speed.

Now, radial flow component at outlet is

.

Flow ratio () =

Now, assuming outlet blade angle 2and inlet blade angle 1,

Figure 5.4: Relation between impeller radius, velocity and vane angle

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15.Inlet Velocity and Inlet Blade Angle:And,

= radial component at inlet= absolute velocity at inlet must be greater than .

Figure 5.5: Inlet velocity triangle

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16.Velocity Triangle:

The inertia of the rotating fluid causes a circulatory flow opposite to the direction of rotation

of the impeller. This flow, superimposed on the outward flow, results in the fluid leaving the

impeller at an angle less than that calculated from angular momentum theory. Thus 2 must

be decreased and , therefore, the absolute angle, 2 , increased. The effect of circulatory flow

is to reduce V2 and the theoretical head.

Now,

And,

From the triangle we can calculate and .

Figure 5.6: Outlet velocity triangle

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17.Number of Vanes:Now we have to check the number of vanes.

18.Impeller Width:

Now, correction to impeller width, in order to allow a margin for wear, the leakage loss q1

may be taken as 2% of the net flow q.

We know,

Corrected inlet width = Corrected outlet width =

Figure 5.7: Layout of the designed pump impeller

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Chapter 6: Application and Future

Scope

In particular radial-flow pumps are used for liquid-delivering in a dominant number of

constructions. Beside water every other liquid is applicable as delivery medium. In particular

oil, but in addition, aggressive liquids or liquid solid mixtures can be delivered with

centrifugal pumps. The exact application fields or plants for centrifugal pump are mentioned

in details below;

Water Management (water supply, irrigation, drainage, sewage disposal) Plant: Heredifferent centrifugal pumps are used in forms of celler drainage pump, booster pump,

sprinkling pump, sewage pump etc.

Power Plants: This is a place for broad application of centrifugal pumps in forms ofcirculation pump, reactor pump, boiler feed pump, condensate pump, storage pump

etc.

Shipbuilding: Here centrifugal pumps are very widely used in different applications,in forms of bilge pump, ballast pump, dock pump, ship pump, fuel pump etc.

Other intended purposes: Except the mentioned sectors, centrifugal pump has a broadapplication in other specific fields like fire-fighting pump (in fire brigade), water

circulation pump (in water treatment plant) etc.

Our designed pump impeller can serve any of the above mentioned purposes as per pumping

requirement.

The design can be rechecked for further betterment and advancement. It can be

studied if there is any problem in practical application of the design. Several fluid analyses

can be done on this by using different software. The design can be improved for higher

efficiency. Blade angles can be further developed for better performance under the given

conditions. The performance of the design at off design conditions could be checked. The

manufacturability of this design can be tested by rapid prototyping. The stability of the

impeller at the running conditions, different stresses, existence of any unbalanced load,

signature of vibration, formation of cavitation under specified situation are subjects of

experiment and further broad study. The result of this project is definitely creating a plenty of

issues that can be taken as a lead of further research and development.

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Chapter 7: Conclusion

Centrifugal pump and specifically its impeller is a very common device for a present

day child of either urban or rural area in our country. It looks very simple and common in

design and construction. There nothing looks special in this device that can make it

attractively charming. During four years of our B. Tech. course, we have studied different

theories about different technologies. The application fields and results were also studied by

us in details. But, before going through this project, we never could imagine how hard to try

to implement a theory in practical application. We didnt know the importance of studying

any simple looking element in details; we just studied it for exams. I was unknown to us that

how much knowledge it needs to design something. But, finally we got success in designing a

centrifugal pump impeller.

This project is a successful design of a centrifugal pump impeller. We designed a

double suction impeller to solve the problem of unbalanced axial force on the shaft. We

studied and understood the theory of centrifugal pump. More importantly we learnt the

limitations of the theory. A large number of previous works were reviewed. Different

relationship curves were studied and some values were retrieved from those curves and

charts. A little hardship were faced in selection of the outlet vane angle and pump

characteristics.

Completion of this project was impossible without the help of our project guide, our

other teachers and some of our friends. This project has left a considerable space for further

study and research. We hope it will help others in future to study about design of centrifugal

pump impeller.

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References

Books:

[1] Kumar. D. S., Fluid Mechanics and Fluid Power Engineering, Katson Books, 2010.

[2] Daugherty. R. L., Centrifugal Pumps, McGraw-Hill Book Company, 1915.

[3] Evolution of the Turbine Pump, Proc. Inst. Of Mech. Eng., 1912.

[4] Webber. W. O., Trans. Amer. Soc. Of Mech. Eng., 1905.

[5] Greene, Pumping Machinery.

Journals and Proceedings:

[6] Neff. M., Bearingless Centrifugal Pump for Highly Pure Chemicals, 8th

International

Symposium on Magnetic Bearing, August 2002, Mito, Japan, P: 283-288.

[7] Case. G., Centrifugal Pump Mechanical Design, Analysis and Testing, Proceedings of

the 18th

International Pump Users Symposium, P: 119-133.

[8] Bacharoudis. E. C., Parametric Study of a Centrifugal Pump Impeller by Varying the

Outlet Blade Angle, The Open Mechanical Engineering Journal, 2008, 2, P: 75-83.

[9] Swanbom. M., Centrifugal Pump Design, Fabrication and Characterization: A Project-

driven Freshman Experience, American Society for Engineering Education, 2008.

[10] Cheah. K. W., Numerical Flow Simulation in a Centrifugal Pump at Design and Off-

design Conditions, International Journal of Rotating Machinery, Vol. 2007, doi:

10.1155/2007/83641.

[11] Zhou. W., Investigation of Flow Through Centrifugal Pump Impellers using

Computational Fluid Dynamics, International Journal of Rotating Machinery, 2003, P: 49-61.

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[12] Asuaje. M., Numerical Modelization of the Flow in Centrifugal Pump: Volute Influence

in Velocity and Pressure Fields, International Journal of Rotating Machinary, 2005:3, P: 244-

255.

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