Successfully reported this slideshow.
We use your LinkedIn profile and activity data to personalize ads and to show you more relevant ads. You can change your ad preferences anytime.

RPT

377 views

Published on

  • Be the first to comment

  • Be the first to like this

RPT

  1. 1. 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. 2. 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. 3. 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. 4. 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. 5. 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. 6. 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. 7. 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. 8. 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. 9. 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. 10. 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. 11. 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. 12. 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. 13. 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. 14. 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. 15. 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. 16. 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. 17. 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. 18. 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. 19. 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. 20. 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. 21. 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. 22. 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. 23. 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. 24. 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 + vs 2 / 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. 25. 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. 26. 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. 27. 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. 28. 28 4. DESIGN PARAMETERS AND OUTPUTS:
  29. 29. 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. 30. 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. 31. 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. 32. 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. 33. 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. 34. 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. 35. 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. 36. 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. 37. 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. 38. 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. 39. 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. 40. 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. 41. 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. 42. 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

×