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Influence of air supply on the performance and internal flow characteristics of a cross flow turbine
Renewable Energy xxx (2014) 1e8
Contents lists available at ScienceDirect
Renewable Energy journal homepage: www.elsevier.com/locate/renene
Influence of air supply on the performance and internal flow characteristics of a cross flow turbine Zhenmu Chen a, Young-Do Choi b, * a
Graduate School, Department of Mechanical Engineering, Mokpo National University, Muan-gun 530-729, Jeollanam-do, Republic of Korea Department of Mechanical Engineering, Institute of New and Renewable Energy Technology Research, Mokpo National University, Muan-gun 530-729, Jeollanam-do, Republic of Korea
b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 12 March 2014 Accepted 8 August 2014 Available online xxx
This study attempts to improve the efficiency of a given type of cross flow turbine by supplying air from air suction holes. A newly developed air supply method is adopted. CFD analysis of the cross flow turbine is carried out to investigate the performance and internal flow characteristics of the turbine in detail. The air layer prevents shock loss between water flow and axis and suppresses recirculation flow in the runner passage. Hence, it is necessary to measure the amount of air layer in the runner passage and examine its effect on the performance of the cross flow turbine. The result shows that the turbine efficiency has improved more as the newly developed air supply method is applied effectively. © 2014 Elsevier Ltd. All rights reserved.
Keywords: Cross flow turbine Performance improvement Internal flow Air layer effect
1. Introduction Recently, the environmental problems such as global warming, pollution problems and soon, have been considered as inevitable issue. Therefore, the necessity of the use of renewable energy as one of the clean and sustainable natural energy resources has become high [1]. In addition, the cross flow turbine has a great potential for small hydropower development for both developed and undeveloped countries. Amongst renewable energy resources, the cross flow turbine is relatively reliable and cost effective with a very simple structure. The nozzle and the runner are made by processing steel plate and the blades are made by processing steel pipe giving it a simple design and manufacture. Therefore it can save more manufacturing cost and maintenance cost. Most present studies of cross flow turbines have optimized the configuration of turbines to improve the efficiency by experiment. The turbine design is based on Banki's one-dimensional analysis method [2]. Fiuzat et al. [3] has carried out an experiment to identify the contribution of each one of the cross flow turbine stages to the power output generation. Khosrowpanah et al. [4] and Fukutomi et al. [5] have tried to improve the performance of the turbine by changing the number of blades, the angles of water entry to the runner and the inner-to-outer diameter ratios. Kokubu et al.
* Corresponding author. E-mail addresses: [email protected] (Z. Chen), [email protected] (Y.-D. Choi).
[6] has improved the performance by changing the structure of turbine. On the other hand, Choi et al. [7,8] has investigated the effect of air layer located in the turbine runner passage. The air layer plays a significant role to improve the turbine performance. However, the internal flow in the turbine passage is complex, especially with the air layer. The flow with air layer has not been analyzed deeply enough. Therefore, this study attempts to improve the efficiency of the cross flow turbine by supplying air into the turbine chamber passage. Experiments are performed in order to measure the efficiency of the cross flow turbine by the conventional method to supply air. CFD analysis on the performance and internal flow of turbine is conducted in order to investigate the air layer effect on the turbine performance by newly developed method. 2. Test turbine and numerical methods 2.1. Experiment setup and cross flow turbine Fig. 1 shows the cross flow turbine for field test. The turbine is installed and operated for commercial production of electric power. The number of the runner blades is Z ¼ 30 and the diameter of the runner is d ¼ 340 mm. The inlet and outlet angles of the blade are b1 ¼ 34 and b2 ¼ 92.5 , respectively. The widths of the nozzle, runner and the draft tube are all the same, the width is b ¼ 500 mm. The guide vane angle ag ranges from 0 to 32 , which means the turbine starts from completely closed to open states, respectively. The design point of the present test turbine at the best efficient
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Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
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Z. Chen, Y.-D. Choi / Renewable Energy xxx (2014) 1e8 Table 1 Cases of air suction hole methods by experiment and CFD analysis. Analysis cases
Air suction Hole A
Air suction Hole B
Experiment CFD
Ab 0 Ab Ab
0 0 0 Bb
a b
Case 1a Case 2a Case 3a
Guide vane angles ag: 15 , 18 , 20 , 22 , 25 . Air flow rate is determined for each guide vane angle (refer to Table 2).
every case. Table 2 shows the water flow rate and air flow rate ratio for each guide vane angle in detail. 2.2. Numerical methods
Fig. 1. Cross flow turbine for field test.
point is H ¼ 20 m for the effective head, n ¼ 530 min1 for rotational speed. Fig. 2 shows the efficiency curve of test turbine model by experiment [8]. The experiment is performed under the condition of fixing head (H) of the turbine, and only the air suction valve is open to supply a constant air flow rate. The result indicated that with increasing the guide vane angle, the efficiency increases as well at first, after it reaches the maximum efficiency, it decreases again slightly. Moreover, uncertainty in the measurement is indicated as error bars in the efficiency curve as well. In addition, Table 1 shows the cases of air suction methods by experiment and CFD analysis. CFD analysis for the Cases 1, 2 and 3 are conducted by variation of guide vane angle ag. For the Case 1, the working fluid is only water. There is no air supply into the turbine chamber. The condition of Case 2 is the same as the experiment which is only air suction Hole A is adopted (the conventional air supply method). Case 3 adopts two air suction holes of a conventional type by Hole A, as well as a type by Hole B which is new method of air supply as shown in Fig. 3. In Case 3, the air flow rate of Hole A is correspondingly the same with that in Case 2 at each guide vane angle. The water flow rates from inlet pipe are all the same in the experiment and CFD analysis. The air flow rate from the air suction hole is determined for each guide vane angle in
Fig. 2. Efficiency curve with uncertainty error bars of test turbine model by experiment.
A commercial code of ANSYS CFX is adopted for the numerical analysis on the turbine performance and internal flow [9]. Fig. 3 shows the schematic view of the turbine model for CFD analysis. As the flow in the turbine can be assumed to be uniform to the direction of main stream, which means no flow velocity component in the direction to the runner depth, the turbine model adopted for the CFD analysis is conducted by two-dimensional shape for considering the calculation time and computer capacity. The air suction Holes A and B are installed on the casing chamber upper wall and at the back of the draft tube, respectively. The area of runner passage can be divided into three regions by the flow patterns in the runner passage. Stage 1 obtains first output power and Stage 2 takes second output power. However, Region 1 consumes output power by hydraulic loss. To ensure relatively high accuracy of calculated results, fine hexahedral grids are generated for the turbine model. The grid element number of about 1.6 106 for the whole flow field has been used. Fig. 4 shows the fine hexahedral numerical mesh for whole view of test turbine model and the mesh for runner with high dense grids. Moreover, the value of yþ, which means non-dimensional distance from wall [10,11], is determined to be below 9 for the runner blade passage and below 22 for the other parts of the turbine passage as shown in Table 3. A suitable turbulence model is required for complex flow phenomena that occur on the internal flow. Therefore, the turbulence model dependence test is carried out by using three representative turbulence models: RNG k- ε, k-ε and SST turbulence models. From
Fig. 3. Schematic view of the numerical model.
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
Z. Chen, Y.-D. Choi / Renewable Energy xxx (2014) 1e8 Table 2 The water flow rate and air flow rate ratio for each guide vane angle.
3
Table 3 Non-dimensional wall distance.
Guide vane angle [+]
Water flow rate [m3/s]
Air flow rate ratio [Qair/Qwater]
Sections
yþ value
Hole A
Hole B
15 18 20 22 25
0.3625 0.4250 0.4650 0.5160 0.5624
0.0124 0.0105 0.0100 0.0145 0.0133
0.0289 0.0318 0.0323 0.0320 0.0320
Inlet Runner blade Center of runner Outlet
18 9 21 22
3. Results and discussion the validation test of the numerical model using turbulence models available from the present numerical code, SST model showed reasonable turbine performance and good convergence of calculation in comparison with the other turbulence models as shown in Fig. 5. Therefore, SST model is used for the single-phase flow and two-phase flow calculations. Unsteady numerical analysis is conducted in the CFD analysis for both single-phase and two-phase flow calculations with the effect of gravity. Total pressure at inlet and constant mass flow rate at outlet of the calculation domain are used for the boundary conditions. The boundary condition of velocity at air suction hole is used.
3.1. Performance curves Fig. 6 shows the performance curves of turbine model under the condition of various guide vane angle. The two efficiency curves by experiment and CFD analysis results of Case 2 are agree well with small deviation in the ranges of the partial and excessive load, which makes the methods and the computational results reliable. The inlet water flow increases along with the increase in guide vane angle accordingly. The efficiency of Case 3 shows the maximum value at every guide vane angles, while that of Case 1 shows the minimum value. These results imply that the air supply through air suction Hole B gives considerable effect on the improvement of the turbine efficiency. Moreover, the effect of air suction both from Holes A and B (Case 3) to improve efficiency is obvious at the medium load in comparison with the result of Cases 2, but the effectiveness becomes poor at the partial and excessive loads. 3.2. Air layer effect on the turbine performance To investigate the influence of air layer on the turbine performance, the two-dimensional air layer area in the runner blade passage and center is examined using the results of CFD analysis under the condition of water and air two-phase flow for Cases 2 and 3 as shown in Fig. 7. The figure shows the air volume fraction distributions of Case 3 (with air supply from both Holes A and B) at 20 guide vane angle. The red part (in web version) is the air layer, and the blue part is water. The air layer in the runner blade passage and center in Region 1 plays the role of preventing a shock loss in the runner axis and suppressing a recirculation flow in the runner [7,8]. Hence, introducing air layer in this region can improve the performance of the turbine. Furthermore, as the pressure in the center of the runner is relatively lower than that outside the turbine as shown in Fig. 8, the air is sucked into the center of runner from air
Fig. 4. Fine hexahedral numerical mesh for test turbine model (a)Whole view of numerical mesh for test turbine (b)Fine hexahedral numerical mesh for runner.
Fig. 5. Turbulence model dependence test for the cross flow turbine model (Case 2).
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
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Z. Chen, Y.-D. Choi / Renewable Energy xxx (2014) 1e8
Fig. 6. Performance curves of turbine model.
suction Holes A and B without additional blower or compressor. Therefore, it needs to examine the relation between the turbine performance and the amount of air layer in the runner passage and its center. Fig. 9 shows the best efficiencies at different guide vane angles under the condition of various air layer area. The air layer area is measured in the runner blade passage and center. The efficiency increase is approximately proportional with increasing air layer area in the turbine chamber. Nevertheless, the efficiency growth rate by the guide vane angle of 25 is very small with increasing air layer area. The cases of 20 guide vane angle have the maximum efficiency compared with the others. With air suction from both Holes A and B (Case 3), there is more than 4% efficiency improvement at 15, 18 and 20 guide vane angles from the efficiencies by Case 1. Moreover there are 3.5% and 0.6% efficiency improvement at 22 and 25 guide vane angle by the air supply conditions from Cases 1 to 3, respectively. Fig. 10 presents the air layer area variation by guide vane angle in the runner blade passage and center for Cases 2 and 3. The air
layer area of Case 2 (air supply by Hole A only) gradually decreases along with increasing guide vane angle. Moreover, for Case 3, the air is sucked from both Holes A and B. The air flow rate from Hole A of Case 3 is the same with that of Case 2. Hence, the remaining air layer area is from Hole B which is marked by AB in Fig. 10. The AB is calculated by following equation:
AB ¼ AAB AA
where the AA and AAB represent the air layer area of Case 2 and Case 3, respectively. The AAB is much larger than AA and the overall trend of AB is upward along with the increase of the guide vane angle. Accordingly, the efficiency of Case 3 is higher than that of Case 2 (refer to Figs. 6 and 9) 3.3. Velocity distribution around the runner For the continuous flow condition, the relation between output torque (T) and the tangential velocity are calculated by following equation:
_ qr T ¼ mv
Fig. 7. Air volume fraction (Case 3 at 20 guide vane angle).
(1)
(2)
where r is radius of the runner, m_ is the mass of the water flow, and the vq is the tangential velocity. Therefore, it is very important to estimate the tangential velocity to predict the turbine output power, because output torque and output power of the turbine is derived from the tangential velocity. In order to investigate the internal flow in detail, averaged tangential velocity distribution around the runner at each Stages 1 and 2 is examined. Fig. 11 shows the schematic of runner regions for the investigation of velocity distribution on the runner circumference at Stages 1 and 2. Figs. 12 and 13 show that the averaged velocity at the runner circumferential passage from the suction side to pressure side of the blades at the inlet and outlet of Stages 1 and 2 by Cases 1 to 3 at 20 guide vane angle, respectively. The average data from the tangential velocity values at several runner passages are averaged to one runner passage with a single rotation of runner as shown in Fig 11. The symbols of q1ave: and q2ave: . in the abscissa denote the averaged passage location between the suction to pressure sides of runner blade at the Stages 1 and 2, respectively.
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
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5
Fig. 8. Pressure contours on the runner passage.
The ordinate is defined as velocity ratio (vq/u0), where the vq is the tangential flow velocity at the local radius and u0 is tangential runner velocity by the following equation:
u0 ¼ ru
(3)
difference ratio increases significantly, especially for Case 3. The tangential velocity difference ratio of Case 3 is the highest at Stage 2. The reason lies in the largest air layer area in the runner passage in Case 3. The results imply that the air from air suction Hole B gives considerable effect on the tangential velocity difference of the cross flow turbine at Stage 2, but the effectiveness is poor at Stage 1.
where u is angular velocity, r is the runner radius. From the results of Figs. 12 and 13, the tangential velocity ratio (vq/u0) at inlet is relatively higher in comparison with that at outlet. The difference of tangential velocity ratio at the inlet and outlet is directly proportional to the amount of torque of the water flow through the runner blade passage as shown in Eq. (2), which means it is proportional to the output power simultaneously. The vq/u0 at the outlet of Stage 1 shows lower than 1.0, but the vq/u0 at inlet of Stage 2 shows higher than 1.0, indicating that the water flow gets the kinetic energy again before the water flow enters into the Stage 2. It can be conjectured that due to the low pressure in the center of the runner, the water flow gains the kinetic energy again. Moreover, the air layer in the runner center plays a role of suppressing the collision loss between water flow and the runner axis. Fig. 14 shows the effective tangential velocity difference ratio at 20 guide vane angle. All the velocity difference area values are divided by that of Case 1 at Stage 1 to compare the effective tangential velocity difference ratio, which means the effective velocity difference area ratio of Case 1 at Stage 1 is 1.0. This figure presents that the tangential velocity difference ratio is almost similar each other among the Cases 1, 2 and 3 at Stage 1. However, at Stage 2, with air supply the tangential velocity
where pout is the static pressure at the outlet of the turbine and p is the local static pressure around the surface of runner blade location, r is the radius of runner and u is angular velocity. From Figs. 15 and 16, there are two parts of DCpn and DCpp in the closed area filled with the Cp curve. DCpn is the closed area filled with the Cp curve that the suction side pressure is larger than that of pressure side, which means the output torque generated by this part of pressure difference is negative. However, for the DCpp, it is contrary to that of DCpn, and the output torque made by this part of pressure difference is positive. Therefore, in order to have power to rotate the runner blade, the effective pressure difference (DCp ¼ DCppDCpn) should be positive.
Fig. 9. Best efficiencies by different air layer area.
Fig. 10. Air layer area variation by guide vane angle.
3.4. Pressure distribution around the runner blade The pressure distribution around the surface of runner blades of Stages 1 and 2 at 20 guide vane angle is shown in Figs. 15 and 16. The pressure coefficient (Cp) is calculated by following equation:
Cp ¼
p pout 0:5rðruÞ2
(4)
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
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Z. Chen, Y.-D. Choi / Renewable Energy xxx (2014) 1e8
Fig. 11. Selected runner passages for the investigation of velocity distribution.
Fig. 12. Averaged velocity of Stage 1 at 20 guide vane angle.
Fig. 14. Effective tangential velocity difference ratio at 20 guide vane angle.
Fig. 17 shows the effective pressure difference ratio at 20 guide vane angle. The ordinate is defined as effective pressure difference ratio. All the effective pressure differences (DCp) are divided by that of Case 1 at Stage 1 to compare the effective pressure difference ratio, which means that the effective pressure difference ratio of Case 1 at Stage 1 is 1.0. The equation of effective pressure difference ratio is calculated by following:
where the DCp1 is the effective pressure difference of Case 1 at Stage 1. The effective pressure difference ratio increases rapidly at Stages 1 and 2 with supplying air from air suction Hole B, which means that the runner blades get more output power by the pressure difference. Moreover, the pressure difference ratio of Stage 1 is larger than that of Stage 2 at each case, which means that the pressure transfer to output power at Stage 1 is more than that at
Cpe ¼
DCp DCp1
Fig. 13. Averaged velocity of Stage 2 at 20 guide vane angle.
(5)
Fig. 15. Averaged pressure distribution around the surface of runner blades of Stage 1 at 20 guide vane angle.
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Z. Chen, Y.-D. Choi / Renewable Energy xxx (2014) 1e8
Fig. 16. Averaged pressure distribution around the surface of runner blades of Stage 2 at 20 guide vane angle.
Stage 2 as well. From this result, the effectiveness of air supply from air suction Hole B on the pressure difference is considerable at both Stages 1 and 2. 3.5. Average local output torque on the blade Fig. 18 shows the local output torque distribution at the local circumferential locations at 20 guide vane angle. The local output torque is monitored for a single blade at the local circumferential locations of one period. It is clearly visible that the Region 1 appears a negative torque by Cases 1 and 2 where the torque losses occur remarkably. However, the Case 3 shows that there is almost rare torque loss at the Region 1, which means the effectiveness of output torque improvement by supplying air from air suction Hole B. Fig. 19 shows the averaged output torque on the blades at each stage by each case at 20 guide vane angle. All of the torque values are divided by the total torque of Case 1, which means the total torque of Case 1 is 1.0. The Stage 1 appears to have the highest torque in all cases. There is more than 70% of local output torques at the Stage 1 by the all test cases, which is almost similar to those by Mockmore et al. [2] and Choi et al. [8,12]. Furthermore, the maximum positive torque is achieved by Case 3 at Stages 1 and 2 with the least negative torque at Region 1. Hence, Case 3 shows the maximum overall total torque output.
Fig. 17. Effective pressure difference ratio at 20 guide vane angle.
7
Fig. 18. Local output torque distribution at the local circumferential locations at 20 guide vane angle.
The result of output power analysis implies that proper air layer formation in the runner passage by the air supply from air suction Hole B gives significant effect on the improvement of output power in the Stages 1 and 2. Moreover, the air layer from air suction Hole B suppresses the output torque loss at Region 1 by reducing the negative torque considerably. 4. Conclusions According to the investigation of the influence of air supply on the performance and the internal flow characteristic of a cross flow turbine, the following conclusions are obtained: 1. The improvement of output torque and efficiency are considerable with supplying air from air suction Hole B. The proper air layer formation in the runner passage by air supply from air suction Hole B gives significant effect on the improvement of output torque at Stages 1 and 2, and the effect to suppress the negative torque at Region 1 is obvious. The efficiency changed from 77.8% (Case 1) to 81.3% (Case 3) by air supply from air suction Hole B at 20 guide vane angle. 2. The efficiency increase is approximately proportional with increasing air layer area in the turbine chamber. The air layer area is the largest by air supply from air suction Hole B in comparison to other Cases. The air layer from air suction Hole B is larger at high guide vane angle than that at low guide vane angle.
Fig. 19. Averaged local output torque ratio for each case at 20 guide vane angle.
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
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Z. Chen, Y.-D. Choi / Renewable Energy xxx (2014) 1e8
3. The air from air suction Hole B gives significant effect on the tangential velocity difference of the cross flow turbine at Stage 2, but the effect is poor at Stage 1. Moreover, the effect on the pressure difference is strong, both at Stages 1 and 2. Therefore, the output torque increases at Stage 1 by increased pressure difference, but at Stage 2 the output torque increases by the increase of both the tangential velocity difference and pressure difference. 4. With air supply from air suction Hole B, suppression of output torque loss at Region 1 by reducing the negative torque is obvious. Therefore, the total torque is the highest one with air supply from air suction Hole B. References [1] Sorensen B. Renewable energy conversion, transmission and storage. Elsevier; 2007. [2] Mockmore CA, Merryfield F. The banki water turbine. Corvallis: Engineering Experiment Station, Oregon State System of Higher Education, Oregon State College; 1949. Bulletin Series No. 25.
[3] Fiuzat AA, Akerkar BP. Power outputs of two stages of cross-flow turbine. J Energy Eng 1991;117:57e70. [4] Khosrowpanah S, Fiuzat AA, Albertson ML. Experimental study of cross-flow turbine. J Hydraul Eng 1988;114:229e314. [5] Fukutomi J, Nakase Y, Watanabe T. A numerical method of free jet from a cross-flow turbine nozzle. Jpn Soc Mech Eng 1985;28:1436e40. [6] Kokubu K, Kanemoto T, Son SW, Choi YD. Performance improvement of a micro eco cross-flow hydro turbine. J Korean Soc Mar Eng 2012;36: 902e9. [7] Choi YD, Kim JI, Lee YH. Performance and internal flow characteristics of a cross-flow hydro turbine by the shapes of nozzle and runner blade. J Fluid Sci Technol 2008;3:398e409. [8] Choi YD, Shin BR, Lee YH. Air layer effect on the performance improvement of a cross-flow hydro turbine. J Korean Fluid Mach Assoc 2010;13:37e43. [9] ANSYS Inc. ANSYS CFX documentation. ver. 14; 2013., http://www.ansys.com. [10] Ariff M, Salim SM, Cheah SC. Wall Yþ approach for dealing with turbulent flow over a surface mounted cube: part 1-low reynolds number. In: Proc. Int. Conf. on CFD in the minerals and process industries; 2009. [11] Ariff M, Salim SM, Cheah SC. Wall Yþ approach for dealing with turbulent flow over a surface mounted cube: part 2-high reynolds number. In: Proc. Int. Conf. on CFD in the minerals and process industries; 2009. [12] Choi YD, Son WS. Shape effect of inlet nozzle and draft tube on the performance and internal flow of cross-flow hydro turbine. J Korean Soc Mar Eng 2012;36:351e7.
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
Contents lists available at ScienceDirect
Renewable Energy journal homepage: www.elsevier.com/locate/renene
Influence of air supply on the performance and internal flow characteristics of a cross flow turbine Zhenmu Chen a, Young-Do Choi b, * a
Graduate School, Department of Mechanical Engineering, Mokpo National University, Muan-gun 530-729, Jeollanam-do, Republic of Korea Department of Mechanical Engineering, Institute of New and Renewable Energy Technology Research, Mokpo National University, Muan-gun 530-729, Jeollanam-do, Republic of Korea
b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 12 March 2014 Accepted 8 August 2014 Available online xxx
This study attempts to improve the efficiency of a given type of cross flow turbine by supplying air from air suction holes. A newly developed air supply method is adopted. CFD analysis of the cross flow turbine is carried out to investigate the performance and internal flow characteristics of the turbine in detail. The air layer prevents shock loss between water flow and axis and suppresses recirculation flow in the runner passage. Hence, it is necessary to measure the amount of air layer in the runner passage and examine its effect on the performance of the cross flow turbine. The result shows that the turbine efficiency has improved more as the newly developed air supply method is applied effectively. © 2014 Elsevier Ltd. All rights reserved.
Keywords: Cross flow turbine Performance improvement Internal flow Air layer effect
1. Introduction Recently, the environmental problems such as global warming, pollution problems and soon, have been considered as inevitable issue. Therefore, the necessity of the use of renewable energy as one of the clean and sustainable natural energy resources has become high [1]. In addition, the cross flow turbine has a great potential for small hydropower development for both developed and undeveloped countries. Amongst renewable energy resources, the cross flow turbine is relatively reliable and cost effective with a very simple structure. The nozzle and the runner are made by processing steel plate and the blades are made by processing steel pipe giving it a simple design and manufacture. Therefore it can save more manufacturing cost and maintenance cost. Most present studies of cross flow turbines have optimized the configuration of turbines to improve the efficiency by experiment. The turbine design is based on Banki's one-dimensional analysis method [2]. Fiuzat et al. [3] has carried out an experiment to identify the contribution of each one of the cross flow turbine stages to the power output generation. Khosrowpanah et al. [4] and Fukutomi et al. [5] have tried to improve the performance of the turbine by changing the number of blades, the angles of water entry to the runner and the inner-to-outer diameter ratios. Kokubu et al.
* Corresponding author. E-mail addresses: [email protected] (Z. Chen), [email protected] (Y.-D. Choi).
[6] has improved the performance by changing the structure of turbine. On the other hand, Choi et al. [7,8] has investigated the effect of air layer located in the turbine runner passage. The air layer plays a significant role to improve the turbine performance. However, the internal flow in the turbine passage is complex, especially with the air layer. The flow with air layer has not been analyzed deeply enough. Therefore, this study attempts to improve the efficiency of the cross flow turbine by supplying air into the turbine chamber passage. Experiments are performed in order to measure the efficiency of the cross flow turbine by the conventional method to supply air. CFD analysis on the performance and internal flow of turbine is conducted in order to investigate the air layer effect on the turbine performance by newly developed method. 2. Test turbine and numerical methods 2.1. Experiment setup and cross flow turbine Fig. 1 shows the cross flow turbine for field test. The turbine is installed and operated for commercial production of electric power. The number of the runner blades is Z ¼ 30 and the diameter of the runner is d ¼ 340 mm. The inlet and outlet angles of the blade are b1 ¼ 34 and b2 ¼ 92.5 , respectively. The widths of the nozzle, runner and the draft tube are all the same, the width is b ¼ 500 mm. The guide vane angle ag ranges from 0 to 32 , which means the turbine starts from completely closed to open states, respectively. The design point of the present test turbine at the best efficient
http://dx.doi.org/10.1016/j.renene.2014.08.024 0960-1481/© 2014 Elsevier Ltd. All rights reserved.
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
2
Z. Chen, Y.-D. Choi / Renewable Energy xxx (2014) 1e8 Table 1 Cases of air suction hole methods by experiment and CFD analysis. Analysis cases
Air suction Hole A
Air suction Hole B
Experiment CFD
Ab 0 Ab Ab
0 0 0 Bb
a b
Case 1a Case 2a Case 3a
Guide vane angles ag: 15 , 18 , 20 , 22 , 25 . Air flow rate is determined for each guide vane angle (refer to Table 2).
every case. Table 2 shows the water flow rate and air flow rate ratio for each guide vane angle in detail. 2.2. Numerical methods
Fig. 1. Cross flow turbine for field test.
point is H ¼ 20 m for the effective head, n ¼ 530 min1 for rotational speed. Fig. 2 shows the efficiency curve of test turbine model by experiment [8]. The experiment is performed under the condition of fixing head (H) of the turbine, and only the air suction valve is open to supply a constant air flow rate. The result indicated that with increasing the guide vane angle, the efficiency increases as well at first, after it reaches the maximum efficiency, it decreases again slightly. Moreover, uncertainty in the measurement is indicated as error bars in the efficiency curve as well. In addition, Table 1 shows the cases of air suction methods by experiment and CFD analysis. CFD analysis for the Cases 1, 2 and 3 are conducted by variation of guide vane angle ag. For the Case 1, the working fluid is only water. There is no air supply into the turbine chamber. The condition of Case 2 is the same as the experiment which is only air suction Hole A is adopted (the conventional air supply method). Case 3 adopts two air suction holes of a conventional type by Hole A, as well as a type by Hole B which is new method of air supply as shown in Fig. 3. In Case 3, the air flow rate of Hole A is correspondingly the same with that in Case 2 at each guide vane angle. The water flow rates from inlet pipe are all the same in the experiment and CFD analysis. The air flow rate from the air suction hole is determined for each guide vane angle in
Fig. 2. Efficiency curve with uncertainty error bars of test turbine model by experiment.
A commercial code of ANSYS CFX is adopted for the numerical analysis on the turbine performance and internal flow [9]. Fig. 3 shows the schematic view of the turbine model for CFD analysis. As the flow in the turbine can be assumed to be uniform to the direction of main stream, which means no flow velocity component in the direction to the runner depth, the turbine model adopted for the CFD analysis is conducted by two-dimensional shape for considering the calculation time and computer capacity. The air suction Holes A and B are installed on the casing chamber upper wall and at the back of the draft tube, respectively. The area of runner passage can be divided into three regions by the flow patterns in the runner passage. Stage 1 obtains first output power and Stage 2 takes second output power. However, Region 1 consumes output power by hydraulic loss. To ensure relatively high accuracy of calculated results, fine hexahedral grids are generated for the turbine model. The grid element number of about 1.6 106 for the whole flow field has been used. Fig. 4 shows the fine hexahedral numerical mesh for whole view of test turbine model and the mesh for runner with high dense grids. Moreover, the value of yþ, which means non-dimensional distance from wall [10,11], is determined to be below 9 for the runner blade passage and below 22 for the other parts of the turbine passage as shown in Table 3. A suitable turbulence model is required for complex flow phenomena that occur on the internal flow. Therefore, the turbulence model dependence test is carried out by using three representative turbulence models: RNG k- ε, k-ε and SST turbulence models. From
Fig. 3. Schematic view of the numerical model.
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Z. Chen, Y.-D. Choi / Renewable Energy xxx (2014) 1e8 Table 2 The water flow rate and air flow rate ratio for each guide vane angle.
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Table 3 Non-dimensional wall distance.
Guide vane angle [+]
Water flow rate [m3/s]
Air flow rate ratio [Qair/Qwater]
Sections
yþ value
Hole A
Hole B
15 18 20 22 25
0.3625 0.4250 0.4650 0.5160 0.5624
0.0124 0.0105 0.0100 0.0145 0.0133
0.0289 0.0318 0.0323 0.0320 0.0320
Inlet Runner blade Center of runner Outlet
18 9 21 22
3. Results and discussion the validation test of the numerical model using turbulence models available from the present numerical code, SST model showed reasonable turbine performance and good convergence of calculation in comparison with the other turbulence models as shown in Fig. 5. Therefore, SST model is used for the single-phase flow and two-phase flow calculations. Unsteady numerical analysis is conducted in the CFD analysis for both single-phase and two-phase flow calculations with the effect of gravity. Total pressure at inlet and constant mass flow rate at outlet of the calculation domain are used for the boundary conditions. The boundary condition of velocity at air suction hole is used.
3.1. Performance curves Fig. 6 shows the performance curves of turbine model under the condition of various guide vane angle. The two efficiency curves by experiment and CFD analysis results of Case 2 are agree well with small deviation in the ranges of the partial and excessive load, which makes the methods and the computational results reliable. The inlet water flow increases along with the increase in guide vane angle accordingly. The efficiency of Case 3 shows the maximum value at every guide vane angles, while that of Case 1 shows the minimum value. These results imply that the air supply through air suction Hole B gives considerable effect on the improvement of the turbine efficiency. Moreover, the effect of air suction both from Holes A and B (Case 3) to improve efficiency is obvious at the medium load in comparison with the result of Cases 2, but the effectiveness becomes poor at the partial and excessive loads. 3.2. Air layer effect on the turbine performance To investigate the influence of air layer on the turbine performance, the two-dimensional air layer area in the runner blade passage and center is examined using the results of CFD analysis under the condition of water and air two-phase flow for Cases 2 and 3 as shown in Fig. 7. The figure shows the air volume fraction distributions of Case 3 (with air supply from both Holes A and B) at 20 guide vane angle. The red part (in web version) is the air layer, and the blue part is water. The air layer in the runner blade passage and center in Region 1 plays the role of preventing a shock loss in the runner axis and suppressing a recirculation flow in the runner [7,8]. Hence, introducing air layer in this region can improve the performance of the turbine. Furthermore, as the pressure in the center of the runner is relatively lower than that outside the turbine as shown in Fig. 8, the air is sucked into the center of runner from air
Fig. 4. Fine hexahedral numerical mesh for test turbine model (a)Whole view of numerical mesh for test turbine (b)Fine hexahedral numerical mesh for runner.
Fig. 5. Turbulence model dependence test for the cross flow turbine model (Case 2).
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024
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Fig. 6. Performance curves of turbine model.
suction Holes A and B without additional blower or compressor. Therefore, it needs to examine the relation between the turbine performance and the amount of air layer in the runner passage and its center. Fig. 9 shows the best efficiencies at different guide vane angles under the condition of various air layer area. The air layer area is measured in the runner blade passage and center. The efficiency increase is approximately proportional with increasing air layer area in the turbine chamber. Nevertheless, the efficiency growth rate by the guide vane angle of 25 is very small with increasing air layer area. The cases of 20 guide vane angle have the maximum efficiency compared with the others. With air suction from both Holes A and B (Case 3), there is more than 4% efficiency improvement at 15, 18 and 20 guide vane angles from the efficiencies by Case 1. Moreover there are 3.5% and 0.6% efficiency improvement at 22 and 25 guide vane angle by the air supply conditions from Cases 1 to 3, respectively. Fig. 10 presents the air layer area variation by guide vane angle in the runner blade passage and center for Cases 2 and 3. The air
layer area of Case 2 (air supply by Hole A only) gradually decreases along with increasing guide vane angle. Moreover, for Case 3, the air is sucked from both Holes A and B. The air flow rate from Hole A of Case 3 is the same with that of Case 2. Hence, the remaining air layer area is from Hole B which is marked by AB in Fig. 10. The AB is calculated by following equation:
AB ¼ AAB AA
where the AA and AAB represent the air layer area of Case 2 and Case 3, respectively. The AAB is much larger than AA and the overall trend of AB is upward along with the increase of the guide vane angle. Accordingly, the efficiency of Case 3 is higher than that of Case 2 (refer to Figs. 6 and 9) 3.3. Velocity distribution around the runner For the continuous flow condition, the relation between output torque (T) and the tangential velocity are calculated by following equation:
_ qr T ¼ mv
Fig. 7. Air volume fraction (Case 3 at 20 guide vane angle).
(1)
(2)
where r is radius of the runner, m_ is the mass of the water flow, and the vq is the tangential velocity. Therefore, it is very important to estimate the tangential velocity to predict the turbine output power, because output torque and output power of the turbine is derived from the tangential velocity. In order to investigate the internal flow in detail, averaged tangential velocity distribution around the runner at each Stages 1 and 2 is examined. Fig. 11 shows the schematic of runner regions for the investigation of velocity distribution on the runner circumference at Stages 1 and 2. Figs. 12 and 13 show that the averaged velocity at the runner circumferential passage from the suction side to pressure side of the blades at the inlet and outlet of Stages 1 and 2 by Cases 1 to 3 at 20 guide vane angle, respectively. The average data from the tangential velocity values at several runner passages are averaged to one runner passage with a single rotation of runner as shown in Fig 11. The symbols of q1ave: and q2ave: . in the abscissa denote the averaged passage location between the suction to pressure sides of runner blade at the Stages 1 and 2, respectively.
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Fig. 8. Pressure contours on the runner passage.
The ordinate is defined as velocity ratio (vq/u0), where the vq is the tangential flow velocity at the local radius and u0 is tangential runner velocity by the following equation:
u0 ¼ ru
(3)
difference ratio increases significantly, especially for Case 3. The tangential velocity difference ratio of Case 3 is the highest at Stage 2. The reason lies in the largest air layer area in the runner passage in Case 3. The results imply that the air from air suction Hole B gives considerable effect on the tangential velocity difference of the cross flow turbine at Stage 2, but the effectiveness is poor at Stage 1.
where u is angular velocity, r is the runner radius. From the results of Figs. 12 and 13, the tangential velocity ratio (vq/u0) at inlet is relatively higher in comparison with that at outlet. The difference of tangential velocity ratio at the inlet and outlet is directly proportional to the amount of torque of the water flow through the runner blade passage as shown in Eq. (2), which means it is proportional to the output power simultaneously. The vq/u0 at the outlet of Stage 1 shows lower than 1.0, but the vq/u0 at inlet of Stage 2 shows higher than 1.0, indicating that the water flow gets the kinetic energy again before the water flow enters into the Stage 2. It can be conjectured that due to the low pressure in the center of the runner, the water flow gains the kinetic energy again. Moreover, the air layer in the runner center plays a role of suppressing the collision loss between water flow and the runner axis. Fig. 14 shows the effective tangential velocity difference ratio at 20 guide vane angle. All the velocity difference area values are divided by that of Case 1 at Stage 1 to compare the effective tangential velocity difference ratio, which means the effective velocity difference area ratio of Case 1 at Stage 1 is 1.0. This figure presents that the tangential velocity difference ratio is almost similar each other among the Cases 1, 2 and 3 at Stage 1. However, at Stage 2, with air supply the tangential velocity
where pout is the static pressure at the outlet of the turbine and p is the local static pressure around the surface of runner blade location, r is the radius of runner and u is angular velocity. From Figs. 15 and 16, there are two parts of DCpn and DCpp in the closed area filled with the Cp curve. DCpn is the closed area filled with the Cp curve that the suction side pressure is larger than that of pressure side, which means the output torque generated by this part of pressure difference is negative. However, for the DCpp, it is contrary to that of DCpn, and the output torque made by this part of pressure difference is positive. Therefore, in order to have power to rotate the runner blade, the effective pressure difference (DCp ¼ DCppDCpn) should be positive.
Fig. 9. Best efficiencies by different air layer area.
Fig. 10. Air layer area variation by guide vane angle.
3.4. Pressure distribution around the runner blade The pressure distribution around the surface of runner blades of Stages 1 and 2 at 20 guide vane angle is shown in Figs. 15 and 16. The pressure coefficient (Cp) is calculated by following equation:
Cp ¼
p pout 0:5rðruÞ2
(4)
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Fig. 11. Selected runner passages for the investigation of velocity distribution.
Fig. 12. Averaged velocity of Stage 1 at 20 guide vane angle.
Fig. 14. Effective tangential velocity difference ratio at 20 guide vane angle.
Fig. 17 shows the effective pressure difference ratio at 20 guide vane angle. The ordinate is defined as effective pressure difference ratio. All the effective pressure differences (DCp) are divided by that of Case 1 at Stage 1 to compare the effective pressure difference ratio, which means that the effective pressure difference ratio of Case 1 at Stage 1 is 1.0. The equation of effective pressure difference ratio is calculated by following:
where the DCp1 is the effective pressure difference of Case 1 at Stage 1. The effective pressure difference ratio increases rapidly at Stages 1 and 2 with supplying air from air suction Hole B, which means that the runner blades get more output power by the pressure difference. Moreover, the pressure difference ratio of Stage 1 is larger than that of Stage 2 at each case, which means that the pressure transfer to output power at Stage 1 is more than that at
Cpe ¼
DCp DCp1
Fig. 13. Averaged velocity of Stage 2 at 20 guide vane angle.
(5)
Fig. 15. Averaged pressure distribution around the surface of runner blades of Stage 1 at 20 guide vane angle.
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Fig. 16. Averaged pressure distribution around the surface of runner blades of Stage 2 at 20 guide vane angle.
Stage 2 as well. From this result, the effectiveness of air supply from air suction Hole B on the pressure difference is considerable at both Stages 1 and 2. 3.5. Average local output torque on the blade Fig. 18 shows the local output torque distribution at the local circumferential locations at 20 guide vane angle. The local output torque is monitored for a single blade at the local circumferential locations of one period. It is clearly visible that the Region 1 appears a negative torque by Cases 1 and 2 where the torque losses occur remarkably. However, the Case 3 shows that there is almost rare torque loss at the Region 1, which means the effectiveness of output torque improvement by supplying air from air suction Hole B. Fig. 19 shows the averaged output torque on the blades at each stage by each case at 20 guide vane angle. All of the torque values are divided by the total torque of Case 1, which means the total torque of Case 1 is 1.0. The Stage 1 appears to have the highest torque in all cases. There is more than 70% of local output torques at the Stage 1 by the all test cases, which is almost similar to those by Mockmore et al. [2] and Choi et al. [8,12]. Furthermore, the maximum positive torque is achieved by Case 3 at Stages 1 and 2 with the least negative torque at Region 1. Hence, Case 3 shows the maximum overall total torque output.
Fig. 17. Effective pressure difference ratio at 20 guide vane angle.
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Fig. 18. Local output torque distribution at the local circumferential locations at 20 guide vane angle.
The result of output power analysis implies that proper air layer formation in the runner passage by the air supply from air suction Hole B gives significant effect on the improvement of output power in the Stages 1 and 2. Moreover, the air layer from air suction Hole B suppresses the output torque loss at Region 1 by reducing the negative torque considerably. 4. Conclusions According to the investigation of the influence of air supply on the performance and the internal flow characteristic of a cross flow turbine, the following conclusions are obtained: 1. The improvement of output torque and efficiency are considerable with supplying air from air suction Hole B. The proper air layer formation in the runner passage by air supply from air suction Hole B gives significant effect on the improvement of output torque at Stages 1 and 2, and the effect to suppress the negative torque at Region 1 is obvious. The efficiency changed from 77.8% (Case 1) to 81.3% (Case 3) by air supply from air suction Hole B at 20 guide vane angle. 2. The efficiency increase is approximately proportional with increasing air layer area in the turbine chamber. The air layer area is the largest by air supply from air suction Hole B in comparison to other Cases. The air layer from air suction Hole B is larger at high guide vane angle than that at low guide vane angle.
Fig. 19. Averaged local output torque ratio for each case at 20 guide vane angle.
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3. The air from air suction Hole B gives significant effect on the tangential velocity difference of the cross flow turbine at Stage 2, but the effect is poor at Stage 1. Moreover, the effect on the pressure difference is strong, both at Stages 1 and 2. Therefore, the output torque increases at Stage 1 by increased pressure difference, but at Stage 2 the output torque increases by the increase of both the tangential velocity difference and pressure difference. 4. With air supply from air suction Hole B, suppression of output torque loss at Region 1 by reducing the negative torque is obvious. Therefore, the total torque is the highest one with air supply from air suction Hole B. References [1] Sorensen B. Renewable energy conversion, transmission and storage. Elsevier; 2007. [2] Mockmore CA, Merryfield F. The banki water turbine. Corvallis: Engineering Experiment Station, Oregon State System of Higher Education, Oregon State College; 1949. Bulletin Series No. 25.
[3] Fiuzat AA, Akerkar BP. Power outputs of two stages of cross-flow turbine. J Energy Eng 1991;117:57e70. [4] Khosrowpanah S, Fiuzat AA, Albertson ML. Experimental study of cross-flow turbine. J Hydraul Eng 1988;114:229e314. [5] Fukutomi J, Nakase Y, Watanabe T. A numerical method of free jet from a cross-flow turbine nozzle. Jpn Soc Mech Eng 1985;28:1436e40. [6] Kokubu K, Kanemoto T, Son SW, Choi YD. Performance improvement of a micro eco cross-flow hydro turbine. J Korean Soc Mar Eng 2012;36: 902e9. [7] Choi YD, Kim JI, Lee YH. Performance and internal flow characteristics of a cross-flow hydro turbine by the shapes of nozzle and runner blade. J Fluid Sci Technol 2008;3:398e409. [8] Choi YD, Shin BR, Lee YH. Air layer effect on the performance improvement of a cross-flow hydro turbine. J Korean Fluid Mach Assoc 2010;13:37e43. [9] ANSYS Inc. ANSYS CFX documentation. ver. 14; 2013., http://www.ansys.com. [10] Ariff M, Salim SM, Cheah SC. Wall Yþ approach for dealing with turbulent flow over a surface mounted cube: part 1-low reynolds number. In: Proc. Int. Conf. on CFD in the minerals and process industries; 2009. [11] Ariff M, Salim SM, Cheah SC. Wall Yþ approach for dealing with turbulent flow over a surface mounted cube: part 2-high reynolds number. In: Proc. Int. Conf. on CFD in the minerals and process industries; 2009. [12] Choi YD, Son WS. Shape effect of inlet nozzle and draft tube on the performance and internal flow of cross-flow hydro turbine. J Korean Soc Mar Eng 2012;36:351e7.
Please cite this article in press as: Chen Z, Choi Y-D, Influence of air supply on the performance and internal flow characteristics of a cross flow turbine, Renewable Energy (2014), http://dx.doi.org/10.1016/j.renene.2014.08.024