1

DESIGN AND ANALYSIS OF REVERSIBLE PUMP TURBINE

An undergraduate thesis submitted in partial fulfillment of the requirements

for the award of the degree of

Bachelor of Technology in Mechanical Engineering

By

Arun Kumar Yadav (12/ME/24)

Mustafizur Rahaman (12/ME/25)

Souvik Das (12/ME/45)

Arvind Kumar Gupta (12/ME/47)

Abhishek Paul (12/ME/48)

Aman Verma (12/ME/101)

Nihal Kumar (12/ME/103)

Under the supervision of:

Dr. SUBHAS CHANDRA RANA

Assistant Professor

DEPARTMENT OF MECHANICAL ENGINEERING

NATIONAL INSTITUTE OF TECHNOLOGY DURGAPUR

DURGAPUR-713209

WEST BENGAL, INDIA

2016

2

CERTIFICATE OF APPROVAL

The following thesis is hereby approved as a creditable

study of an Engineering subject, carried out and presented

in a manner satisfactory to warrant its acceptance as a

prerequisite to the degree for which it has been submitted.

It is to be understood that this approval of the undersigned

do not necessarily endorse or approve any statement made,

opinion expressed or conclusion drawn therein, but approve

the thesis only for the purpose for which it has been

submitted.

SIGNATURE OF SUPERVISOR THESIS EXAMINER

3

ACKNOWLEDGEMENT:

We take this opportunity to express our profound gratitude and deep regards

to my supervisor Dr Subhas Chandra Rana for his exemplary guidance,

monitoring and constant encouragement throughout the duration of this

project. His valuable suggestions were of immense help throughout our

research work. The blessings, help and guidance given by him time to time shall

carry us a long way in the journey of life on which we are about to embark.

4

CONTENTS

TOPICS PAGE NUMBER

1. Abstract ………………………………………………………………………………………………………………………..5

2. Brief Introduction…………………………………………………………………………………………………….…….5

3. Hydraulic Design Issues………………………………………………………………………………………………….7

3.1. Euler’s Equation………………………………………………………………………………………………………8

3.2. Stability of pumps……………………………………………………………………………..…….…………….14

4. Design process…………………………………………………………………………………………………………….19 4.1. Maximum Efficiency………………………………………………………………………………….….……….21

4.2. Cavitation……………………………………………..……………………………………………………...…..….23

4.3. Net positive suction head(NPSH)……………………….………………………………………..….….…25

4.4. Free Vortex Theory………………………………………………………………………………………………..28

5. Design parameters and output……………………………………………………………………………………..29

6. Modelling……………………………………………………………………………………………………………………..34

6.1. Assembly of RPT Housing……………………………………………………………………………………….35

7. CFD ……………………………………………………………………………………………………………………………...37

7.1. CFD Analysis…………………………………………………………………………………………………………..39

7.2. Results of CFD Analysis…………………………………………………………………………………………..41

8. Conclusion…………………………………………………………………………………………………………………….41

9. References……………………………………………………………………………………………………………………42

10. Websites……………………………………………………………………………………………………………………….42

5

1. ABSTRACT:

A reversible pump turbine is a machine that can operate in three modes of

operation i.e. in pumping mode, in turbine mode and in phase compensating

mode (idle speed). Most hydropower plants in are run-off type, which cannot

supply the designed amount of energy during dry season and peak demand

period resulting in energy crisis. Therefore, pumped storage plants can be an

ideal solution to meet the current energy needs of the country. Most of these

storage plants make use of a single unit acting both as turbine as well as pump;

hence aptly called Reversible Pump Turbine (RPT). This paper explains the new

application concept for use of such RPTs as auxiliary unit to supplement the

main power unit in hydropower plants.

The design is performed in two steps. The first step is an analytical design,

which gives an initial geometry of the RPT runner. The process resulted in a

runner of larger size than a normal Francis runner for same parameters since it

has to work as pump as well. The next step involves CFD analysis under which

the simulation of the RPT in turbine mode is done.

2. BRIEF INTRODUCTION:

A Pumped-storage plant stores energy by pumping water from a lower

reservoir at off peak hours of electric demand by means of surplus power into a

high level reservoir, in order to utilize the stored energy at periods when it is

6

most needed.It is probably the best way to compensate for the gap between

produced and consumed power. These pumped storage system use a single

pump/turbine unit i.e. Reversible Pump Turbine to efficiently and economically

store electrical energy during periods of low demand to meet peak load

demands.As far as our scope of work is concerned we have immaculately tried

to design the differents parts of the RPT using SOLIDWORKS,CATIA for the

modelling part and ANSYS for consequent simulation and analysis.

Schematic diagram of pumped storage plant

7

2. HYDRAULIC DESIGN ISSUES:

An RPT acts as a water passageway with a shiftable body component

selectively displaceable to achieve alternatively, either an energy generation or

an energy accumulation mode The design of a pump turbine is similar to the

design of a high head Francis turbine. However, the factors that need to be

considered while designing an RPT are as follows:

A. Pump head should be higher than in turbine mode due to loss in the

waterways.

8

3.1 EULER’S EQUATION –

The Euler's equation for steady flow of an ideal fluid along a streamline is a

relation between the velocity, pressure and density of a moving fluid. It is based

on the Newton's Second Law of Motion. The integration of the equation gives

Bernoulli's equation in the form of energy per unit weight of the following fluid.

It is based on the following assumptions:

The fluid is non-viscous (i,e., the frictional losses are zero).

The fluid is homogeneous and incompressible (i.e., mass density of the

fluid is constant).

The flow is continuous, steady and along the streamline.

The velocity of the flow is uniform over the section.

No energy or force (except gravity and pressure forces) is involved in the

flow.

9

DERIVATION –

Let us consider a steady flow of an ideal fluid along a streamline and small

element AB of the flowing fluid as shown in figure.

Let,

dA = Cross-sectional area of the fluid element

ds = Length of the fluid element

dW = Weight of the fluid element

P = Pressure on the element at A

P+dP = Pressure on the element at B

v = velocity of the fluid element

10

We know that the external forces tending to accelerate the fluid element in the

direction of the streamline

(1) We also know that the weight of the fluid element,

From the geometry of the figure, we find that the component of the weight of

the fluid element in the direction of flow,

11

(2)

Mass of the fluid element =

We see that the acceleration of the fluid element

(3) Now, as per Newton's second law of motion, we know that Force = Mass

*Acceleration

Dividing both sides by

or,

(4) This is the required Euler's equation for motion as in the form of a

differential equation. Integrating the above equation,

12

or in other words,

, which proves the Bernoulli's equation.

Below Figures depict the velocity triangles at inlet and outlet ,along with the

cross section of the turbine blade:

13

NOMENCLATURE –

Thus, according to the Euler’s equation for turbine and pump we have,

EFFICIENCY OF TURBINE

EFFICIENCY OF PUMP

Now, we assume frictionless flow i.e. Hp =Ht and no swirl at the outlet during

turbine mode i.e. Cu2 =0 . The reduced speed for turbine and pump can then, be

calculated as

14

U*1t = Flow speed ratio of turbine U

*1p = Flow speed ratio of pump

Assuming speed is the same in both pump and turbine i.e. U*1t=U

*1p

This means,

Therefore, the pump turbine has to be designed for a higher head than the

theoretical head such that . This ensures that the pump head should

be greater than the turbine head. But, the function of the RPT defines it using as

both pump and turbine which means the design head should be compatible with

both modes ensuring highest efficiency delivered. The solution can be met by

having fewer and longer blades than the traditional Francis runner.

B . The pump should be stable pump and not oscillating.

3.2 Stability of pumps – A basic idea when it comes to stability, is to look at

what happens if a system experiences a small perturbation. If the system

manages to come back to it’s starting point, the system is stable. It is a condition

of pump where the static head or the differential head continuously falls with

15

increasing flow rate.Practically it suggests that the pump will not oscillate for

any two combinations of head flow.Suppose water is to be pumped from one

level to higher level,there will be no vibrations induced during the

operation.The flow rate will not modulate and the pipeline doesn’t vibrate.

stable head-flow

Whereas, Operation of an unstable pump may lead to either exponential or

oscillatory unstable behaviour of both pressure and volume flow of the water in

the conduit system. This is of course is unwanted, and fulfilling the stability

criteria is desirable.

For unstable head-flow characteristic the differential head - h - rises to a

maximum and then progressively falls with increasing flow rate - q.

16

For stability of the pump, the inlet angle β1 should be less than 90 degrees. such

that for higher value for flow (Q) the power stabilizes.With β (outlet

angle)larger than 90 degrees, the slope will be positive.When the value of U1 is

increased the (inlet angle) β1 gets smaller ,which is preferable considering

pump characteristics.

The above figure shows the graph plot between “Power” on Y-axis Vs

“Discharge” on X-axis.When β(outlet angle)=0 degrees, it indicates that it is a

radial curved blade.Simultaneously, when β>90 degrees,it indicates that it is a

forward curved blade.Similarly,when β<90 degrees,it shows that it is a

backward curved blade.

17

C. The system should deliver high efficiency in both modes.

The higher efficiency in both modes can be achieved by designing the system

for best efficiency head rather than the required head at the site. Adjustment of

the best efficiency head for pump is possible by varying its rotational speed as

head produced is dependent on the rpm of the impeller in the pump mode. High

specific speed pumps have relatively steep head-discharge curves so that they

are able to operate at wide variation of head. However, the allowed variation of

operating head range is narrow for RPT than in case of real turbine such that the

head range is only allowed between 65% to 125% of the design head.

MATHEMICAL PROOF - A centrifugal pump and a Francis turbine is in fact

the same kind of machine, both obeying the Euler equation. Assuming rotation

free inlet, the theoretical head of a pump may be expressed by the Euler

equation:

where H – Theoretical head u2 – outlet water velocity β – outlet angle c2 –

radial velocity at outlet B2 – width of impeller D2 – diameter of impeller Q –

volumetric discharge

18

At a given speed of rotation, i.e. for a given u2, The QH-characteristics will be

ascending if the outlet angle β>90o and descending if β<90o ,i.e forward or

backward leant runner blades.In order to get stable pump

characteristics,centrifugal pumps must have backward leant runner blades.This

is the case of RPT,when running as a pump.However,when the speed of rotation

changes direction to turbine mode of operation,the pump effect will be even

stronger because of forward leant blades.

Theoretical Q-H characteristics of centrifugal pumps.They are dependent

on outlet angle β2.

D. The relative velocity should remain almost constant.

Acceleration through the runner is undesired since it turns into deceleration

when shifting operation mode. A decelerated flow is more vulnerable to

secondary flow effects and separation. Thus, a small difference between the

magnitudes of relative velocity at the inlet and the outlet should be the goal of

19

the design, which can be achieved by increasing U1. To increase U1, the value

of U1 reduced should be chosen near to 1. This value gives steep pump

characteristics and just a small increase of W1 through the runner channel.

3. DESIGN PROCESS -

The first choice in the design process is to find a suitable existing plant site that

can be used as a starting point. The power plant should have a turbine with a

speed number between 0.27 and 0.35.The available head and flow is 270 m and

4 m3/s respectively (From input parameters)

\

A.Main Dimensions

4.1 MAXIMUM PUMP EFFICIENCY –

A pump does not completely convert kinetic energy to pressure energy. Some of

the energy is always internal or external lost.

Internal losses are

hydraulic - like disk friction in the impeller, loss due to rapid change in

flow direction and change in velocities throughout the pump

volumetric losses - internal recirculation due to wear in rings and bushes

20

External losses

mechanical losses - friction in seals and bearings

The efficiency of a pump at it's design point is normally it's maximum and

called

BEP - Best Efficiency Point

It is normal to operate a pump in other positions than BEP, but the efficiency

will always be lower than in BEP.

The dimensioning of the outlet starts with assuming no rotational speed at best

efficiency point (BEP) i.e Cu2 (Tangential component of absolute velocity

at outlet) = 0. In addition, the values for outlet angle, β2, and (peripheral speed

at outlet) U2, are chosen from empirical data:

13° < β2 < 22° Lowest value for highest head

35 m/s < U2 < 42 m/s Highest value for highest head

21

The outlet diameter and the speed are found by reorganizing the expressions for

flow rate and rotational speed, respectively. Cm2 ( Radial component of

absolute velocity at outlet) is obtained from the known geometry in the velocity

triangles. The number of poles, Z, in the generator depends on the rotational

speed and net frequency. With a net frequency of 50 Hz, the number of poles is

calculated using equation-

Since the number of poles is an integer, the value obtained from equations must

be round up. With the correct number of poles, equations is used to find the

synchronous speed, ncorrected, which in turn, is again used to calculate the

corrected diameter at the outlet, D2corrected .

4.2 CAVITATION:

Cavitation is the formation of vapour cavities in a liquid – i.e. small liquid-free

zones ("bubbles" or "voids") – that are the consequence of forces acting upon

the liquid. It usually occurs when a liquid is subjected to rapid changes

of pressure that cause the formation of cavities where the pressure is relatively

22

low. When subjected to higher pressure, the voids implode and can generate an

intense shock wave.

Cavitation is a significant cause of wear in some engineering contexts.

Collapsing voids that implode near to a metal surface cause cyclic

stress through repeated implosion. This results in surface fatigue of the metal

causing a type of wear also called "cavitation". The most common examples of

this kind of wear are to pump impellers, and bends where a sudden change in

the direction of liquid occurs. Cavitation is usually divided into two classes of

behavior: inertial (or transient) cavitation and non-inertial cavitation.

Inertial cavitation is the process where a void or bubble in a liquid rapidly

collapses, producing a shock wave. Inertial cavitation occurs in nature in the

strikes of mantis shrimps and pistol shrimps, as well as in the vascular tissues of

plants. In man-made objects, it can occur in control

valves, pumps, propellers and impellers.

Non-inertial cavitation is the process in which a bubble in a fluid is forced to

oscillate in size or shape due to some form of energy input, such as an acoustic

field. Such cavitation is often employed in ultrasonic cleaningbaths and can also

be observed in pumps, propellers, etc.

23

To avoid cavitation at the runner outlet, high head turbines usually need to be

submerged. The level of submergence is calculated. The NPSH required is

calculated from equation

4.3 NET POSITIVE SUCTION HEAD (NPSH) –

In a hydraulic circuit, net positive suction head (NPSH) may refer to one of two

quantities in the analysis of cavitation.

1. The Available NPSH (NPSH A ): a measure of how close the fluid at a given

point is to flashing and so to cavitation.

2. The Required NPSH (NPSH R ): the head value at a specific point (e.g. the

inlet of a pump) required to keep the fluid from cavitating.

NPSH is particularly relevant inside centrifugal pumps and turbines which are

parts of a hydraulic system that are most vulnerable to cavitation. If cavitation

occurs, the drag coefficient of the impeller vanes will increase drastically -

possibly stopping flow altogether - and prolonged exposure will damage the

impeller.

The minimum pressure required at the suction port of the pump to keep the

pump from cavitating. Available NPSH is a function of your system and must

24

be calculated, whereas Required NPSH is a function of the pump and must be

provided by the pump manufacturer.

The Net Positive Suction Head - NPSH - can be defined as

the difference between the Suction Head, and

the Liquids Vapor Head

and can be expressed as

NPSH = hs - hv

Hs = suction head Hv = vapour head

or, by combining (1) and (2)

NPSH = ps / γ + vs2 / 2 g - pv / γ

Y = specific weight

Where NPSH = Net Positive Suction Head (m, in)

The next step in the design is to calculate the inlet parameters, diameter, D1

(Diameter of runner), height of the inlet, B1, and inlet angle, β1. In order to find

these values, the Euler Equation i.e. equation 1 is used. By introducing reduced

dimensionless values and assuming no rotation at the outlet, the efficiency of

turbines and pumps equation be rewritten as-

25

The efficiency, ηh, is set to 0.96. This value accounts for the friction in the

runner and draft tube. For the high head Francis Turbine, U1 is chosen in the

interval of 0.7 to 0.75. However, in case of RPT, the U1 (Peripheral velocity at

inlet) is taken to be nearly equal to 1.

The inlet diameter can now be found by using equation mentioned below. From

the velocity triangle, the expression for the inlet angle can be derived . The

value for Cm1 in is calculated by using the continuity equation . Assuming

minimal acceleration of 10% in the runner, the inlet height is calculated by

using equation-

After the main dimensions of the runner are known, the shape of runner blade

can be designed. The procedure starts by determining the shape of the blade in

axial view, then the radial view is established, and finally, the runner blade can

be plotted in three dimensions.

B. Guide vanes

4.4 FREE VORTEX THEORY –

In free vortex,a fluid gets its rotational motion due to fluid pressure itself and

gravity.It does not require any external force to cause rotation.

26

In a centrifugal pump,the flow of water in the pump casing after it is left the

impeller is considered as free vortex.In free vortex,no energy is imparted to the

liquid in rotation,and as such total head H in bernoullis equation is constant for

all stream lines.In free vortex,the torque remains constant as there is no external

force coming into action.

Since the flow from the guide vanes outlet to the runner inlet is not affected by

the runner blades, free vortex theory is used to define the velocities in the guide

vane path. In general, longer the guide vane, better the directed path of water

but it also results in more friction loses. Hence, suitable length with overlapping

of about 12% need to be selected. Afterwards, the number of guide vanes is

selected such that no water enters the runner in its full closed condition. For the

RPT design, appropriate number of guide vanes was chosen to be 17 and NACA

4415 profile was selected for guide vane shape. The guide vane shaft was fixed

at 2/3 of the guide vane length from guide vane outlet.

27

C. Design of Spiral Casing

For the spiral casing design, the Cm component of velocity remains same

throughout the sections of the spiral casing. The flow of the turbine is 4 m3/s

and Cm velocity at inlet is about 10.84 m/s. The section of the spiral casing is

selected such that it is slightly more than the number of stay vanes, which

results in even flow inside the casing. For our design case, we have selected 20

sections for spiral casing where there are 17 stay vanes. Hence, each section is

at an angular interval of 18 degrees . Now the diameter of each section of spiral

casing is calculated using. The flow in the next section is then calculated. The

diameter and flow for a section can be calculated as-

D. Draft tube

The draft tube recovers kinetic energy at the outlet of the turbine to pressure

energy. It allows the turbine outlet pressure to be lower than the atmospheric

pressure by gradually increasing the cross section area of the tube. A cone type

draft tube was used for the design of RPT. To avoid unfavorable flow patterns

as backflow, the angle ϒ between centerline and the wall was chosen to be 5

degrees. The outlet diameter of the draft tube cone which depends on the length,

which is selected to be little larger than suction head of the turbine to minimize

cavitation effect.

28

4. DESIGN PARAMETERS AND OUTPUTS:

29

5. MODELLING:

The modelling of the different parts are carried out on the basis of the above

mentioned design parameters:

1) Runner

ISOMETRIC VIEW

30

TOP VIEW

The above modelling was carried out in SOLIDWORKS. It involves the

combined modelling of two parts: geometry of runner and guide vane

profile.Exterior geometry of runner was created first in 2-D Sketch using given

input parameters(i.e.diameter,width,thickness etc),later the profile was extruded

in 3-D platform to obtain the model.The guide vanes was separately created and

drafted in a 2-D sketch platform and assembled to the runner using “join”

command.

31

2) Spiral Casing

The modelling of the casing was carried out in SOLIDWORKS.At first,in the

drafting process, the necessary shape was initialised by input parameters.The

geometry was obtained by creating the vertices of casing.These vertices are then

joined to create the edges of the casing.These vertices can be joined by

arc,spline curves etc.Spilne curves were used to create the edges of the

casing.The edges are further joined to create the face of casing.The faces are

finally joined with ‘stitch’ command to obtain the requisite volume.

32

3) Guide Vanes

Figure shows the modelling of a guide vane in PRO-E

The modelling was carried out in PRO-E. The main element in the construction

of a guide-vane is the drafting process.The drafting includes matching the

respective coordinates and elementary profile of NACA-4415 in a 2-D sketch

platform.Joining of the points and coordinates was carried out by the help of

spline-curve or three-point arc option.Once the profile is obtained in 2-D ,it is

then extruded in 3-D platform and with symmetry option,as many number of

guide vanes can be obtained depending on the external diameter of the casing.

33

4) Draft tube:

The modelling of Draft tube was carried out in SOLIDWORKS. The draft angle

was taken as 5 degrees,in order to avoid formation of eddies and flow

separation.The shape was not restricted to cone-frustum ,since the RPT has to

be assembled vertically and the water has to be drained to the lower reservoir

which is situated few metres away from RPT housing.Instead, a rectangular-

frustum shape was chosen for the outlet portion of the draft tube.

34

6.1 ASSEMBLY OF THE RPT HOUSING:

The above two figure shows views of RPT assembly from different

projections.The parts,which are individually created as parts in SOLIDWORKS

are combined together under assembly command.We can easily switch to any

part and the drafted model can then be separately analysed.For analysis using

35

suitable input parameters,the assembly can then be exported to ANSYS

workbench or CFX Fluent.

6. CFD:

The CFD analysis of the RPT was carried out inside the premises of ANSYS

Workbench 14.0. The code uses finite volume approach to solve the governing

equations of fluid motion numerically on a user-defined computational grid. In

this simulation, there are in total four different domains which were meshed

separately in ANSYS meshing. In this study, all the simulations have been

performed in turbine mode and the setups were done for the same condition.

The details of the numerical models for each of the domains are described

below.

A. Spiral Casing

Spiral Casing was discretized with tetrahedral mesh with total node count of

241929. The discretization was made finer towards the interface between the

casing and the guide vane, with mapped mesh property. The inlet of the spiral

casing is the inlet of the whole domain, which have mass flow rate of 4000

kg/m. Towards the outlet of this casing, an interface was defined with the guide

vane inlet. Other regions were defined to be no slip wall. The roughness of the

wall was not included in this analysis.

36

B. Guide Vane

Guide vane was meshed with a total node count of 1038232. In this case, the

size of the mesh is uniform throughout the domain, whereas inlet and outlet of

the domain was considered to be mapped in order to have proper mapping with

neighboring domains. The inlet of the domain was interfaced with spiral casing

outlet and the outlet was interfaced with the runner inlet. The mixing model

between the stationary and rotating domain was done through Stage averaging

between the blade passages. Other boundaries in this domain were considered to

be no slip walls, including the blades.

C. Runner

As runner is the most important region of the whole domain, it was meshed with

finest discretization. The total node count of 5635095 was used to mesh the

runner with finer mesh towards the leading and trailing edges. All boundaries

except the interfaces with stage mixing model were considered to be no slip

walls.

37

D. Draft tube

Draft tube consists of hexahedral mesh with total node count of 832477. The

outlet of the domain was taken as outlet with average static pressure of 1atm.

The inlet of the domain was interfaced with the runner having stage mixing

model. Other boundaries were considered to be no slip wall without roughness.

The type of mesh selected for all the cases is defined in Fig 7. A total of

7747733 mesh nodes were used for the analysis.

7.1 CFD ANALYSIS:

FLOW TRAJECTORIES IN THE ASSEMBLY WITH PRESSURE

VARIATION

The above figure shows the motion trajectory of fluid(i.e.water) entering,and its

variations along the entire spiral casing.The volute profile of spiral casing is

done in order to obtain uniform variation of fluid velocity.The kinetic energy of

38

water increases as it enters and decreases towards the point of exit.The above

shown pressure distribution colludes with the reason stated above,as the

pressure is more along the periphery of casing,and simultaneously decreases

towards the exit,after crossing the guide vanes,as a part of kinetic energy gets

converted into potential head.

PRESSURE VARIATION THROUGHOUT THE CASING

The above analysis of the assembly was done in CFX Fluent platform in

ANSYS Workbench.It shows surface contour and pressure variation along the

periphery of spiral casing.The pressure is less at the entry because of straight

flow path,as the profile changes to volute(curved),the water impinges with a

certain force on the walls of the casing,thereby increasing the pressure at that

region.The pressure remains same,till the fluid(i.e.water) loses its velocity and

then it starts decreasing.Consequently,the pressure becomes minimum at the

centre,where the water loses its velocity and can no longer exert any pressure.

39

PRESSURE VARIATION ALONG THE BLADES

The analysis of above assembly was done in CFX Fluent platform of ANSYS

Workbench.In this analysis,only flow simulation along the guide vanes has been

shown.The water after flowing from spiral casing enters the guide vanes.Since

the main mandate of water is to rotate th guide vanes,it can be done by force of

impact.The pressure distribution along the top surface of the blade is uniform

and same(assuming that the water impinges the blade on its top

surface).Hence,pressure in this area will be more than the bottom surface of the

blade,since there is no impact force of water acting on it.

7.2 RESULTS OF CFD ANALYSIS-

The efficiency of this turbine was measured based on the output power obtained

from the torque produced on the runner and the input power based on the

40

available head. The efficiency of the turbine was found to be 86.71%. From the

runner, the water flows towards the draft tube with high velocity. This velocity

head is converted into the pressure head generating more power and efficiency.

The diverging passage of the draft tube allows for the decrease in the velocity at

the outlet from the equation of continuity. As it can be seen, the maximum

velocity of the flow in the draft tube is around 17 m/s towards inlet, which is

reduced to less than 1 m/s towards the outlet boundary.

VELOCITY STREAMLINES IN TURBINE MODE

The above two figure shows the flow simulation through the casing and draft

tube.It specifically shows the variation in velocity along the casing and draft

tube.Since the velocity of water remains constant along the periphery of spiral

casing,it increases only when it impinges the blades with relatively more higher

velocity,as can be seen from the figure.Whereas ,the main function of the draft

41

tube is to reduce the exit velocity of water and convert it into potential head by

having a larger length tube.It can be seen that the velocity is more at the inlet of

draft tube(i.e. exit of casing) which gets decreased towards the exit of tube.

7. CONCLUSION

The RPT can be used as a secondary unit to supplement the existing main plant

so that it can operate at best efficiency point for most part of the year. The CFD

analysis of the RPT in turbine mode of operation has been completed. The RPT

is a hybrid of centrifugal pump and Francis runner with speed number 0.337. It

has an inlet diameter of 1.39 m and outlet diameter of 0.68 m with inlet height

0.09 m. Despite the simple design techniques involved in the design, the runner

performance is quite satisfactory with a numerical computed efficiency above

88 percent .

8. REFERENCES

O. T. Gabriel Dan Ciocan, Francois Czerwinski, "Variable Speed Pump

Turbines Technology," U.P.B. Science Bulletin, vol. 74, p. 10, 2012.

"Hydroelectric pumped storage technology: international experience,"

Task Committee on Pumped Storage, Committee on Hydropower of the

Energy Division of the American Society of Civil Engineers, New York:

American Society of Civil Engineers.

J. Lal, Hydraulic Machines. India: Metropolitan Book Company, 1994.

42

G. Olimstad, "Characteristics of Reversible- Pump Turbines," PhD,

NTNU, Norway, 2012.

H. Brekke, Hydraulic Turbines Design, Erection and Operation: NTNU

Publication, 2000.

Z. W. Grunde Olimstad, Pål Henrik Finstad, Eve Walseth, Mette Eltvik,

"High Pressure Hydraulic Machinery," NTNU, Norway2009

9. WEBSITES

1)www.sciencedirect.com

2)www.elsevier.com

3)www.ijsrp.org