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Design and performance evaluation of a pump-as-turbine micro-hydro test facility with incorporated inlet flow control
Renewable Energy 78 (2015) 1e6
Contents lists available at ScienceDirect
Renewable Energy journal homepage: www.elsevier.com/locate/renene
Technical note
Design and performance evaluation of a pump-as-turbine micro-hydro test facility with incorporated inlet flow control D.R. Giosio a, *, A.D. Henderson a, J.M. Walker b, P.A. Brandner b, J.E. Sargison c, P. Gautam d a
School of Engineering, The University of Tasmania, Hobart, Tasmania, 7001, Australia National Centre for Maritime Engineering & Hydrodynamics, Australian Maritime College, University of Tasmania, Launceston, Tasmania, 7248, Australia c JSA Consulting Engineers, Sandy Bay, Tasmania, 7005, Australia d Genesis Energy, Hamilton, Waikato, 3204, New Zealand b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 5 February 2014 Accepted 10 December 2014 Available online
In the context of micro-hydro power schemes the initial cost of conventional Francis turbine units is often prohibitive. As such there is growing interest in pump-as-turbine (PAT) technology offering a more cost effective, yet still highly efficient, power generating alternative, finding uses in remote area power supply and energy recovery systems. However, the implementation of a PAT is highly problematic in terms of predicting the installed best operating point coupled with poor off-design performance due to the fixed geometry and absence of inlet flow control. In the current work a micro-hydro test facility and turbine unit is developed utilising a commercially available pump impeller together with a customised housing for incorporation of inlet flow control. Working initially from established PAT theory, this paper presents the design and performance testing of a hydraulic turbine unit suitable for use in rural microhydro, and energy recovery installations. Maximum efficiency of the unit was found to be 79%, marginally higher than that of the parent pump, while the off-design efficiency offered considerable improvement over previously published data of traditional PAT systems. The design provides a cost effective power generator in comparison to small scale Francis turbines, while providing a greater operational range than traditional PAT units. © 2014 Published by Elsevier Ltd.
Keywords: Micro-hydro Hydro power Pump-as-turbine Remote area power generation Waste energy recovery systems
1. Introduction The ability for pumps to operate efficiently in reverse as turbines was first established by Thoma [1] in 1931 while mapping the full operating characteristic of a centrifugal pump. In recent decades there has been renewed interest in pump-as-turbine (PAT) technology which has found significant use in remote area power supply installations, both on- and off-grid, as well as industrial application in energy recovery systems where a high pressure water source exists that would otherwise require throttling. In such instances, within the range designated micro-hydro power (<100 kW), the initial capital cost of a conventional Francis turbine generating set is often prohibitive, making PAT installations an attractive alternative with significantly shorter payback periods.
* Corresponding author. E-mail addresses: [email protected] (D.R. Giosio), Alan.Henderson@utas. edu.au (A.D. Henderson), [email protected] (J.M. Walker), P.Brandner@ utas.edu.au (P.A. Brandner), [email protected] (J.E. Sargison), prakash. [email protected] (P. Gautam). http://dx.doi.org/10.1016/j.renene.2014.12.027 0960-1481/© 2014 Published by Elsevier Ltd.
However, a major drawback of PATs is that the performance away from best efficiency point (BEP) is extremely poor due to the fixed internal geometry and absence of flow regulation. Various authors have provided, with positive results, a number of relatively simple modifications such as impeller tip and hub/shroud rounding in order to increase overall PAT performance [2e4]. However, the rapid efficiency drop-off at off-design conditions remains an inherent and major limitation of PATs. This is further compounded by the current lack of accurate and reliable methods for predicting expected PAT performance (BEP) from available pump manufacturer data. In a review of PAT performance prediction methods Williams [5] defines a prediction criterion, C, as a means to assess the accuracy of eight prediction models suggested by various authors for predicting the turbine BEP based on either pump performance data at BEP, or dimensional pump specific speed, nq ¼ NQ0.5/H0.75 where N [rpm], Q [m3/s] and H [m] at rated. Each method was compared to turbine test data of 35 pumps of various sizes and with nq ranging from 12.7 to 183.3. The method proposed by Sharma [6] was found to be the most accurate, however, 20% of the tested pumps were still outside the
2
D.R. Giosio et al. / Renewable Energy 78 (2015) 1e6
acceptable range of the prediction criterion. More recently, Ventrone et al. [7] presented a detailed method of turbine performance prediction through the definition of a runner momentum coefficient, jR, which describes the specific momentum work performed within the runner proper assuming zero incident losses. For the four pumps tested the proposed method gives values within ±4% of experimental values, although a larger sample size is needed to verify the reliability of the proposed method. Moreover, the method requires specific knowledge of pump geometric parameters, as well as performance curves, and is therefore limited in its ability as a pump selection tool as these details are often not readily available. Consequently, due to the nature and constraints of micro-hydro sites, if the installed BEP is found to differ to some degree from the predicted operating point the PAT will operate at sub-optimal efficiency for a high proportion of its operational life. Recent work by Alexander and Giddens [8], and Alexander et al. [9] has investigated an extensive range of modular, fixed geometry microhydro turbine units. The turbine units described are designed for ease of manufacture and consistantly obtain hydraulic efficiencies greater than 70%. Most importantly, perhaps, is the scaleability of the units, providing micro-hydro developers with units whereby the performance at site conditions may be well predicted. However, as with traditional PAT units, the lack of inlet flow control may somewhat limit the efficient turbine operating range. This paper presents the development of a turbine using an offthe-shelf pump impeller together with a customised housing incorporating inlet flow control. Working initially from established PAT theory the design and performance testing of a hydraulic turbine unit suitable for use in rural micro-hydro and energy recovery installations is given. The resulting design provides a cost effective power generator in comparison to small scale Francis turbines while providing a greater operational range than traditional pumpas-turbine units. In typical micro-hydro power applications, where water storage catchments or industrial operating conditions may vary, the improved off-design efficiency is highly desirable.
water storage, adjustable turbine and generator support frame, pump-turbine unit, clear acrylic circular section draft tube, and open discharge tank with return channel to underfloor storage pit. Flow may be provided directly from the 22 kW supply pump, delivering up to 10 m head at 150 L/s, or via a hydraulically isolated storage tank for investigation of penstock pressure pulsations associated with transient operation, see Figs. 1 and 2. 2.1. Pump as turbine unit design principles In most site installations of PAT units an induction motor is used in place of a synchronous generator [10]. A directly coupled induction generator removes the need for any belts or gearing, minimises any lateral force thereby prolonging bearing life, eliminates the requirement for a turbine shaft bearing, and is low cost particularly compared to synchronous generators of sizes up to 30 kW [10]. However this requires that one of the synchronous speeds, corresponding to the number of poles of the induction generator, must be chosen as the turbine operating speed [4]. As a first approximation the method of Sharma [6] was used to determine the required pump characteristics. In order to allow for operation at a reduced speed an additional correction was made based on turbine affinity laws as Q fN and HfN2 such that
Qt ¼
Nt QpBEP ; Np h0:8 pBEP
Ht ¼
Nt Np
2
HpBEP h1:2 pBEP
(1)
Knowing the available turbine net head and flow rate, Ht and Qt, and the required turbine speed, Nt, Eq. (1a) and Eq. (1b) can be used to select a suitable pump impeller based on readily available manufacturers data. The selected pump impeller was a KSB Ajax I.S. series impeller with turbine runner outlet diameter, D ¼ 226 mm. No blade geometry modifications were performed. The rated values (BEP) of the pump unit from which the impeller is taken, and the predicted and actual turbine performance are given later in Table 1 (Section 4). 2.2. Incorporation of inlet flow control to the PAT unit
2. Experimental facility design A micro-hydro experimental facility was developed comprising a supply pump fitted with variable speed drive (VSD), elevated
To provide optimum inlet incidence angle over a range of flow conditions a guide vane assembly consisting of 13 hydrofoil shaped vanes, with individual linkages to a common ring, was fitted within
Fig. 1. Pump-turbine unit installed at the University of Tasmania micro-hydro test facility.
D.R. Giosio et al. / Renewable Energy 78 (2015) 1e6
3
Fig. 2. Schematic of the micro-hydro test facility at the University of Tasmania.
a simple square section spiral case (Fig. 3). Position is controlled by a programmable 5.3 kN SEW electro-mechanical actuator, however, this could also be achieved by manual operation of a hand wheel depending on the requirements and budget of the given installation.
2.3. Instrumentation For determination of test specific energy Druck UNIK 5000 gauge pressure transducers, 0e100 kPa range, were fitted to ring manifolds with individually valved tappings located at spiral inlet, and draft tube outlet. Each unit has a frequency response of 3.5 kHz and was calibrated over the full range with a maximum uncertainty of 0.06% full scale. Flow rate measurement is achieved using a Yokogawa US300PM ultrasonic flow meter fitted upstream the turbine inlet with an accuracy of ±0.01 m/s over the sonic path. An optical sensor and 500 hole disc was mounted on the nondrive end of the generator for test speed measurement. For steady state tests a rolling average was used, updated every pulse. An independent induction pick-up was mounted on the drive end for grid synchronisation and overspeed protection. Shaft torque measurement was achieved with a full bridge 350 U strain gauge installed on an exposed section of turbine shaft with data acquisition via a KMT digital telemetry system. The encoder scanning rate was 6.94 kHz with a system accuracy of ±0.2%. Turbine unit efficiency is therefore stated with an uncertainty of 1.3% based on the propagation of errors method [11].
E¼
p1 p2 v21 v22 þ þ gðz1 z2 Þ½J=kg r 2
(2)
where subscripts 1 and 2 refer to measurements at high and low pressure side, respectively. The hydraulic power available to the turbine is therefore a product of the test specific energy and the mass flow rate calculated as
Ph ¼ EðrQ Þ½W
(3)
The power transferred to the turbine shaft is given by
Ps ¼ 2pn$T½W
(4)
where the torque, T, is calculated from the measured shaft strain as
T¼
pG D4s Fc ½Nm=v 8D 106
(5)
where G is the shear modulus of the solid stainless steel shaft (of diameter Ds) and Fc is the strain gauge calibration coefficient. Turbine performance is presented in terms of three key dimensionless parameters: energy coefficient, EnD, flow coefficient, QnD, and the power coefficient, PnD, defined as
EnD ¼
E Q Pm ; QnD ¼ ; PnD ¼ 3 5 n2 D 2 nD3 rn D
(6)
The hill chart of turbine efficiency contours shown in Fig. 4 is given in terms of overall efficiency, ho, which is defined as the power transferred to the turbine shaft relative to the available hydraulic power.
3. Experimental methodology During steady state performance testing data was recorded at each operational point with a sample frequency of 100 Hz, acquired by an NI M Series Multifunction DAQ SCB-68 card. For the construction of the turbine efficiency hill chart tests were conducted for guide vane opening angles ranging between 20 and 35 , equivalent to a range of between 53% and 92% full stroke. Tests were conducted at net specific energy values of 34.3 J/kg up to 83.3 J/kg in 5 J/kg increments to ensure a thorough exploration of the entire operating range. In accordance with IEC Standards [12] the net specific energy across the machine was calculated as
Table 1 Best efficiency point values of the selected PAT operating as pump, turbine predicted & turbine actual. Values at BEP
nq [e]
H [m]
Q [m3/s]
N [rpm]
P [kW]
h [%]
Pump as rated Turbine predicted Turbine actual
104.3 92.4 71.9
12.25 4.38 5.98
0.222 0.139 0.133
1450 750 754
34.0 4.68 6.20
78.5 78.5 79.0
ho ¼
Ps 2pn$T ¼ Ph E$ðrQ Þ
(7)
As such the overall efficiency incorporates mechanical efficiency, taking into account mechanical losses in shaft seals and bearings; volumetric efficiency, taking into account leakage and cooling water flow; and hydraulic efficiency, defining the amount of power extracted by the runner from the power available from the water. Owing to the tight internal tolerances of the casing the volumetric efficiency can be approximated to hv z 1, with only small losses due to cooling water flow, so that
ho ¼ hm hv hh zhm hh
(8)
Mechanical efficiency due to losses in bearings, shaft seals, and wear rings was estimated by running the turbine up to rated speed and connecting to the grid. Guide vanes were then closed and air injected into the draft tube in order to fill the runner chamber. With cooling lines open the torque required to maintain constant speed, Tloss, was recorded.
4
D.R. Giosio et al. / Renewable Energy 78 (2015) 1e6
Fig. 3. Square section spiral case incorporating the guide vane mechanism prior to final assembly.
hm ¼
Pt 2pnðT þ Tloss Þ
(9)
The mechanical efficiency as determined by Eq. (9), was assumed constant over the entire operating range, however in practice this will vary to a relatively small degree.
4. Results The efficiency of the designed micro-hydro pump-turbine unit over the entire operating range is illustrated in the hill chart shown in Fig. 4. Overall efficiency contours are given, while guide vane opening angles are represented by the dotted lines running diagonally left to right. The best efficiency point was found to be at EnD ¼ 7.27, QnD ¼ 0.925, giving a maximum overall turbine efficiency of 79.0%, measured at turbine shaft. This result is in good
agreement with the predicted best efficiency of a pump impeller acting as turbine which is generally reported to be within the range of ±2% of the best efficiency in pump operation [5]. Performance of the micro pump-turbine unit in comparison to the parent pump and predicted turbine values is given in Table 1. While the maximum efficiency shows a slight increase of 0.5% from that in pumping mode, the net head and output power at rated were considerably higher than predicted by the method of Sharma. Furthermore the flow rate in turbine operation was slightly less than that expected. It is widely accepted that PATs require a higher head and flow at BEP point than in pump mode, with the head ratio generally the larger of the two. Ventrone et al. [7] suggests that this is due to a combination of a significant increase in head coefficient with increasing flow and the simultaneous decrease in mechanical losses relative to power output as well as a residual negative tangential velocity at operating condition just greater than design.
Fig. 4. Performance Hill Chart diagram of the micro-hydro turbine unit.
D.R. Giosio et al. / Renewable Energy 78 (2015) 1e6
5
Fig. 5. Turbine efficiency (B) and power coefficient (,) at constant specific energies corresponding to tank min. level, BEP and tank max. level.
This is in contrast with the belief that slip, due to a relative rotational velocity within the blade passages, is only present in pump mode such that flow should be increased in order to minimise incidence losses. However, a degree of slip is also present in turbine operation, and an increase in incidence loss due to higher flow is balanced in some degree by the mechanical efficiency. This decreasing influence of mechanical losses can be seen in Fig. 6 whereby the overall efficiency approaches the hydraulic efficiency at normalised flow Q/QBEP > 1. Of practical interest is the general flatness of the efficiency map, particularly around the BEP. This is also shown in Fig. 5, presented for three cases of constant specific energy corresponding to tank minimum supply level, BEP condition and tank maximum supply level. The variation in head represents a range of static head between approximately 0.85HBEP and 1.2HBEP. The efficiency in each case remains considerably high even with significant decreases in flow. Indeed, in all cases a 30% flow reduction is required to see a 10% reduction in overall efficiency, and a further 20% decrease in
flow for an additional 10% efficiency reduction at rated EnD. This compares favourably with previously presented data on PAT performance by Yang et al. [13], Williams et al. [10], Derakhshan and Nourbakhsh [14] and Singh [2] that all report significant reduction in turbine operation efficiency due to the inherent lack of inlet flow control and associated large incidence losses. In the sphere of micro-hydro remote area power installations the ability to operate efficiently at off-design conditions is extremely important as water storage reservoirs are often limited in capacity and catchments are subject to highly fluctuating, often seasonal inflows. 4.1. Weighted-average-efficiency (WAE) The weighted average efficiency provides developers of microhydro installations a means to quantitatively compare generation alternatives, taking into consideration the likely percentage of time the unit will be operating under off-design conditions. As defined by Eq. (10), the WAE is determined at 1.0Pr, 0.8Pr and 0.6.Pr
Fig. 6. Turbine efficiency (B) and power (△) at constant guide vane opening of 30 . Power and flow rate normalised by values at B.E.P.
6
D.R. Giosio et al. / Renewable Energy 78 (2015) 1e6
acknowledgements to Prakash Gautam and Paul Coull for their contributions to the turbine design. Nomenclature
Fig. 7. Normalised efficiency of the new micro-hydro turbine unit (UTAS-MH) and that of an optimised PAT [2] at three load points 1.0Pr, 0.8Pr, and 0.6Pr.
WAE ¼ a1 $ht;1:0 þ a2 $ht;0:8 þ a3 $ht;0:6
(10)
where a1 ¼ 0.6, a2 ¼ 0.2, a3 ¼ 0.2 are weight factors representing the expected percentage of operational time at each off-design loading. Values given were choosen such that a direct comparision may be made with data presented by Singh [2], although these may be changed to reflect conditions at a given site. The performance of the designed micro-hydro turbine unit with incorporated inlet flow control is compared to performance results of a similar traditional PAT unit presented by Singh [2]. The traditional PAT tested by Singh [2] had undergone numerous stages of optimisation including impeller blade tip rounding and tapered ring inserts, resulting in an increased peak efficiency of 83.4%, while the designed micro-hydro turbine runner had not undergone any modifications. Even so, the improved performance at off-design conditions has essentially made up for a 4.4% disparity in peak efficiency with WAEs of 76.8% and 76.9% for the new micro-hydro turbine and the optimised PAT, respectively. This is further illustrated in Fig. 7. 5. Conclusions The development of a 6.2 kW micro-hydro turbine unit and test facility was presented. The turbine unit was designed to incorporate a commercially available pump impeller operated in turbine mode, together with custom spiral casing and guide vane mechanism for inlet flow control. Steady state experimental testing over the full operating range indicated a maximum overall efficiency of 79%, in good agreement with PAT theory. The net head and output power at rated conditions were somewhat higher than predicted by the method of Sharma [6], while the flow rate at rated was slightly less than expected. Importantly, in the context of micro-hydro and energy recovery systems, while slightly more complex than a traditional PAT system the turbine demonstrated near peak efficiency operation over a wide range of flow conditions, as indicated by the weighted-average-efficiency. The designed micro-hydro turbine unit thereby addresses the main drawbacks of pumps operating in turbine mode, providing a low cost alternative generating solution for application in remote area power supply and industrial energy recovery systems. Acknowledgements This project was funded through ARC Linkage Grant (LP 110200244) and industry partner Hydro Tasmania. Turbine design in collaboration with Pentair Flow Technologies, Tasmania. Special
D Ds E Fc G H N P Q g n nq p z v
h p r
turbine outlet diameter [m] shaft diameter [m] specific energy [J/kg] strain gauge coefficient [e] shear modulus [Pa] head [m] rotational speed [rpm] power [W] flow rate [m3/s] acceleration due to gravity [m/s2] rotational speed [s1] specific speed, dimensional [rpm] static pressure [Pa] elevation [m] flow velocity [m/s] efficiency [%] mathematical constant [e] density [kg/m3]
Abbreviations BEP best efficiency point PAT pump-as-turbine Subscripts h hydraulic m mechanical o overall p pump r rated s shaft t turbine v volumetric References [1] Thoma D, Kittredge CP. Centrifugal pumps operated under abnormal conditions. Power 1931;73:881e4. [2] Singh P. Optimization of the internal hydraulics and of system design for pumps as turbines with field implementation and evaluation. Ph.D. thesis. Germany: University of Karlsruhe; 2005. [3] Singh P, Nestmann F. Internal hydraulic analysis of impeller rounding in centrifugal pumps as turbines. Exp Therm Fluid Sci 2011;35(No. 1):121e34. [4] Derakhshan S, Mohmmadi B, Nourbakhsh A. Efficiency improvement of centrifugal reverse pumps. J Fluids Eng 2009;131(No. 2). Article ID 021103. [5] Williams AA. The turbine performance of centrifugal pumps: a comparison of prediction methods. Proc Inst Mech Eng 1994;208:59e66. [6] Sharma KR. Small hydroelectric projects e use of centrifugal pumps as turbines. Bangalore, India: Kirloskar Electric Co; 1985. [7] Ventrone G, Ardizzon G, Pavesi G. Direct and reverse flow conditions in radial flow hydraulic turbomachines. Proc Inst Mech Eng Part A: J Power Energy 2000;214(6):635e44. [8] Alexander K, Giddens E. Microhydro: cost-effective, modular systems for low heads. Renew Energy 2008;33(6):1379e91. [9] Alexander K, Giddens E, Fuller A. Radial- and mixed-flow turbines for low head microhydro systems. Renew Energy 2009;34(7):1885e94. [10] Williams AA, Smith NPA, Bird C, Howard M. Pumps as turbines and induction motors as generators for energy recovery in water supply systems. J CIWEM 1998;12:175e8. [11] Figliola RS, Beasley DE. Theory and design for mechanical measurements. 3rd ed. John Wiley & Sons, Inc; 2000. [12] International Standards. IEC 60193 second edition 1999-11. [13] Yang SS, Derakhshan S, Kong FY. Theoretical, numerical and experimental prediction of pump as turbine performance. Renew Energy 2012;48:507e13. [14] Derakhshan S, Nourbakhsh A. Experimental study of characteristic curves of centrifugal pumps working as turbines in different specific speeds. Exp Therm Fluid Sci 2008;32:800e7.
Contents lists available at ScienceDirect
Renewable Energy journal homepage: www.elsevier.com/locate/renene
Technical note
Design and performance evaluation of a pump-as-turbine micro-hydro test facility with incorporated inlet flow control D.R. Giosio a, *, A.D. Henderson a, J.M. Walker b, P.A. Brandner b, J.E. Sargison c, P. Gautam d a
School of Engineering, The University of Tasmania, Hobart, Tasmania, 7001, Australia National Centre for Maritime Engineering & Hydrodynamics, Australian Maritime College, University of Tasmania, Launceston, Tasmania, 7248, Australia c JSA Consulting Engineers, Sandy Bay, Tasmania, 7005, Australia d Genesis Energy, Hamilton, Waikato, 3204, New Zealand b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 5 February 2014 Accepted 10 December 2014 Available online
In the context of micro-hydro power schemes the initial cost of conventional Francis turbine units is often prohibitive. As such there is growing interest in pump-as-turbine (PAT) technology offering a more cost effective, yet still highly efficient, power generating alternative, finding uses in remote area power supply and energy recovery systems. However, the implementation of a PAT is highly problematic in terms of predicting the installed best operating point coupled with poor off-design performance due to the fixed geometry and absence of inlet flow control. In the current work a micro-hydro test facility and turbine unit is developed utilising a commercially available pump impeller together with a customised housing for incorporation of inlet flow control. Working initially from established PAT theory, this paper presents the design and performance testing of a hydraulic turbine unit suitable for use in rural microhydro, and energy recovery installations. Maximum efficiency of the unit was found to be 79%, marginally higher than that of the parent pump, while the off-design efficiency offered considerable improvement over previously published data of traditional PAT systems. The design provides a cost effective power generator in comparison to small scale Francis turbines, while providing a greater operational range than traditional PAT units. © 2014 Published by Elsevier Ltd.
Keywords: Micro-hydro Hydro power Pump-as-turbine Remote area power generation Waste energy recovery systems
1. Introduction The ability for pumps to operate efficiently in reverse as turbines was first established by Thoma [1] in 1931 while mapping the full operating characteristic of a centrifugal pump. In recent decades there has been renewed interest in pump-as-turbine (PAT) technology which has found significant use in remote area power supply installations, both on- and off-grid, as well as industrial application in energy recovery systems where a high pressure water source exists that would otherwise require throttling. In such instances, within the range designated micro-hydro power (<100 kW), the initial capital cost of a conventional Francis turbine generating set is often prohibitive, making PAT installations an attractive alternative with significantly shorter payback periods.
* Corresponding author. E-mail addresses: [email protected] (D.R. Giosio), Alan.Henderson@utas. edu.au (A.D. Henderson), [email protected] (J.M. Walker), P.Brandner@ utas.edu.au (P.A. Brandner), [email protected] (J.E. Sargison), prakash. [email protected] (P. Gautam). http://dx.doi.org/10.1016/j.renene.2014.12.027 0960-1481/© 2014 Published by Elsevier Ltd.
However, a major drawback of PATs is that the performance away from best efficiency point (BEP) is extremely poor due to the fixed internal geometry and absence of flow regulation. Various authors have provided, with positive results, a number of relatively simple modifications such as impeller tip and hub/shroud rounding in order to increase overall PAT performance [2e4]. However, the rapid efficiency drop-off at off-design conditions remains an inherent and major limitation of PATs. This is further compounded by the current lack of accurate and reliable methods for predicting expected PAT performance (BEP) from available pump manufacturer data. In a review of PAT performance prediction methods Williams [5] defines a prediction criterion, C, as a means to assess the accuracy of eight prediction models suggested by various authors for predicting the turbine BEP based on either pump performance data at BEP, or dimensional pump specific speed, nq ¼ NQ0.5/H0.75 where N [rpm], Q [m3/s] and H [m] at rated. Each method was compared to turbine test data of 35 pumps of various sizes and with nq ranging from 12.7 to 183.3. The method proposed by Sharma [6] was found to be the most accurate, however, 20% of the tested pumps were still outside the
2
D.R. Giosio et al. / Renewable Energy 78 (2015) 1e6
acceptable range of the prediction criterion. More recently, Ventrone et al. [7] presented a detailed method of turbine performance prediction through the definition of a runner momentum coefficient, jR, which describes the specific momentum work performed within the runner proper assuming zero incident losses. For the four pumps tested the proposed method gives values within ±4% of experimental values, although a larger sample size is needed to verify the reliability of the proposed method. Moreover, the method requires specific knowledge of pump geometric parameters, as well as performance curves, and is therefore limited in its ability as a pump selection tool as these details are often not readily available. Consequently, due to the nature and constraints of micro-hydro sites, if the installed BEP is found to differ to some degree from the predicted operating point the PAT will operate at sub-optimal efficiency for a high proportion of its operational life. Recent work by Alexander and Giddens [8], and Alexander et al. [9] has investigated an extensive range of modular, fixed geometry microhydro turbine units. The turbine units described are designed for ease of manufacture and consistantly obtain hydraulic efficiencies greater than 70%. Most importantly, perhaps, is the scaleability of the units, providing micro-hydro developers with units whereby the performance at site conditions may be well predicted. However, as with traditional PAT units, the lack of inlet flow control may somewhat limit the efficient turbine operating range. This paper presents the development of a turbine using an offthe-shelf pump impeller together with a customised housing incorporating inlet flow control. Working initially from established PAT theory the design and performance testing of a hydraulic turbine unit suitable for use in rural micro-hydro and energy recovery installations is given. The resulting design provides a cost effective power generator in comparison to small scale Francis turbines while providing a greater operational range than traditional pumpas-turbine units. In typical micro-hydro power applications, where water storage catchments or industrial operating conditions may vary, the improved off-design efficiency is highly desirable.
water storage, adjustable turbine and generator support frame, pump-turbine unit, clear acrylic circular section draft tube, and open discharge tank with return channel to underfloor storage pit. Flow may be provided directly from the 22 kW supply pump, delivering up to 10 m head at 150 L/s, or via a hydraulically isolated storage tank for investigation of penstock pressure pulsations associated with transient operation, see Figs. 1 and 2. 2.1. Pump as turbine unit design principles In most site installations of PAT units an induction motor is used in place of a synchronous generator [10]. A directly coupled induction generator removes the need for any belts or gearing, minimises any lateral force thereby prolonging bearing life, eliminates the requirement for a turbine shaft bearing, and is low cost particularly compared to synchronous generators of sizes up to 30 kW [10]. However this requires that one of the synchronous speeds, corresponding to the number of poles of the induction generator, must be chosen as the turbine operating speed [4]. As a first approximation the method of Sharma [6] was used to determine the required pump characteristics. In order to allow for operation at a reduced speed an additional correction was made based on turbine affinity laws as Q fN and HfN2 such that
Qt ¼
Nt QpBEP ; Np h0:8 pBEP
Ht ¼
Nt Np
2
HpBEP h1:2 pBEP
(1)
Knowing the available turbine net head and flow rate, Ht and Qt, and the required turbine speed, Nt, Eq. (1a) and Eq. (1b) can be used to select a suitable pump impeller based on readily available manufacturers data. The selected pump impeller was a KSB Ajax I.S. series impeller with turbine runner outlet diameter, D ¼ 226 mm. No blade geometry modifications were performed. The rated values (BEP) of the pump unit from which the impeller is taken, and the predicted and actual turbine performance are given later in Table 1 (Section 4). 2.2. Incorporation of inlet flow control to the PAT unit
2. Experimental facility design A micro-hydro experimental facility was developed comprising a supply pump fitted with variable speed drive (VSD), elevated
To provide optimum inlet incidence angle over a range of flow conditions a guide vane assembly consisting of 13 hydrofoil shaped vanes, with individual linkages to a common ring, was fitted within
Fig. 1. Pump-turbine unit installed at the University of Tasmania micro-hydro test facility.
D.R. Giosio et al. / Renewable Energy 78 (2015) 1e6
3
Fig. 2. Schematic of the micro-hydro test facility at the University of Tasmania.
a simple square section spiral case (Fig. 3). Position is controlled by a programmable 5.3 kN SEW electro-mechanical actuator, however, this could also be achieved by manual operation of a hand wheel depending on the requirements and budget of the given installation.
2.3. Instrumentation For determination of test specific energy Druck UNIK 5000 gauge pressure transducers, 0e100 kPa range, were fitted to ring manifolds with individually valved tappings located at spiral inlet, and draft tube outlet. Each unit has a frequency response of 3.5 kHz and was calibrated over the full range with a maximum uncertainty of 0.06% full scale. Flow rate measurement is achieved using a Yokogawa US300PM ultrasonic flow meter fitted upstream the turbine inlet with an accuracy of ±0.01 m/s over the sonic path. An optical sensor and 500 hole disc was mounted on the nondrive end of the generator for test speed measurement. For steady state tests a rolling average was used, updated every pulse. An independent induction pick-up was mounted on the drive end for grid synchronisation and overspeed protection. Shaft torque measurement was achieved with a full bridge 350 U strain gauge installed on an exposed section of turbine shaft with data acquisition via a KMT digital telemetry system. The encoder scanning rate was 6.94 kHz with a system accuracy of ±0.2%. Turbine unit efficiency is therefore stated with an uncertainty of 1.3% based on the propagation of errors method [11].
E¼
p1 p2 v21 v22 þ þ gðz1 z2 Þ½J=kg r 2
(2)
where subscripts 1 and 2 refer to measurements at high and low pressure side, respectively. The hydraulic power available to the turbine is therefore a product of the test specific energy and the mass flow rate calculated as
Ph ¼ EðrQ Þ½W
(3)
The power transferred to the turbine shaft is given by
Ps ¼ 2pn$T½W
(4)
where the torque, T, is calculated from the measured shaft strain as
T¼
pG D4s Fc ½Nm=v 8D 106
(5)
where G is the shear modulus of the solid stainless steel shaft (of diameter Ds) and Fc is the strain gauge calibration coefficient. Turbine performance is presented in terms of three key dimensionless parameters: energy coefficient, EnD, flow coefficient, QnD, and the power coefficient, PnD, defined as
EnD ¼
E Q Pm ; QnD ¼ ; PnD ¼ 3 5 n2 D 2 nD3 rn D
(6)
The hill chart of turbine efficiency contours shown in Fig. 4 is given in terms of overall efficiency, ho, which is defined as the power transferred to the turbine shaft relative to the available hydraulic power.
3. Experimental methodology During steady state performance testing data was recorded at each operational point with a sample frequency of 100 Hz, acquired by an NI M Series Multifunction DAQ SCB-68 card. For the construction of the turbine efficiency hill chart tests were conducted for guide vane opening angles ranging between 20 and 35 , equivalent to a range of between 53% and 92% full stroke. Tests were conducted at net specific energy values of 34.3 J/kg up to 83.3 J/kg in 5 J/kg increments to ensure a thorough exploration of the entire operating range. In accordance with IEC Standards [12] the net specific energy across the machine was calculated as
Table 1 Best efficiency point values of the selected PAT operating as pump, turbine predicted & turbine actual. Values at BEP
nq [e]
H [m]
Q [m3/s]
N [rpm]
P [kW]
h [%]
Pump as rated Turbine predicted Turbine actual
104.3 92.4 71.9
12.25 4.38 5.98
0.222 0.139 0.133
1450 750 754
34.0 4.68 6.20
78.5 78.5 79.0
ho ¼
Ps 2pn$T ¼ Ph E$ðrQ Þ
(7)
As such the overall efficiency incorporates mechanical efficiency, taking into account mechanical losses in shaft seals and bearings; volumetric efficiency, taking into account leakage and cooling water flow; and hydraulic efficiency, defining the amount of power extracted by the runner from the power available from the water. Owing to the tight internal tolerances of the casing the volumetric efficiency can be approximated to hv z 1, with only small losses due to cooling water flow, so that
ho ¼ hm hv hh zhm hh
(8)
Mechanical efficiency due to losses in bearings, shaft seals, and wear rings was estimated by running the turbine up to rated speed and connecting to the grid. Guide vanes were then closed and air injected into the draft tube in order to fill the runner chamber. With cooling lines open the torque required to maintain constant speed, Tloss, was recorded.
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Fig. 3. Square section spiral case incorporating the guide vane mechanism prior to final assembly.
hm ¼
Pt 2pnðT þ Tloss Þ
(9)
The mechanical efficiency as determined by Eq. (9), was assumed constant over the entire operating range, however in practice this will vary to a relatively small degree.
4. Results The efficiency of the designed micro-hydro pump-turbine unit over the entire operating range is illustrated in the hill chart shown in Fig. 4. Overall efficiency contours are given, while guide vane opening angles are represented by the dotted lines running diagonally left to right. The best efficiency point was found to be at EnD ¼ 7.27, QnD ¼ 0.925, giving a maximum overall turbine efficiency of 79.0%, measured at turbine shaft. This result is in good
agreement with the predicted best efficiency of a pump impeller acting as turbine which is generally reported to be within the range of ±2% of the best efficiency in pump operation [5]. Performance of the micro pump-turbine unit in comparison to the parent pump and predicted turbine values is given in Table 1. While the maximum efficiency shows a slight increase of 0.5% from that in pumping mode, the net head and output power at rated were considerably higher than predicted by the method of Sharma. Furthermore the flow rate in turbine operation was slightly less than that expected. It is widely accepted that PATs require a higher head and flow at BEP point than in pump mode, with the head ratio generally the larger of the two. Ventrone et al. [7] suggests that this is due to a combination of a significant increase in head coefficient with increasing flow and the simultaneous decrease in mechanical losses relative to power output as well as a residual negative tangential velocity at operating condition just greater than design.
Fig. 4. Performance Hill Chart diagram of the micro-hydro turbine unit.
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5
Fig. 5. Turbine efficiency (B) and power coefficient (,) at constant specific energies corresponding to tank min. level, BEP and tank max. level.
This is in contrast with the belief that slip, due to a relative rotational velocity within the blade passages, is only present in pump mode such that flow should be increased in order to minimise incidence losses. However, a degree of slip is also present in turbine operation, and an increase in incidence loss due to higher flow is balanced in some degree by the mechanical efficiency. This decreasing influence of mechanical losses can be seen in Fig. 6 whereby the overall efficiency approaches the hydraulic efficiency at normalised flow Q/QBEP > 1. Of practical interest is the general flatness of the efficiency map, particularly around the BEP. This is also shown in Fig. 5, presented for three cases of constant specific energy corresponding to tank minimum supply level, BEP condition and tank maximum supply level. The variation in head represents a range of static head between approximately 0.85HBEP and 1.2HBEP. The efficiency in each case remains considerably high even with significant decreases in flow. Indeed, in all cases a 30% flow reduction is required to see a 10% reduction in overall efficiency, and a further 20% decrease in
flow for an additional 10% efficiency reduction at rated EnD. This compares favourably with previously presented data on PAT performance by Yang et al. [13], Williams et al. [10], Derakhshan and Nourbakhsh [14] and Singh [2] that all report significant reduction in turbine operation efficiency due to the inherent lack of inlet flow control and associated large incidence losses. In the sphere of micro-hydro remote area power installations the ability to operate efficiently at off-design conditions is extremely important as water storage reservoirs are often limited in capacity and catchments are subject to highly fluctuating, often seasonal inflows. 4.1. Weighted-average-efficiency (WAE) The weighted average efficiency provides developers of microhydro installations a means to quantitatively compare generation alternatives, taking into consideration the likely percentage of time the unit will be operating under off-design conditions. As defined by Eq. (10), the WAE is determined at 1.0Pr, 0.8Pr and 0.6.Pr
Fig. 6. Turbine efficiency (B) and power (△) at constant guide vane opening of 30 . Power and flow rate normalised by values at B.E.P.
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acknowledgements to Prakash Gautam and Paul Coull for their contributions to the turbine design. Nomenclature
Fig. 7. Normalised efficiency of the new micro-hydro turbine unit (UTAS-MH) and that of an optimised PAT [2] at three load points 1.0Pr, 0.8Pr, and 0.6Pr.
WAE ¼ a1 $ht;1:0 þ a2 $ht;0:8 þ a3 $ht;0:6
(10)
where a1 ¼ 0.6, a2 ¼ 0.2, a3 ¼ 0.2 are weight factors representing the expected percentage of operational time at each off-design loading. Values given were choosen such that a direct comparision may be made with data presented by Singh [2], although these may be changed to reflect conditions at a given site. The performance of the designed micro-hydro turbine unit with incorporated inlet flow control is compared to performance results of a similar traditional PAT unit presented by Singh [2]. The traditional PAT tested by Singh [2] had undergone numerous stages of optimisation including impeller blade tip rounding and tapered ring inserts, resulting in an increased peak efficiency of 83.4%, while the designed micro-hydro turbine runner had not undergone any modifications. Even so, the improved performance at off-design conditions has essentially made up for a 4.4% disparity in peak efficiency with WAEs of 76.8% and 76.9% for the new micro-hydro turbine and the optimised PAT, respectively. This is further illustrated in Fig. 7. 5. Conclusions The development of a 6.2 kW micro-hydro turbine unit and test facility was presented. The turbine unit was designed to incorporate a commercially available pump impeller operated in turbine mode, together with custom spiral casing and guide vane mechanism for inlet flow control. Steady state experimental testing over the full operating range indicated a maximum overall efficiency of 79%, in good agreement with PAT theory. The net head and output power at rated conditions were somewhat higher than predicted by the method of Sharma [6], while the flow rate at rated was slightly less than expected. Importantly, in the context of micro-hydro and energy recovery systems, while slightly more complex than a traditional PAT system the turbine demonstrated near peak efficiency operation over a wide range of flow conditions, as indicated by the weighted-average-efficiency. The designed micro-hydro turbine unit thereby addresses the main drawbacks of pumps operating in turbine mode, providing a low cost alternative generating solution for application in remote area power supply and industrial energy recovery systems. Acknowledgements This project was funded through ARC Linkage Grant (LP 110200244) and industry partner Hydro Tasmania. Turbine design in collaboration with Pentair Flow Technologies, Tasmania. Special
D Ds E Fc G H N P Q g n nq p z v
h p r
turbine outlet diameter [m] shaft diameter [m] specific energy [J/kg] strain gauge coefficient [e] shear modulus [Pa] head [m] rotational speed [rpm] power [W] flow rate [m3/s] acceleration due to gravity [m/s2] rotational speed [s1] specific speed, dimensional [rpm] static pressure [Pa] elevation [m] flow velocity [m/s] efficiency [%] mathematical constant [e] density [kg/m3]
Abbreviations BEP best efficiency point PAT pump-as-turbine Subscripts h hydraulic m mechanical o overall p pump r rated s shaft t turbine v volumetric References [1] Thoma D, Kittredge CP. Centrifugal pumps operated under abnormal conditions. Power 1931;73:881e4. [2] Singh P. Optimization of the internal hydraulics and of system design for pumps as turbines with field implementation and evaluation. Ph.D. thesis. Germany: University of Karlsruhe; 2005. [3] Singh P, Nestmann F. Internal hydraulic analysis of impeller rounding in centrifugal pumps as turbines. Exp Therm Fluid Sci 2011;35(No. 1):121e34. [4] Derakhshan S, Mohmmadi B, Nourbakhsh A. Efficiency improvement of centrifugal reverse pumps. J Fluids Eng 2009;131(No. 2). Article ID 021103. [5] Williams AA. The turbine performance of centrifugal pumps: a comparison of prediction methods. Proc Inst Mech Eng 1994;208:59e66. [6] Sharma KR. Small hydroelectric projects e use of centrifugal pumps as turbines. Bangalore, India: Kirloskar Electric Co; 1985. [7] Ventrone G, Ardizzon G, Pavesi G. Direct and reverse flow conditions in radial flow hydraulic turbomachines. Proc Inst Mech Eng Part A: J Power Energy 2000;214(6):635e44. [8] Alexander K, Giddens E. Microhydro: cost-effective, modular systems for low heads. Renew Energy 2008;33(6):1379e91. [9] Alexander K, Giddens E, Fuller A. Radial- and mixed-flow turbines for low head microhydro systems. Renew Energy 2009;34(7):1885e94. [10] Williams AA, Smith NPA, Bird C, Howard M. Pumps as turbines and induction motors as generators for energy recovery in water supply systems. J CIWEM 1998;12:175e8. [11] Figliola RS, Beasley DE. Theory and design for mechanical measurements. 3rd ed. John Wiley & Sons, Inc; 2000. [12] International Standards. IEC 60193 second edition 1999-11. [13] Yang SS, Derakhshan S, Kong FY. Theoretical, numerical and experimental prediction of pump as turbine performance. Renew Energy 2012;48:507e13. [14] Derakhshan S, Nourbakhsh A. Experimental study of characteristic curves of centrifugal pumps working as turbines in different specific speeds. Exp Therm Fluid Sci 2008;32:800e7.