Performance of a bulb turbine suitable for low prototype head: model test and transient numerical simulation

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1 IOP Conference Series: Earth and Environmental Science Performance of a bulb turbine suitable for low prototype head: model test and transient numerical simulation To cite this article: L Zhu et al IOP Conf. Ser.: Earth Environ. Sci Related content - A modification method on runner blades in a Bulb turbine W Yang, Y Wu and S Liu - Pressure pulsation in Kaplan turbines: Prototype-CFD comparison A Rivetti, C Lucino, S Liscia et al. - Numerical simulation of turbulence flow in a Kaplan turbine -Evaluation on turbine performance prediction accuracy- P Ko and S Kurosawa View the article online for updates and enhancements. This content was downloaded from IP address on /9/8 at 3:4

2 6th IAHR Symposium on Hydraulic Machinery and Systems Performance of a bulb turbine suitable for low prototype head:model test and transient numerical simulation L Zhu, H P Zhang, J G Zhang, X C Meng and L Lu Laboratory of Hydraulic Machinery, China Institute of Water Resources and Hydropower Research (IWHR), A FUXING Road, Beijing, 38, P.R. China zhulei@iwhr.com Abstract. In this paper, a bulb turbine, with unit specific speed of n q =3. min -, suitable for low prototype head was studied from aspect of its performance. Hydraulic model of the turbine was developed firstly, and then model turbine was designed and manufactured. Performance tests were carried out on high-accuracy hydraulic machinery model universal test rig located at IWHR, including energy, cavitation and pressure fluctuation tests, etc. In order to investigate internal flow field, three-dimensional transient turbulence numerical simulation was conducted on the tested turbine, adopting Reynolds-averaged Navier-Stocks control equations and RNG k-ε turbulence model. Test and simulation results show that: () hydraulic efficiency of model turbine η M is up to 9.7%, at the optimum operating point of n o =65.54 r/min versus Q o =.93 m 3 /s; () numerical results agree well with experimental resultsby comparing pressure fluctuation, which shows that pressure amplitude is very low at the optimum operating point; (3) hydraulic loss in Outflow domain accounts for more than 5% total hydraulic loss due to flow separation and secondary flow.. Introduction Tubular turbines are applicable to low water head hydropower plant, which can be used for less than 3 m head hydropower resources development, can also be used for the ocean tidal power exploitation. In China, the East and South Regions are rich in low head water resources. At present, in the situation of high-head water resources have been fully utilized, low head water resources has become an important development trend, which will lead to tubular turbine broad and excellent application prospect. Meanwhile, the tubular turbine can also be used as fish-friendly turbine owing to its features of small blade number and low passage pressure []. With advantage of large flow capacity, smooth tubular passage and superior performance, bulb turbine is the most widely used tubular turbine type in large low head hydropower plant. However, bulb turbines for less than m prototype head, due to its relatively large head loss proportion to net head and peculiar cavitation similarity problem, have around more attention from researchers and engineers. Wang [] () researched hydraulic performance optimization of a bidirectional bulb turbine for tidal power plant; W Yang Y Wu [3] et al () proposed a modified method for runner blades, combining with Computational Fluid Dynamics (CFD) method to improve a H r =5.8m lowhead bulb turbine; LI Fengchao, FAN Honggang [4] et al () developed full three-dimensional sealing design for tubular conical space guide vanes based on optimization technique, using a full three-dimensional inverse model; Takashi Kubota, Tasashi Tsukamotao [5] (99) studied scale effect Published under licence by Ltd

3 6th IAHR Symposium on Hydraulic Machinery and Systems on cavitation of low head bulb turbine; and Arno Gehrer, Helmut Benigni [6] (4) carried out transient numerical simulation for a n q =9min - bulb turbine. Generally, there are two ways to research turbine s performance, among which one is model test and the other numerical simulation. For model test, with limitation of bulb body size, the model unit cannot be the same as the prototype arranging generator inside the bulb body, but need specialized shafting structure that transfers shaft torque to the external generator, so the test accuracy is affected. For numerical simulation, three-dimensional unsteady flow simulation is still needed to solve transient flow phenomenon such as flow separation, rotating stall and Rotor-Stator Interaction, etc., though the steady simulation can already accurately predict hydraulic characteristics near the rated operating point. In order to study performance of bulb turbine suitable for low prototype head, model test and numerical simulation are adopted on a n q =3.min - bulb turbine. Firstly, a model bulb turbine and its bulb body was developed and manufactured; then model tests including energy, cavitation and pressure fluctuation tests were carried out on the horizontal-shaftturbine test rig; and finally, threedimensional unsteady turbulence numerical simulation was performed to analysis internal flow of the model turbine.. Turbine model design.. Model turbine geometry Hydraulic model of bulb turbine was developed. Its sketch is shown in Figure, and parameters are summarized in Table. The model has bulb ratio of.6, runner blade number of 3 (Z r =3) and guide vane number of 6 (Z g =6). Runner is mounted in spherical chamber, with hub ratio of.3. B D b D h D Figure. Sketch of bulb turbine hydraulic model Table. Parameters of model turbine Parameters Symbol Value Runner chamber shape - Spherical Bulb ratio D /D.6 Hub ratio D h /D.3 Height at inflow B /D.9 Height of guide vane b /D.35 Runner blade number Z r 3 Guide vane number Z g 6 Cone angle of guide vane α/ 6.. Model unit According to model turbine geometry, turbine unit for model tests was manufactured. Before the manufacture, with purpose to accurately measure shaft torque, bulb body with hydrostatic bearing structure had been designed on the basis of flow passage dimension. Photograph of the tested turbine is given in Figure. Test section includes six components, that is, inflow channel, bulb body, conical guide vane, runner chamber, coned pipe and outflow channel. In order to observe flow pattern or cavitation phenomenon at runner outlet, cone-shaped pipe is made of transparent material. 3. Model test method Model tests were carried out on test rig TP, a high-accuracy hydraulic machinery model universal test rig locating in IWHR. The test rig is a close-circuit system, equipped with high accuracy original calibration and test system, as shown in Figure 3. This test stand is mainly used to test horizontal-shaft

4 6th IAHR Symposium on Hydraulic Machinery and Systems low head turbine model. Test stand s maximum model test head is m, with discharge range of ~.5 m 3 /s. Total error for model efficiency test is better than ±.% in this system. Tests were performed under condition of Reynolds number Re u.5 6, which satisfies test requirement specified in IEC code 693 [7]. Model tests consist of energy test, cavitation test and pressure fluctuation tests. Blade angle Φ was adjusted at a total of 6 blade angles from to 35. At first, energy characteristic was tested at different blade position. According to single blade angle test results, the on-cam operating points were obtained, furthermore, comprehensive characteristic curves were drawn. Pressure fluctuations were tested at on-cam operating condition, whose data was analyzed using FFT method to get dominating amplitudes and corresponding frequency.. Low pressure tank; 5.Electromagnetic valve for discharge adjusting;. Dynamometer; 6.Electromagnetic flowmeter; 3. Bulb turbine model; 7.main pump; 8. Tank for pressure stabilization 4. High pressure tank; 8 Figure. Photograph of the tested bulb turbine Figure 3. Close-circuit test system for horizontal-shaft turbine 4. Numerical simulation method 4.. Domains and Grids Choose the optimum operating point for numerical simulating. As is presented in Figure 4, the full passage was divided into 4 domains, that is, Inflow, Guide Vane (GV), Runner and Outflow domains. Hexahedral structural mesh was generated for domains of Guide Vane and Runner, and tetrahedral unstructural mesh for domain Inflow and Outflow. Total grid elements number of the full-passage is 8.8 million, and single component grid numbers are listed in table.. Inflow;. Guide Vane; 3. Runner; 4.Outflow Figure 4. Computational domains Table. Domain grid number Inflow GV Runner Outflow Numerical method and boundary conditions Three-dimensional transient turbulence numerical simulation was conducted on the tested bulb turbine, adopting Reynolds-averaged Navier-Stocks control equations and RNG k-ε turbulence model. The solver of simulation is commercial CFD code, Ansys CFX.. Advection scheme was set as high resolution, and convergence residue as -4. Steady flow simulation was computed firstly, and then used as initial value for unsteady simulations. Total calculation time of unsteady simulation is.6 rotation cycle (.6 T). Boundary conditions were set as follows: static pressure condition was set for inlet, mass flow rate boundary conditions for outlet, and no-slip condition for walls. Domain Runner is a rotating domain and the others are stationary domains. Rotating domain connects to stationary domains by Transient Rotor Stator interfaces. 3

5 ETAM (%) 6th IAHR Symposium on Hydraulic Machinery and Systems 5. Results and analysis 5.. Hydraulic performance 5... Comprehensive characteristics. Figure 5 gives best efficiency test results. In this figure, three on-cam operating curves near optimum blade angle (Φ=, 5, 3 ) are shown, and the discrete data point is CFD result. Table 3 gives the contrast of experimental and numerical result at optimal operating point. It can clearly be seen from the figure and table that: () CFD results are in good agreement with experimental results, with error less than.%; () Optimum efficiency η M of bulb turbine is up to 9.7% measured by model test, while 9.8% calculated by CFD method; (3) The optimum operating point, with characteristic parameter of n o =65.54 r/min versus Q o =.93 m 3 /s, appears at position of blade angle Φ= 5 together with guide vane opening a = Head loss. Head loss ΔH of single domain is calculated based on CFD result. Relative hydraulic loss ΔH/H net of various stationary parts is shown in table 4. The table indicates, head loss in Outflow domain accounts for the largest proportion up to 56.8% of total loss. 94 Table 3. Optimum operating point parameters Exp. Result Phi= deg. Phi=5 deg. Phi=3 deg. CFD Result n (r/min) Figure 5. Optimum efficiency test curves Units EXP. CFD Error n o r/min % Q o l/s % η M % % Table 4. Component relative head loss Inflow GV Outflow.4%.74%.44% 5.. Pressure fluctuation and analysis Analyze pressure fluctuation result at optimum operating point. Because pressure fluctuation data was acquired at off-optimum points (n =5 r/min and 8 r/min) rather than optimum point (n =65.5 r/min) at the optimum guide vane opening (a =6 ), herein pressure fluctuation data at the optimum operating point is obtained by CFD method. As a comparison, data of those two off-optimum operating points is also given. Figure 6 shows the pressure fluctuation spectrum, and table 5 gives the first 3 dominating frequencies and their amplitude. Pressure fluctuation parameters are processed into normalized form. In order to display clearly, only data of 3 cycles is given in time domain, while ~6 times rotation frequency data is given in frequency domain. The contrast shows that: () dominating frequency of 3 times rotation frequency is get at all 3 investigated operating points, which correspondents to runner blade number of Z r =3; () pressure fluctuation amplitude at optimum operating point is computed as.36% relative to net head by CFD, this calculated result is reasonable because the value is in the same order of magnitude as test results, but a little less than the two nonoptimum values; (3) dominating frequency of about 4 times frequency appears in model test results, this number is half of runner blade number multiplied by guide vane number (/ Z r Z g ), which reveals rotor-stator interaction between guide vane and runner that propagates downstream to cone pipe. 4

6 H/H (%) Amplitude (%) 6th IAHR Symposium on Hydraulic Machinery and Systems n=8. rpm, EXP. -4 n=5. rpm, EXP. n=65.5 rpm, CFD THETA/36 (REV).8.4. n=8 r/min, EXP. n=5 r/min, EXP. n=65.5 r/min, CFD Frequency/fn (-) Figure 6. Pressure fluctuation spectrum. a) time domain (left), b) frequency domain (right) Table 5. Normalized dominating frequency and amplitude of pressure fluctuation n Parameter st -order nd -order 3 rd -order 8. r/min, f/f n EXP. Amp. (%) r/min, f/f n EXP. Amp. (%) r/min, f/f n CFD Amp. (%) Distribution of flow field Flow field in GV and runner domains. Figure 7 gives the flow field in GV and Runner domains, in which (a) shows flow patterns whose parameters are circumferential averaged on axial plane, and (b) shows pressure distribution on span plane (Span=.5) from blade-to-blade view. Pressure coefficient C p and dimensionless meridional velocity v m,av are defines as: C p =(p-p Outlet )/ (/ρu ), and v m,ave = v m /(gh).5. Along flow direction, internal flow in these two areas is smooth, pressure distributes uniform and no mutation region appears. The pressure coefficient C p is in range from -.65 to.4, and normalized meridional velocity v m,ave is about.5 in the regions Flow pattern in outflow domain. Because largest head loss was found in Outflow domain, special attention is pay on this domain. Flow pattern in Outflow domain is shown in Figure 8, which was extracted from slice of plane X=, where (a) is velocity distribution at.6t, and (b) is pressure contour at various time in one rotation period. Figures show that: () Seeing from velocity distribution, flow pattern in Outflow channel looks a little uniform, but flow separation occurs in near wall region (region B and C), and secondary flow appears in downstream near hub head and outlet position (region A~D), which is the reason for largest loss proportion in this component; () within one rotation cycle from pressure contour, periodicity displayed at the moment of /3T, /3T and.t, which correspondent to dominating frequency of 3 times rotation frequency found in pressure fluctuation analysis; (3) pressure distribution at inlet area of Outflow pipe looks symmetric about rotating axis, but the symmetry gradually disappears in downstream due to the complex flow phenomenon. 5

7 R/R (-) R/R (-) R/R (-) R/R (-) R/R (-) R/R (-) 6th IAHR Symposium on Hydraulic Machinery and Systems Level Level 3Cp: Cp: Z/R (-) Z/R (-) Z/R (-) v m,ave v m,ave Z/R (-) Z/R (-) Z/R (-) Figure 7. Flow pattern in GV and Runner domains (n =65.54 r/min, Q =.93 m 3 /s). a) Circumferential mean pressure contour (Top-left) and velocity vector (Top-right) on axial plane, b) Pressure contour on Span=.5 from blade-toblade view (Bottom) X Y Z A C B D t=/3t t=/3t-/t B A D t=/3t t=/3t+/t C t=/3t+5/t t=.t Figure 8. Flow pattern in Outflow domain on plane X= (n =65.54 r/min, Q =.93 m 3 /s). a) velocity distribution at t=.6t (Left), b) pressure contour of one rotation period (Right) 6

8 Cp (-) Flow Angle (Deg.) 6th IAHR Symposium on Hydraulic Machinery and Systems Runner dynamic behavior. Choose the optimum operating point to analysis dynamic behavior on runner, as shown in Figure 9 and, among which Figure 9 is curves of hydraulic parameters, and Figure demonstrates the dynamic behavior..5.5 Pressure Side 75 5 Alpha Beta Suction Side Streamwise (-) Span Normalized (-) Figure 9. Hydraulic parameter curves (n =65.54 r/min, Q =.93 m 3 /s). a) blade loading on Span=.5 (Left), (b) flow angle of runner leading edge from hub to shroud (Right) P.S. S.S. Figure. Runner dynamic behavior (n =65.54 r/min, Q =.93 m 3 /s). a) velocity distribution from blade-to-blade view (Left), b) pressure contour In Figure 9, abbreviations of P.S. and S.S. are short for pressure side and suction side, respectively. Figure (a) is blade loading curve on span=.5 surface from blade-to-blade view; and figure (b) reveals flow angle at runner leading edge from hub to shroud, where Alpha is absolute flow angle, and Beta is relative flow angle. As curves show in the figure, pressure coefficient on blade surface is between -.5 to.5, which changes gently along streamwise, but relative flow angle near hub in spanwise direction drop sharply, where the incidence angle is large, which leads to more hydraulic loss. Figure (a) presents velocity vector distribution on Span=.5 surface from blade-to-blade view, and Figure (b) pressure contour on runner surface. At the optimum operating point, flow exhibits a uniformity and stability. Unsteady flow phenomenon, such as vortex, flow separation or rotating stall, does not appear, only a little flow impact happens at blade leading edge. Pressure on blade surface is of homogeneous. The runner dynamic behavior reflects perfect flow characteristics. 7

9 6th IAHR Symposium on Hydraulic Machinery and Systems 6. Conclusions In this paper, performance of a low-head bulb turbine is studied using experimental and numerical methods. Conclusion comes out as follows: () The bulb turbine has good hydraulic performance. The optimum model efficiency is up to 9.7% by model test, at the optimum operating point of n o =65.54 r/min versus Q o =.93 m3/s; () Through pressure fluctuation analysis, the numerical results coincide with experimental results well, so the numerical method is effective. The relative pressure fluctuation amplitude is only.36% at optimum operating point; (3) Head loss in Outflow channel occupies more than 5% proportion of turbine hydraulic loss, where flow separation and secondary flow occur severely; (4) Bulb turbine model has fine hydraulic characteristics, with a good prospect of promotion. Acknowledgments This work was supported by China Institute of Water Resources and Hydropower Research (IWHR 7) and BITC Company (SY--ZY-4). Nomenclature C p Pressure coefficient (=(p-p Outlet )/ (/ρu )) D Diameter, [m] f frequency, [Hz] g acceleration of gravity, [ms - ] H net Net head, [m] ΔH Head loss, [m] n Rotating speed, [r min - ] n Unit speed (=nd/h.5 ), [r/min] n q Unit specific speed (=nq.5 /H.75), [r min - m.5 s.5 /m.75 ] Q Q Re u Z v m,ave η M Φ Discharge, [m 3 /s] Unit discharge (=Q/(D H.5 )), [m 3 /s] Reynolds number (=u D / ) Velocity, [m/s] Blade or vane number Normalized averaged meridional speed (=v m /(gh).5 ) Efficiency (=Tω/ρgQH) Runner blade angle, [ ] Fluid Density References [] Odeh M and Sommers G Hydropower & Dams [] Wang Z W, Yang X S and Xiao Y X J. of Drainage and Irrigation Machinery Engineering 8( 5) (in Chinese) [3] Yang W, Wu Y and Liu S A modification method on runner blades in a bulbturbine 5th IAHR Symp. on Hydraulic Machinery and Systems (Timisoara, Romania, September -4, ) ///3 [4] Li F C, Fan H G, Wang Z W and Chen N X J. Tsinghua Univ (Sci & Tech) 5() (in Chinese) [5] Takashi K and Tasashi T 99 Fuji Electric Rev. 36() [6] Arno G, Helmut B, Martin K 4 Unsteady simulation of the flow through a horizontal-shaft bulbturbine nd IAHR Symp. on Hydraulic Machinery and Systems (Stockholm, Sweden, June 9 - July, 4) [7] IEC Hydraulic Turbines, Storage Pumps and Pump-Turbines - Model Acceptance Test 8