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The 11th Asian International Conference on Fluid Machinery and Paper ID: AICFM_TM_012 The 3rd Fluid Power Technology Exhibition November 21-23, 2011, IIT Madras, Chennai, India
Original Paper
Internal Flow Analysis on a Micro Cross-Flow Type Hydro Turbine at Very Low Specific Speed Range
Kiyoshi Kokubu1, Sung-Woo Son2, Toshiaki Kanemoto3 and Young-Do Choi4
1Technical Department, Tanaka Suiryoku, 5-18-34, Hibarigaoka, Zama, Kanagawa, 252-0003, Japan, [email protected]
2Department of Mechanical Engineering, Graduate school, Mokpo National University 1666 Yeongsan, Chonggye, Muan, Jeonnam, Mokpo, 534-729, Korea, [email protected]
3Faculty of Engineering, Kyushu Institute of Technology Sensui 1-1, tobata, kitakyushu 804-8550, Japan, [email protected]
4Department of Mechanical Engineering, Mokpo National University 1666 Yeongsan, Chonggye, Muan, Jeonnam, Mokpo, 534-729, Korea, [email protected]
Abstract
Renewable energy has been interested because of fluctuation of oil price, depletion of fossil fuel resources and environmental impact. Amongst renewable energy resources, hydropower is most reliable and cost effective way. In this study, to develop the new type of micro hydro turbine which has very low specific speed, simple structure and high efficiency, Eco Cross-Flow Hydro Turbine (ECFT) is proposed. ECFT can be used at very low specific speed range of hydropower resources, such as very high-head and considerably small-flow rate water resources. CFD analysis on the internal flow and performance characteristics of the turbine is conducted to obtain a basic data for a new design method of ECFT.
Keywords: Micro hydropower, Eco cross flow turbine, Very low specific speed range, Performance, Internal flow.
1. Introduction
For hundreds of years, humankind has heavily relied on fossil fuel to generate heat and electric power. However, this energy source is gradually depleting. Burning of fossil fuel leads to pollution and many negative environmental impacts. Moreover, fluctuating/rising oil prices, increases in demand, supply uncertainties and other factors have led to increased calls for alternative energy sources. Therefore, there has been considerable interest recently in the topic of renewable energy. This is primarily due to concerns about environmental damage, especially acid rain and global warming, resulting from the burning of fossil fuels.
Among the renewable energy resources, micro-hydropower systems are more preferable for stand-alone/residential electric generation than other renewable resources. This is because micro hydropower systems offer a stable, economical and renewable source of electricity that uses proven and available technologies.
There are many available micro hydropower resources surrounding us for extraction, such as rivers, municipal water supply systems of towns, drainage water from houses, irrigation canals and so on. Several kinds of micro hydropower turbines have been developed and used to extract energy from micro hydropower resources in the last century. Different kinds of turbines require different conditions for operation. In the case of very high head and critical low flow rate range of micro hydropower resources, impulse turbine has long been used widely in the specific speed range of ns < 70[min-1, kW, m]). Common impulse turbines are Pelton, Turgo and cross-flow turbines. Especially, in the specific speed range of ns ≦ 40, Pelton turbine has been used conventionally as shown in Fig. 1 [1].
However, recently, there exists strong request of developing a very low specific speed turbine which has very simple structure and high efficiency to be used at the very low specific speed range of micro hydropower resources. Moreover, widening the operational range of a micro hydro turbine system is essential to expand the introduction of the micro hydropower system.
Therefore, this study is aimed to develop the new type of micro cross-flow hydro turbine which has very low specific speed, simple structure and high efficiency. As the cross-flow type hydro turbine has been used in the range of middle specific speed because of simple structure and wide operational range of flow rate [2-11], this study tries to adopt the turbine in the range of very low specific speed as well to widen the operational range of the turbine. Eco Cross-Flow Hydro Turbine (ECFT) is proposed to
Accepted for publication. Corresponding author: Young-Do Choi, Professor, [email protected]
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apply the turbine operation range to the very low specific speed of ns ≦ 40, in which conventional cross-flow turbine has not been used in the range but pelton hydro turbine has been used conventionally. CFD analysis on the internal flow and performance characteristics of the turbine model is conducted to obtain a basic data for a new design method of ECFT.
2. Turbine Model and Numerical Method
2.1 Eco Cross-Flow Hydro Turbine Model Figure 2 shows the schematic view of ECFT model which is proposed firstly by present study. As seen from Fig. 2, the inlet
pipe and nozzle are vertically installed. The number of runner blade is Z=30 and the diameter of runner is d=250mm. Inlet and outlet angles of the blade are α=30° and β=87°, respectively. The widths of nozzle, runner and runner chamber are all same, b=17mm, which includes tip clearance c=2mm between front wall surface and tip of the components. From the design data of effective head H =19m, flow rate Q = 0.017 m3/s and rotational speed n = 1000 min-1, relatively very narrow internal flow passage is designed to adjust flow rate of the turbine with the specific speed of ns=nP1/2/H5/4=40. Air suction hole is not installed in this study because only hydraulic performance of the turbine model without air layer in the flow passage is investigated. As the main purpose of this study is to confirm the effects of guide vane angle and inner guider angle on the internal flow and performance of the turbine model, internal passage of the turbine model is designed very simply.
Three kinds of guide vane angle θgv and two kinds of inner guider angle θig are selected including no inner guider case for CFD analysis as shown in Fig. 3. Moreover, Table 1 shows test cases with the variation of guide vane angle and inner guider angle. Rotational speed range between 300 to 1500min-1 is selected to examine the turbine performance by CFD analysis. The sections of Stages 1 and 2 in Fig. 3 are determined by the flow passage of main flow in the turbine.
2.2 Numerical analysis method For the numerical analysis on the turbine internal flow and performance by the variation of the turbine structural configuration,
Fig. 4 Numerical grid system
Fig. 3 Variation of guide vane angle
and inner guider angle
Flow
Guidevane
Shaft
Innerguider
Runner
Fig. 2 Schematic view of ECFT model
Fig. 1 Turbine types and their specific speeds
Table 1 Test cases of turbine model
Case
Guide vane angle θgv [°]
Inner guider angle θig [°]
A 0 0
B 5 0
C 15 0
D 15 12
E 15 no inner guider
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a commercial code of ANSYS CFX [12] is adopted. The grid node number of about 7x105 has been used as shown in Fig. 4. Fine tetrahedral-grids are employed to ensure relatively high accuracy of calculated results for the turbine model.
From the validation test of the numerical model using turbulence models available from the present numerical code, k-ω SST model shows reasonable turbine performance and good convergence of calculation. Therefore, k-ω SST model is used for single-phase flow calculation. Constant pressure at inlet and averaged flow rate at outlet are the used boundary conditions. CFD analysis for the test cases by the variation of turbine structure are conducted under the conditions of steady state and working fluid of water
at 25°C.
3. Results and Discussion
3.1 Performance Curve Figure 5 shows performance curve of the turbine model by the rotational speed variation. Moreover, comparison of the
efficiency ratios by the variation of guide vane angle and inner guider angle at the best efficiency point is conducted. The calculated results reveal that Case C, which is composed of the guide vane angle θgv =15°and the inner guider angle θig
=0°, has maximum efficiency among the arrangement of the turbine models. However, influence of the inner guider on the efficiency is negligible. From these results, it is conjectured that present ECFT model is designed by conventional method. Therefore, hydraulic and friction losses in the turbine model should be estimated exactly and quantitatively for the new design method.
Fig. 5 Performance curve of ECFT model by CFD analysis
3.2 Velocity Vectors and Velocity Distribution Figures 6 and 7 show velocity vectors in the internal flow field of the turbine model. The velocity vectors are expressed using
absolute velocity vectors in the flow field of turbine model. It is clear from Fig. 6 that velocity in the internal flow field is dominated directly by the guide vane angle. Velocities in the
nozzle and internal area of the runner change obviously by the guide vane angle. As a whole, fluid velocity becomes accelerated along the contracted nozzle passage from the inlet. After passing through the runner blade passage at Stage 1, cross-flow within the runner center gains accelerated velocity once more and then the flow enters into the inlet of Stage 2.
While, in the region of recirculation flow which is indicated as Stage 3in the Fig. 3, there exists a large recirculating flow. According to the operation condition of guide vane angle, the shape of recirculating flow region within the blade passage changes as well. However, by the variation of the inner guider angle, there is not so considerable change in the flow patterns as shown in Fig. 7.
Figure 8 indicates the averaged velocity distribution in the blade passages at the inlet of Stages 1. Tangential velocity vθat the inlet of Stage 1 is varied with the guide vane and inner guider angles. Especially, Case C (θgv =15°, θig =0°) shows relatively higher tangential velocity compared with those of the other cases. This result confirms that the angular momentum of the Case C by relatively higher fluid tangential velocity is absorbed into the runner blade and the angular momentum changes to larger output power in comparison with the other cases. The result implies that the guide vane angle gives remarkable effect on the internal flow of the cross-flow turbine and thus, performance of the turbine is closely dependent on the guide vane angle. However, influence of the inner guide angle on the tangential velocity at Stage 1 is not so large. Moreover, the overall value distribution of tangential velocity by the guide vane angle at Stage 1 inlet in Fig. 8 is proportional to the efficiency variation as shown in Fig. 5.
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(a) θgv=0° (b) θgv=5° (c) θgv=15°
Fig. 6 Velocity vectors by the variation of guide vane angle (N=1000min-1, θig=0°)
(a) θig=0° (b) θig=12°
Fig. 7 Velocity vectors by the variation of inner guider angle (N=1000min-1, θgv=15°)
0.00
0.50
1.00
1.50
2.00
2.50
3.00
0.00 0.10 0.20 0.30 0.40 0.50 0.60 0.70 0.80 0.90 1.00
1234
Non-dimensional passage length
Vθ/U
1
Case A (θgv=0˚, θig=0˚)
Case B (θgv=5˚, θig=0˚)
Case C (θgv=15˚, θig=0˚)
Case D (θgv=15˚, θig=12˚)
0
1
5
Fig. 8 Velocity distribution by the variation of guide vane angle and inner guider angle at Stage 1 inlet(N=1000min-1)
3.3 Pressure distribution in the turbine passage From the results of the Figs. 6 to 8, it is obvious that the flow velocity passing through the runner passage gives considerable
effect on the output power, which is one of the principal characteristics of impulse turbine. While, static pressures by the guide vane and inner guider angles are investigated in order to examine the pressure change by the component’s configuration.
Figure 9 shows the static pressure contours within the turbine internal flow filed with the variation of guide vane angle. Turbine inlet pressure decreases along the nozzle passage but the pressure at the nozzle outlet is almost uniform. The fluid pressure passing through the passage of runner blades at Stage 1 drops rapidly. From this result, it is assumed that the fluid pressure passing through the passage of runner blade is taken by the runner blades and the pressure changes to output power. In addition, very low pressure region is located at the leading edge of guide vane suction side as shown in Fig. 9(c), at which there is high possibility of cavitation occurrence.
Figure 10 presents pressure contours with the variation of inner guider angle. There is no considerable difference of the pressure contours in the flow passage. Therefore, it is conjectured that radius of the circular arc should be optimized in order to fit the inner guider angle to the outlet angle of the runner blade at Stage 1 and to the inlet angle of the runner blade at Stage 2.
Figure 11 shows averaged pressure distributions on the surface of runner blades by guide vane angle (Cases B and C) at Stages 1 and 2. With the variation of the guide vane angle, there is no significant changes for the area surrounded by the pressure coefficients on the blade pressure and suction sides. This result implies that performance of ECFT in the range of very low specific speed is influenced by mainly flow rate, which is controlled by guide vane angle. Therefore, optimum flow rate at the design point should be determined.
(a) θgv=0° (b) θgv=5° (c) θgv=15°
Fig. 9 Pressure contours by the variation of guide vane angle (N=1000min-1, θig=0°)
(a) θig=0° (b) θig=12°
Fig. 10 Pressure contours by the variation of inner guider angle (N=1000min-1, θgv=15°)
6
-2.5
-2.0
-1.5
-1.0
-0.5
0.0
0.6 0.7 0.8 0.9 1.0
2
1
Case B (?gv=5°, ?ig=0°), Stage 1
Case C (?gv=15°, ?ig=0°), Stage 1
Cp
r/r2 (a) Stage 1
-3.5
-3.0
-2.5
-2.0
-1.5
-1.0
-0.5
0.0
0.6 0.7 0.8 0.9 1.0
2
1
Case B (?gv=5°, ?ig=0°), Stage 2
Case C (?gv=15°, ?ig=0°), Stage 2
Cp
r/r2 (b) Stage 2
Fig. 11 Pressure distribution around the surface of runner blade by guide vane angle (N=1000min-1)
7
-10
-5
0
5
10
15
20
25
1
2
3
4
Case A (θgv=0˚, θig=0˚)
Case B (θgv=5˚, θig=0˚)
Case C (θgv=15˚, θig=0˚)
Case D (θgv=15˚, θig=12˚)
Stage 1
Tor
que[
Nm
]
Stage 2 Stage 3 Total
Fig. 12 Output torque analysis by CFD
3.4 Output Torque Analysis and Loss by Recirculating Flow Figure 12 presents the divided local output torques at each region by the guide vane and inner guide angles. The local output
torque is evaluated from the calculated local output torque at each region of runner blades. Remarkable output torque loss occurs in the recirculation flow region of Stage 3 [13]. Especially, Cases A and B shows considerable values of consumed torque by the recirculating flow.
Therefore, air layer should be allocated in the runner passage to improve the turbine performance. The existence of air layer improves the turbine efficiency not only from the suppression of collision loss between the runner passage flow and shaft, but also from the elimination of loss by recirculating flow in the runner passage.
According to the inner guider angle, variation of the output torque at Stages 2 and 3 is obviously recognizable. Higher output torque is generated by the Case C at Stage 2 in comparison with the Case D, but higher output torque loss at Stage 3 by the Case C. Therefore, total output torque of two Cases C and D does not differ so much. This result also implies that inner guide angle can give considerable effect on the performance improvement by the angle optimization.
Moreover, from the Fig. 12, it is clear that local output torques at Stage 1 are almost controlled by guide vane angle but local output torques at Stages 2 and 3 are strongly governed by inner guide angle.
4. Conclusions
Analysis on the internal flow and performance a micro cross-flow type hydro turbine at very low specific speed range is conducted with the variation of the guide vane and inner guider angles by CFD. From the results of the present study, the following conclusions are obtained.
1. Higher angular momentum with relatively higher fluid tangential velocity by Case C by is absorbed into the runner blade and the angular momentum changes to larger output power in comparison with the other test cases. The result implies that the guide vane angle gives remarkable effect on the internal flow of the cross-flow turbine and thus, performance of the turbine is closely dependent on the guide vane angle and flow rate. Moreover, overall value distribution of tangential velocity by the guide vane angle at Stage 1 inlet is proportional to the efficiency variation of the turbine.
2. Local output torque at Stage 1 is almost controlled by guide vane angle, but local output torques at Stages 2 and 3 are strongly governed by inner guider angle. Therefore, optimization of shape and angle of the guide vane and inner guider is very important in order to improve the turbine performance.
3. Air layer should be allocated in the runner passage to improve the turbine performance. The existence of air layer improves the turbine efficiency not only from the suppression of collision loss between the runner passage flow and shaft, but also from the elimination of loss by recirculating flow in the runner passage.
Nomenclature
H Effective head [m] α Inlet angle of runner blade [°] Q Flow rate [m3/s] β Outlet angle of runner blade [°] Z Number of blades θgv Guide vane angle [°] b Width of runner [mm] θig Inner guider angle [°] d Diameter of runner [mm] η Efficiency [%] n Rotational speed [min-1] ηmax maximum efficiency [%]
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References
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