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Lift-type and drag-type hydro turbine with vertical axis for power generation from water pipelines
Energy 188 (2019) 116070
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Lift-type and drag-type hydro turbine with vertical axis for power generation from water pipelines Wei Yang a, b, *, Yimin Hou c, Huiting Jia a, b, Benqing Liu a, b, Ruofu Xiao a, b a
College of Water Resources and Civil Engineering, China Agricultural University, Beijing, 100083, China Beijing Engineering Research Center of Safety and Energy Saving Technology for Water Supply Network System, Beijing, 100083, China c Research & Test Center, Dongfang Electric Machinery Co. Ltd, Deyang, 618000, Sichuan, China b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 14 December 2018 Received in revised form 15 June 2019 Accepted 4 September 2019 Available online 6 September 2019
Pipeline turbines are designed to supply power for smart sensors installed in water pipelines. They are installed in the water supply pipe network directly and should meet the performance requirements for starting up at low flow rates and having low pressure loss at high flow rates. To this end, we designed two types of vertical axis pipeline turbines with lift-type runner and drag-type runner respectively, and conducted both experimental test and numerical simulations on the hydraulic performance of both the turbines. The results show that both the lift-type and drag-type turbines can meet the performance requirements. Under the same incoming flow conditions, the critical startup flow rate, the tip speed ratio, and the power coefficient of the lift-type turbine is larger. And the startup process of the lift-type turbine is unstable. While the drag-type turbine shows the opposite performances in these respects. In the range of incoming flow velocity studied, the pressure loss of both the turbines increases with the incoming flow velocity, however, the maximum pressure loss is controlled in an acceptable range, which will not affect the normal water supply. © 2019 Elsevier Ltd. All rights reserved.
Keywords: Water pipeline monitor Vertical axis hydro turbine Lift-type runner Drag-type runner Startup performance Numerical simulation and experimental test
1. Introduction With the development of information technology, the automation degree of the water pipeline system is getting higher and higher. Therefore, many smart instruments such as sensors with remote data transmission are installed in the water pipeline system to monitor hydraulic performances and water quality conditions along the water pipelines [1e4]. One of the challenges of such tremendous monitor systems is its limited power resources to make the smart sensors with remote data transmission operate continuously and safely due to poor site conditions. In this paper, two kinds of pipeline hydro turbines with lift-type and drag-type runner respectively have been developed to generate power for the monitor system when extra water head in the pipeline can be consumed. The characteristics of the pipeline hydro turbine are different
* Corresponding author. College of Water Resources and Civil Engineering, China Agricultural University, Beijing, 100083, China. E-mail addresses: [email protected] (W. Yang), [email protected] (Y. Hou), [email protected] (H. Jia), [email protected] (B. Liu), [email protected] (R. Xiao). https://doi.org/10.1016/j.energy.2019.116070 0360-5442/© 2019 Elsevier Ltd. All rights reserved.
from the conventional turbines and energy recovery turbines [5]. The turbine needs to meet the sensor requirements of micro instantaneous power and large cumulative power, as well as micro power output (watt level), startup ability at low flow rate, and small pressure loss at high flow rate [6]. The sensors with remote data transmission are used for the data measurement, collection, and transmission of the water pipeline system, and provide data to support the real-time control of the operation status of the water supply system, especially the monitor of leakage problems [7]. Pipeline turbine needs to be installed directly in the water pipeline. In order to guarantee normal water supply, the following requirements should be met: (1) not changing the original water flow direction; (2) simple structure, small space occupation and easy installation; (3) not affecting the water quality; (4) good startup ability at low flow rate and low pressure loss at high flow rate. In order to meet the above requirements, a vertical axis type of pipeline turbine may be the best choice. And the structure of the existing pipeline turbine is derived from the vertical axis wind turbine. There are mainly two types. One is lift-type (Darrieus type) [8]and the other is drag-type (Savonius type) [9]. These two types of runners have been successfully developed and applied on different occasions.
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A lift-type hydro turbine developed by Lucid Energy Inc. is applied to pipeline with diameter more than 600 mm and a rated flow velocity more than 3 m/s. The purpose of its design and application is to extract the water energy as much as possible, which is different from the present pipeline turbine that consumes extra water head available in the water pipeline system. At the same time, the startup problem of the lift-type turbine with large diameter and large flow rate is easy to solve. However, for present pipeline turbine with lift-type runner the startup performance is a big problem. Since the flow rate is relatively low and the attack angle of the blade is dynamically changing during the working process of the lift-type pipeline turbine and the blade may be dynamically stalled [10]. The blade characteristic curve of the dynamic stall is ring-shaped, and the hysteresis phenomenon occurs [11]. The dynamic stall is a complex process accompanied by the separation of boundary layers, the generation, development and attenuation of stall vortices, and the reattachment of boundary layers [12], which have important influences on the startup performance of the turbine. A drag-type pipeline turbine with a diameter of 100 mm and a design flow rate of 1.5 m/s developed by Chen, J et al. [6] has been successfully applied by Hong Kong Water Supplies Department for supplying power to remote data transmission equipment. However, the pressure loss increases significantly with the increase of the flow velocity, and reaches 12 m when the flow velocity is 2.5 m/s. Therefore, the attention should be paid to the pressure loss problem at a large flow rate when the drag-type runner is applied in the water pipeline system. In summary, the existing lift-type pipeline turbines and the drag-type pipeline turbines are far from the performance requirements of the remote data transmission equipment installed in the water pipeline system. The existing lift-type pipeline turbines cannot meet the requirement of startup ability under low incoming flow velocity. At the same time the main goal of them is to generate electricity as large as possible which is unacceptable for water pipeline system since only the extra water head can be consumed. The pressure loss of the existing drag-type pipeline turbine increases significantly with the flow velocity, which cannot meet the requirement of low pressure loss under high flow rate. In this paper both the lift-type and drag-type turbines will be designed with improvement based on existing work to meet the specific requirements of the data remote transmission equipment listed before. The structure of the paper is as follows: in Section 2 hydraulic design of both the lift-type turbine and drag-type turbine are discussed. Then the hydraulic performances of the turbines are analyzed in Section 3 based on experimental test and numerical simulation. In Section 4 the main conclusions are drawn.
Pipe
Upper bottom
v
Shaft
Lower bottom
Blade
Fig. 1. The main structure of the lift-type turbine.
R
c
Fig. 2. The main design parameters of the lift-type turbine.
g[ ]ddBlade inclination N [-]ddNumber of blades
2. Hydraulic design of the pipeline hydro turbine 2.1. Lift-type turbine When the lift-type runner is applied in the water pipeline, the spherical shape of the runner is evolved to adapt to the circular section of the pipeline [13] as shown in Fig. 1. That is, the scanning trace of the blade is on a spherical surface and the radius of this spherical surface is the radius of the base circle of the runner. The main design parameters of the lift-type runner are shown in Fig. 2. The main design parameters of lift-type turbine are shown in Fig. 2. DN [m]ddThe diameter of the pipe R [m]ddThe radius of the base circle s[-]ddRotor solidity c [m]ddChord length of the blade airfoil
In this design, the pipe diameter DN is 200 mm. Considering the maximum utilization of the inner space of the pipe and more power generation, the radius R of the turbine is as large as possible, and sufficient clearance is retained in the pipe for consideration of high fluid permeability and low pressure loss. The concept of rotor solidity was introduced into the pipeline turbine here, which indicates the ratio of the sum of the blade chords on the equatorial plane (the section perpendicular to the turbine shaft and through the turbine centerline) of the spherical rotor to the circumference of the base circle. By consideration of the blade inclination g the rotor solidity of a lift-type turbine is defined as,
s¼
Nðc=cos gÞ R
(1)
where N is the number of blades, c is the chord length of the blade
W. Yang et al. / Energy 188 (2019) 116070
airfoil, and g is the inclination of the blade. Since the blade inclination angle is limited by the upper and lower support, the rotor solidity cannot be too large, and the influence of the number of blades and the chord length of the blade are important. The number of blades commonly used in the vertical axis wind turbines are 2 for “F“type and 3 for “H” type. The existing lift-type pipeline turbine developed by Lucid Energy Inc. has a runner with four blades [13]. Considering the smaller average flow rate of the pipeline turbines studied here, the number of blades were increased to six. And attention should be payed to the corresponding blade chord length to ensure the rotor solidity is within a reasonable range. Besides, the blade chord length should not be too small to ensure the strength of the blade. The blade inclination angle g as shown in Fig. 2 allows the blades to lie at different phase angles on different sections perpendicular to the axis, and mitigates the influence of the limited blade number. This parameter is similar to the blade helix angle that appears in the vertical axis wind turbine and the tidal current turbine. Since the trajectory of the blade is on the spherical surface rather than the cylindrical surface, the large inclination will make the effective height of the blade too small. So, the angle of the blade inclination angle should not be too large, and its influence to rotor solidity is smaller than the blade number and the chord length of the blade. Finally, the blade inclination angle was set to 15 here. And the standard airfoil NACA 0018 was used. As a result, the design parameters of the final lift-type turbine are shown in Table 1. 2.2. Drag-type turbine The drag-type pipeline turbine is evolved from the Savoniustype runner [14] which usually does not change the diameter of the runner in the axial direction. However, when the Savonius runner is applied in water pipeline, the diameter of the runner in the axial direction has to be variable to adapt to the shape of the pipe. The main structure of the present drag-type turbine is shown in Fig. 3. The main design parameters of the drag-type turbine are shown in Fig. 4. D [m]ddThe diameter of the runner N [-]ddNumber of blades s [m]ddThe diameter of inner circle s/DddOverlap H [m]ddThe height of runner W [ ]ddBlade chord angle It can be seen that for the Savonius wind turbine used in open environment, the increase of blade number does not bring an improvement in performance. When the drag-type runner is used in the water pipeline, it exhibits the opposite characteristics [6]. Whether it is a hollow (non-overlapping) or a non-hollow (overlapping) runner, the output power increases with the blade number in a certain range. And the blade number can exceed ten which is far more than that of the common drag-type wind turbines. This difference is closely related to that the drag-type runner used in the pipeline is constrained by the pipe wall, in which the fluid is not easy to bypass the runner. Instead, more fluid acts on the blade directly and the fluid permeability of the runner is decreased.
3
Referring to the existing study and taking into account the lower flow rate of the design, the blade number of the runner is set to twelve for the present drag-type turbine. The degree of overlap s/D is an important parameter affecting the performance of the drag-type wind turbine. When s/D is negative it is called non-hollow runner with overlapping blades. When s/D is positive it is called hollow runner with a void space between the runner blades as shown in Fig. 4. A certain degree of the blade overlapping can help to direct the fluid to the concave surface of the blade. However, as the number of blades increases, the flow capacity will decrease and the hydraulic loss will increase. Therefore, in the present case where the blade number is large, the hollow runner with s/D ¼ 0.5 was used to improve both the flow capacity and hydraulic loss of the turbine. When the drag-type runner is applied in the water pipeline, the pipe wall will hinder the discharge of the fluid at zone ④ as shown in Fig. 5. The high-pressure zone b is enlarged, and the adverse effect of the leeward blade is intensified. Therefore, some scholars have studied using a flow guiding device to direct the fluid to the windward blade to reduce the adverse effects of the leeward blade [6]. The purpose of the flow guiding device is to direct the flow to the windward blades and to increase the flow velocity by contracting the flow passage, which has similar effect as the inlet nozzle for drag-type hydro turbine [15]. An eye-type deflector baffle which shows a good performance [6] was used in the present case for the drag-type turbine. It should be noted that a stagnant water area can be easily generated between the baffle and the leeward blade, and large-scale vortices are easily generated in this area. The deflector baffle consists of a main baffle and a sub baffle as shown in Fig. 6(b). The baffle is used to direct the flow and increase its velocity, and the sub baffle wraps the runner to prevent the generation of large-scale vortices in the stagnant water area on the back side of the main baffle. For the deflector baffle the shrinkage rate ε is defined as the ratio of the flow path area blocked by the baffle to the cross-sectional area of the pipe. Large shrinkage rate is beneficial to guiding flow while small shrinkage rate is good for decreasing the local hydraulic loss. In this design, the deflector baffle shrinkage rate was optimized to be 0.58. Following the above analysis, the main design parameters of the drag-type turbine were determined as listed in Table 2. 3. Hydraulic performances of both the turbines 3.1. Test rig According to the test requirements, a pipeline turbine test bench was built to test the hydraulic performances of the turbine. The schematic diagram of the test bench circuit is shown in Fig. 7, which mainly includes a pipeline turbine test section, centrifugal pump, water tanks and electromagnetic flowmeters, brake, torque meter, pressure sensors, valves and so on. The brake is used to exert load for the tested turbine. Two pressure sensors were installed upstream and downstream of the turbine respectively. And the pressure test accuracy is 0.25%. The range of the torque meter is 0e2 N m with test accuracy of 0.5%. The brake is a magnetic remanence brake with rated torque of 2 N m and maximum rotational speed of 10000 r/min. The flowmeter is a three-electrode electromagnetic flowmeter with a measuring range of
Table 1 The main parameters of lift-type turbine. Number of blades
Chord length of airfoil blade
Blade inclination
Rotor solidity
Airfoil
6
26.60 mm
15
1.80
NACA 0018
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Shaft
Pipe
v
Blade Fig. 3. The main structure of the drag-type turbine.
s
Outer contour cutting line
D
Blade arc
(a) Front view
Fig. 4. The main design parameters of the drag-type turbine.
(b) Side view Fig. 6. Deflector baffle for drag-type pipeline turbine.
Fig. 5. Drag-type runner used in water pipeline.
14e282 m3/h, and its accuracy is 0.5%. The inlet and outlet valves of the test section are butterfly valves. The flow rate of the pipeline is controlled by adjusting the rotational speed of the pump. The turbine load is controlled by adjusting the current of the brake. The flow rate was directly read by the display on the flowmeter panel. The pressure, rotational speed and torque were collected by a selfdeveloped integrated data acquisition system. The test bench layout is shown in Fig. 8. The test was carried on two parts. One part is to measure the external characteristics of the turbine such as rotational speed, flow rate, pressure loss, torque and so on. Starting from no-load with the flow rate fixed, the current of the brake is gradually increased to
W. Yang et al. / Energy 188 (2019) 116070
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Table 2 The main parameters of the drag-type turbine. Outer diameter D
Inner diameter s
Height of runner H
Blade chord angle W
Shrinkage rate of the baffle ε
12
180 mm
90 mm
104 mm
15
0.58
Pressure Sensor_in
Pressure Sensor_out DN200*100
Turbine
Butterfly valve_out Torque meter
DN200
Butterfly valve_in
Water tank_out
Brake DN200*100 Reducer pipe
Electromagnetic Electromagnetic Flowmeter valve 1
Number of blades N
Electromagnetic Centrifugal pump valve 2 Water tank_in Fig. 7. The schematic diagram of the test bench circuit.
pump Brake Torque meter
Three pressure measuring points
H¼
pi po rg
(2)
P¼
2pMn 60
(3)
h¼
P
rgQH
(4)
where r is the fluid density, g is the gravitational acceleration, and they are treated as constant in the calculation. Here the statistical uncertainty was applied to evaluate the error of the experimental results. The statistical uncertainty u was estimated through the standard deviation from the arithmetic mean value of each measurement by
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u m X 2 u 1 uðxÞ ¼ t x x mðm 1Þ j¼1 j
(5)
where m is the time of the measurement, x is the measured variable, x is the arithmetic mean value of each measurement and it is P calculated by x ¼ m j¼1 xj =m. Since the pressure loss H, the power output P, and the efficiency h are not measured directly from the experiment, transfer formula is necessary to estimate the uncertainties of these parameters. According to equation (2), equation (3) and equation (4) the uncertainty transfer formula for pressure loss H, power output P and efficiency h can be expressed as
UðHÞ ¼
Turbine test section
100%
uðHÞ ¼ H
qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi uðpi Þ2 þ uðpo Þ2
0 1 uðPÞ U @P A ¼ ¼ P
(6)
H sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi uðMÞ 2 uðnÞ 2 þ n M
(7)
Fig. 8. The test bench layout.
0 1
enlarge the load on the turbine, and the external characteristics of the turbine are measured under several operating points. The rotational speed of the turbine, the torque, the inlet pressure, and the outlet pressure were recorded during this process. Second part is to test the turbine startup performance. The flow rate is adjusted gradually from low to high under no-load condition to observe the startup performance of the turbine, and the critical flow rate of the turbine startup was recorded. The critical flow rate of the turbine stop was also recorded by decreasing the flow rate under on-load condition.
uðhÞ U @hA ¼ ¼
h
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi ffi uðPÞ 2 uðHÞ 2 uðQ Þ 2 þ þ P H Q
where u(.) represents the uncertainties of the measured variables
3.2. Uncertainties of the experimental results In the experiment the inlet pressure of the turbine pi, the outlet pressure of the turbine po, the rotational speed n (rpm), the impeller torque M and the flow rate Q are measured. Based on these measured variables the pressure loss H of the turbine, the power output P of the turbine, and the efficiency h can be calculated by
(8)
Exit part Inlet part
Rotation zone
Fig. 9. The schematic diagram of the computation domain.
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W. Yang et al. / Energy 188 (2019) 116070
(a) Lift-type turbine
(b) Drag-type turbine
Fig. 10. Mesh distributions of the rotation zone.
calculated by equation (5), U(.) represents the relative uncertainties of the measured variables, and the overbar represents the arithmetic mean value of the measured variables. By using expression (6), expression (7), and expression (8) the experimental results of the pressure loss H, power output P and efficiency h and their relative uncertainties for lift-type turbine and drag-type turbine are shown in Fig. 13 and Fig. 17 respectively. And the maximum of the relative uncertainties is no more than 5%.
Fig. 11. Mesh of the inlet section (drag-type).
Fig. 12. Mesh of the exit section.
3.3. CFD settings 3.3.1. Computation domain and grid The whole Computation domain of both the lift-type and dragtype turbines is consistent, which is a cylindrical section of water in the pipeline as shown in Fig. 9. The upstream distance which is three times of the pipe diameter from the axis was set as the inlet part of the computation domain. Considering the full development of the flow, the downstream distance which is ten times of the pipe diameter from axis was set as the exit part of the computation domain. The computation domain was divided into three parts, a spherical rotation area with a rotor, an inlet pipe section (including an upstream pipe section and a portion around the rotation zone), and an outlet pipe section. A sphere that wraps the rotor was set as the rotation zone as shown in Fig. 9. Unstructured mesh was used in the inlet section and the rotation zone, while structural mesh was used in the outlet section. In the rotation zone where the runner is located, a finer mesh is required near the surface of the blades as shown in Fig. 10. The mesh for lift-type turbine should be fine enough at both the blade leading edge and trailing edge due to the large difference between the blade chord length and the overall size of the runner. The rotation zone mesh of both the lift-type and drag-type turbine is shown in Fig. 10. Since the rotation zone is located in the inlet section, there is a small gap between the interface and the pipe wall. The mesh size of the gap was generated to ensure that there were at least five layers in the gap. The drag-type turbine has a deflector structure and the mesh near the junction of the main baffle and the sub baffle were refined. Since the downstream flow of the runner is complicated the grid with a smaller scale was applied there. The mesh of the inlet section and exit section are shown in Figs. 11 and 12 respectively. The total number of the mesh for the lift-type and the dragtype turbine are 2.17 million and 2.84 million respectively.
3.3.2. Turbulence model and boundary conditions The ANSYS CFX was used for unsteady calculation. Since the flow inside the pipeline turbine has strong rotation effects, the RNG k ε model was selected for its advantages in considering the rotation and swirl effects in the averaged flow. A high-precision difference scheme was applied for the convection term, and a second order difference scheme was used for the diffusion term. The maximum iterations in each time step were set to 20, and the residual of 104 was set as the convergence criterion for all the simulated variables. The time step was determined according to the Courant number in the calculation process, and the average
W. Yang et al. / Energy 188 (2019) 116070
(a) V=0.79m/s (Re=1.57×105)
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(b) V=1.00m/s (Re=1.99×105)
(c) V=1.29m/s (Re=2.56×105)
Angle of attack [-]
Fig. 13. Power, efficiency and pressure loss curves of the lift-type turbine under different flow rates (P is the power, H is the pressure loss, and h is the efficiency).
Fig. 14. Velocity triangle analysis of the axial section of the runner (v is the incoming flow velocity, u is the angular velocity at which the rotor rotates, u is the peripheral velocity, and w is the relative velocity, a is the attack angle, q is the rotating angle of the runner after time t).
25 20 15 10 5 0 -5 -10 -15 -20 -25
n=200 n=250 n=300 n=348.2 n=439.9 n=480
=0°
0
60
120
180 240 Degree
300
360
Fig. 15. Angle of attack at different blade phase angle based on Eq. (3).
Courant number was kept less than unity. The average value of the y plus near the blade surface was from 30 to 50 and a standard wall function was used. It should be noted that during the transient calculation process, the rotational speed of the runner was fixed for simplicity, and the hydraulic performance of the turbine at different rotational speeds was calculated.
A velocity inlet condition was set at the inlet of the domain with the velocity direction normal to the inlet surface. A static pressure was set at the outlet of the domain. A non-slip condition was set at all the walls in the domain. And a moving mesh scheme was used for the unsteady flow simulation in the present domain including both stationary and rotation parts.
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W. Yang et al. / Energy 188 (2019) 116070
0.4 n=200 n=250 n=300 n=348.2 n=439.9 n=480
0.3
M [Nm]
0.2 0.1 0.0 =0°
-0.1 -0.2 -0.3
0
60
120
180 240 Degree
300
360
Fig. 16. Numerical results of single blade torque at different blade phase angel.
3.4. Experimental and numerical results 3.4.1. Lift-type turbine (1) External performances The power, efficiency and pressure loss curves of both the lift-
(a) v=0.79m/s (Re=1.57×105)
type and drag-type turbine under three different flow rates (Reynolds numbers) are shown in Figs. 13 and 17 respectively. These three flow rates were selected based on the work condition of the turbine. The numerical simulation was performed at the Reynolds number of 1.57 105 with corresponding flow velocity of 0.79 m/s for comparison. Obvious difference between the experimental and numerical results can be seen from Figs. 13(a) and 17(a). Three causes may contribute to the present simulation error. First one is the error from geometry simplification. For lift-type turbine the geometry simplification was performed at the pallet installed on the axis near the pipeline wall to avoid small gap between the pallet and the wall. For drag-type turbine both the axis and the pallet structure were simplified for convenience of grid generation. Second one is the error from uniform boundary conditions setup. Uniform boundary condition was set at both the inlet and outlet of the turbine which might not be the case for the experiment since flow disturbance may happen during the experiment. Third one is the error from the RNG k-epsilon turbulence model. This model is based on the eddy viscosity assumption which has limitations in the simulation of turbulence flow with rotation and curvature. However, the main characteristic of the simulated curves is consistent with the experimental results and qualitative analysis can be conducted based on the simulation results. The rotational speed of lift-type runner is in the range of 200 r/ ~ 700 r/min. The maximum efficiency of the turbine at each flow min rate is higher than 30%. At two large flow rates of V ¼ 1.00 m/s and V ¼ 1.29 m/s, the highest efficiency is more than 40%. The leftmost point of each test curve in Fig. 13 is the brake point of the turbine, and the corresponding rotational speed is the brake speed, under
(b) v=1.00m/s (Re=1.99×105)
(c) v=1.29m/s (Re=2.56×105) Fig. 17. Power, pressure loss and efficiency curve of the drag-type turbine (P is the power, H is the pressure loss, and h is the efficiency).
W. Yang et al. / Energy 188 (2019) 116070
which the turbine stops. The characteristics of the lift-type turbine that is easily braked under the condition of large load are consistent with its poor startup performance. Both the brake and startup performance are caused by the decrease of working ability when the blades work at a high angle of attack under low tip speed ratio. Both the efficiency and power of the turbine under each Reynolds number decrease with the increase of the rotational speed, while the pressure loss increases with the increase of the rotational speed. In order to further analyze the power characteristics of the lifttype turbine, it was assumed that the flow between the blades did not interfere with each other, regardless of the flow influence between the axial sections of the blade. The velocity triangle of the axial section of the runner is shown in Fig. 14, where v is the incoming flow velocity, u is the angular velocity at which the rotor rotates, u is the peripheral velocity, and w is the relative velocity. In a rotating reference frame, the absolute velocity is equal to the vector sum of the relative velocity and the convected velocity. The angle of attack a is the angle between the relative speed w and the circumferential direction. Although the direction of the flow velocity v in the pipe does not change, the angle between the peripheral velocity u and v changes constantly as the rotor rotates. In the velocity triangle, the sine theorem can be used as,
v u ¼ sin a sin½90 a ð90 qÞ
(2)
where q is the rotating angle of the runner after time t and it is defined by q ¼ ut. The formula of the attack angle a can be derived from Eq. (2),
a ¼ arctan
sin q cos q þ l
(3)
where l ¼ u=v is the tip speed ratio. It can be seen from Eq. (3) that the angle of attack is only related to the position of the blade and the tip speed ratio of the runner. The variation range of the attack angle decreases with the tip speed ratio of the runner and the stall is not likely to occur. When the tip speed ratio of the runner increases, the risk of the airfoil working at large angles of attack will be increased accordingly. The numerical results about variation of the single blade torque with the blade phase angle under the Reynolds number of 1.57 105 are shown in Fig. 16, and the corresponding theoretical angle of attack given by Eq. (3) are shown in Fig. 15. The airfoil used in the blade is a symmetrical airfoil. The change of the theoretical angle of attack is symmetrical, while the variation of the torque got by simulation is asymmetry in a rotation period of the runner. Its positive peak is located in the middle of the upstream, which corresponds to the phase angle of the maximum positive angle of attack. However, when the blade reaches the theoretical maximum negative angle of attack, there is no torque output, which may be related to the complicated flow field when the blade is located in downstream. At this time, the blade passes through the low-speed area in the middle of the runner, and the actual angle of attack of the blade is far below the theoretical angle of attack. When the blade is located between 50 150 , the output torque is large, and the theoretical angle of attack of the blade exceeds 5 . When the rotational speed exceeds 300 r/min, the phase angles of the torque peak value under different rotational speed are close. When the rotational speed decreases from 300 r/min, the maximum angle of attack is too large, and the torque peak shifts to the low phase angle, as a result the maximum torque is reduced. When the blade is located between 180 360 , the negative torque generated by the blade in the downstream region increases with the rotational
9
speed. At this time, the blade passes through the low-speed region of the runner, and the low-speed area gradually expands with the increase of the rotational speed and the flow is more chaotic, which causes generation of the negative torque. And these results are consistent with the turbine power characteristics shown in Fig. 13 obtained in experimental tests. (2) Startup characteristics The startup of the lift-type pipeline turbine is an unstable process with the vibration of the runner and flow uncertainty. Therefore, it is difficult to measure the critical startup flow rate steadily and accurately. However, it can still provide a valuable reference for the startup characteristics study of the lift-type turbine. The minimum startup flow rate v0 of the turbine is close to 0.69 m/s which is a relatively large value for the turbine application conditions in water pipeline system. The critical startup process was observed at the flow rate of v0 ¼ 0:69m=s, which can be divided into five stages: i. The runner is shaking dramatically with intermittent slow rotation. ii. Intermittent slow rotation makes the runner reach a phase angle with a large torque, then the runner is accelerated. iii. The accelerating runner does not stop before it reaches the next phase angle with a small torque. iv. The runner changes from intermittent rotation to continuous slow rotation, retaining continuous low speed rotation and slow acceleration. v. The runner passes the startup process and accelerates quickly to reach a steady operation condition. The startup condition of the turbine is that the startup torque is greater than the drag torque. Due to the dynamic change of the blade angle of attack during the rotation of the lift-type runner, the runner torque is also dynamically changed. Therefore, the startup torque will change with the different phase position of the runner, and it may be greater or less than the drag torque, which shows the above stage i and stage ii. When the runner is in the intermittent slow rotation, the rotation will make the runner rotate to a phase angle with a larger startup torque. Whether the runner can successfully reach the stage iii through the phase angle with a small startup torque is the first key for the successful startup of the turbine. After the runner passes the intermittent slow rotation, it will enter the rapid acceleration stage and before reaching the steady operation condition it is in the stage iv. At this stage, the tip speed ratio of the runner is low, the airfoil may have a stall and the generated torque is small. Whether the runner can pass the possible stall state to the stage v is another key issue in the startup of the lift-type turbine. The critical stop flow rate for lift-type turbine was also recorded by adjusting the flow rate from high level to low level, which is about 34% of the critical startup flow rate. 3.4.2. Drag-type turbine (1) External performances The power, pressure loss and efficiency test results of the dragtype turbine under three Reynolds numbers are shown in Fig. 17. The numerical simulation was performed at the Reynolds number of 1.57 105 for comparison. In each operating condition, the rotational speed is significantly less than that of the lift-type turbine. From the form of the power and efficiency curves, there is a peak value of the output power and no brake point at any rotational speed. The efficiency of the drag-type turbine is lower which does not exceed 20% under all operating conditions. The results obtained
10
W. Yang et al. / Energy 188 (2019) 116070
by the simulation of the drag-type turbine are quite different from the experimental values, as shown in Fig. 17(a), which was also reported in Ref. [6]. The pressure loss curve of the drag-type turbine changes gently, and the variation trend of the power curve and the efficiency curve is similar, which is shown in both the simulation and test results. When the rotational speed is low, the output power is low because the fluid bypasses the rotor or passes through the center of the runner directly, without pushing the blade to work. When the rotational speed is large, the flow field structure is disordered, and the negative torque increases. At the same time the efficiency of the turbine and the output power are reduced. The zero-phase angular position of the drag-type turbine blade is shown in Fig. 18. The single-blade torque of the runner rotating at different speeds is shown in Fig. 19. When the rotational speed is more than 100 r/min, the blade generates the maximum torque at the phase angle of 100 200 as shown in Fig. 19. This position is the outlet of the deflector baffle at which the outflow is working directly on the blades, and the maximum torque decreases with the increase of the rotational speed. The negative moment increases when the blade rotates to the junction of the main baffle and the sub baffle. When the rotational speed is less than 100 r/min, the torque peak position appears after 280 as shown in Fig. 19, that's because the fluid fails to do work when it first passes the runner, while the fluid does work when it leaves the runner. It can be seen that although the fluctuation of the single-blade torque is large, the
generation of the negative torque of the drag-type turbine with the deflector baffle is suppressed, and there is no large-scale negative moment region compared with the lift-type turbine. (2) Startup characteristics The startup performance of the drag-type turbine was tested with a slow increase in the flow rate. The critical startup flow rate of the drag-type turbine is 0.18 m/s which is significantly lower than that of the lift-type turbine. The drag-type turbine is stable whether it is starting or stopping, and there is no strong shaking of the runner. The drag-type turbine blade has a large windward projection area, and is subjected to a large force of water flow in a stationary state, which is advantageous for the startup performance. The critical stop flow rate was also recorded, which is about 77% of the critical startup flow rate for the drag-type turbine. 3.4.3. Comparisons In general, the rotational speed of the drag-type turbine is significantly lower than that of the lift-type turbine at the same flow rate, and the lift-type turbine suddenly brakes when the load is increased to a certain value. In order to compare the characteristics of both the lift-type and drag-type turbine with different flow rates and different rotational speeds, the rotational speed of the turbine is nondimensionalized as
l¼
2pRn v
(4)
where l is the tip speed ratio, R is the runner diameter, n is the rotational speed, and v is the incoming flow velocity. The power factor is defined as
CP ¼ 1
P
(5)
3 2 rAv
where A is the cross-sectional area of the pipe, P is the output power of the runner, v is the velocity of the incoming flow, and r is the density of the water. Fig. 20 shows the variation of the power coefficient of both the lift-type and drag-type turbine with the tip speed ratio. Under the different flow rates, the power coefficient curves of the drag-type turbine tend to coincide, and they are Fig. 18. Zero-phase angular position of the drag-type turbine blade.
v=0.34 v=0.50 v=0.79 v=1.00 v=1.29 Solid: lift-type hollow: drag-type
3.5 3.0
0.8
2.5 CP [-]
M [Nm]
0.4 0.0
1.5 1.0
n=40 n=70 n=100
-0.4 -0.8
2.0
0
60
120
180 240 Degree
0.5
n=130 n=160 n=180
300
Fig. 19. Single-blade torque of the runner rotating at different speeds.
0.0 360
-0.5
0
1
2
4
[-]
5
6
7
8
Fig. 20. Power coefficient curve of both the lift-type and drag-type pipeline turbine.
W. Yang et al. / Energy 188 (2019) 116070
v=0.34 v=0.50 v=0.79 v=1.00 v=1.29 Solid: lift-type hollow: drag-type
0.8
H [m]
0.6
0.4
0.2
0.0
0
1
2
4
[-]
5
6
7
8
Fig. 21. Pressure loss curve of both the lift-type and drag-type pipeline turbine.
consistent for the lift-type turbine. The characteristics of the lifttype vertical axis pipeline turbine and the drag-type vertical axis pipeline turbine were compared with the corresponding type of wind turbine [16]. The tip speed ratio of the drag-type pipeline turbine is about unity, which is close to the tip speed ratio of the Savonius type of wind turbine. The tip speed ratio of the lift-type pipeline turbine is in the range of four to seven, which is also close to the Darrieus type of wind turbine. In addition, the multiple relationship between the power coefficient of the lift-type pipeline turbine and the drag-type pipeline turbine is similar to that of the Darrieus-type wind turbine and the Savonius-type wind turbine, which is about a triple relationship between the two turbines. Fig. 21 shows the pressure loss curves of both the lift-type and drag-type turbine with the tip speed ratio. For both turbines, the pressure loss increases significantly with the flow rate, and for the same flow rate of the same runner, the pressure loss increases with the rotational speed. However, the pressure loss of both turbines is less than 3% of the total head in the range of the incoming flow velocity studied, which has little influence on the normal water supply.
4. Conclusions Vertical axis turbines with both the lift-type and drag-type runner for micro power generation in water pipeline system are studied experimentally and numerically. The vertical axis pipeline turbine is directly installed in the water pipeline system to supply power for the data remote transmission equipment. It should meet the requirements of startup at low flow rate and low-pressure loss at high flow rate without affecting the normal water supply. In this paper, both the lift-type pipeline turbine and the dragtype pipeline turbine are designed with reference to the Darrieus wind turbine and the Savonius wind turbine, and the external and startup performances of both the turbines are experimentally measured. It is found that there are five stages during the startup process of the lift-type turbine. The five-stage startup process is unstable and is accompanied by intermittent shaking of the runner, which is caused by the dynamical change of the difference between the startup torque and the drag torque due to the dynamical variation of the attack angle. As a result, the critical incoming flow velocity of the startup process for lift-type turbine is larger than
11
that of the drag-type turbine, which can start at a lower flow rate. And the starting process of the drag-type turbine is stable. Under different incoming flow rates, the power coefficient curves with the change of the tip speed ratio of the lift-type turbine are consistent, while they tend to coincide for the drag-type turbine. The lift-type turbine has a higher speed ratio and higher output power than that of the drag-type turbine. As the turbine load increases, the lift-type turbine will brake at a large incoming flow rate until it stops running, while the drag-type turbine has a significantly lower braking flow rate. It should be noticed that the critical startup flow rate is not consistent with the critical stop flow rate for both the turbines, which indicates that there is hysteresis effect in the flow system especially for the lift-type turbine. In the range of the incoming flow velocity studied, the pressure loss of both the lift-type turbine and the drag-type turbine increases with the incoming flow rate. However, the maximum pressure loss does not exceed 3% of the total water head, which will not affect the normal water supply. In general, both the lift-type turbine and the drag-type turbine can meet the energy requirements of the data remote transmission equipment in water pipeline system, however, there is still room for improvement. For the lift-type turbine, the critical incoming flow rate of the startup can be further reduced, and the stability of the starting process can be improved. For the drag-type turbine, the range of the tip speed ratio can be further expanded, and the output power can be increased. Acknowledgements The author would like to acknowledge the financial support received from National Key Research and Development Plan (grant number 2018YFB0606103) and Beijing Waterworks Group. References [1] Yang X, Ong KG, Dreschel WR, Zeng K, Mungle CS, Grimes CA. Design of a wireless sensor network for long-term, in-situ monitoring of an aqueous environment. Sensors 2002;2(11):455e72. [2] Stoianov IA, Nachman LB, Whittle AC, Madden SD, Kling RB. Sensor networks for monitoring water supply and sewer systems: lessons from Boston. Cincinnati, OH: 8th Annual Water Distribution Systems Analysis Symposium; 2006. [3] Qureshi FU, Muhtaroglu A, Tuncay K. Near-optimal design of scalable energy harvester for underwater pipeline monitoring applications with consideration of impact to pipeline performance. IEEE Sens J 2017;17(7):1981e91. [4] Qureshi FU, Muhtaroglu A, Tuncay K. A method to integrate energy harvesters into wireless sensor nodes for embedded in-pipe monitoring applications. Cairo, EGYPT: 5th International Conference on Energy Aware Computing Systems & Applications (ICEAC); 2015. [5] Du J, Yang H, Shen Z, Chen J. Micro hydro power generation from water supply system in high rise buildings using pump as turbines. Energy 2017;137: 431e40. [6] Chen J, Yang HX, Liu CP, Lau CH, Lo M. A novel vertical axis water turbine for power generation from water pipelines. Energy 2013;54(2):184e93. [7] Mazzolani G, Berardi L, Laucelli D, Simone A, Martino R, Giustolisi O. Estimating leakages in water distribution networks based only on inlet flow data. J Water Resour Plan Manag 2017;143(040170146). [8] Islam M, Ting D, Fartaj A. Aerodynamic models for Darrieus-type straightbladed vertical axis wind turbines. Renew Sustain Energy Rev 2008;12(4): 1087e109. [9] Golecha K, Eldho TI, Prabhu SV. Influence of the deflector plate on the performance of modified Savonius water turbine. Appl Energy 2011;88(9): 3207e17. [10] Fujisawa N, Shibuya S. Observations of dynamic stall on Darrieus wind turbine blades. J Wind Eng Ind Aerodyn 2001;89(2):201e14. [11] Mcalister KW, Carr LW, Mccroskey WJ. Dynamic stall experiments on the NACA 0012 airfoil, NASA TP-1100. NASA; 1978. [12] Carr LW, Mcalister KW, Mccroskey WJ. Analysis of the development of dynamic stall based on oscillating airfoil experiments, NACA TN D-8382. NASA; 1977. [13] Schlabach RA, Cosby MR, Kurth E, Palley I, Smith G. In-pipe hydro-electric power system and turbine. 2011. p. 6e14. United States Patent, US 7959411B2. [14] Akwa JV, Vielmo HA, Petry AP. A review on the performance of Savonius wind
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W. Yang et al. / Energy 188 (2019) 116070
turbines. Renew Sustain Energy Rev 2012;16(5):3054e64. [15] Kai S, Furukawa A, Okuma K, et al. Experimental study on simplification of Darrieus-type hydro turbine with inlet nozzle for extra-low head hydropower
utilization. Renew Energy 2012;41(2):376e82. [16] Wilson RE, Lissaman PBS. Applied aerodynamics of wind power machines. Nasa Sti/recon Technical Report N; 1974.
Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
Lift-type and drag-type hydro turbine with vertical axis for power generation from water pipelines Wei Yang a, b, *, Yimin Hou c, Huiting Jia a, b, Benqing Liu a, b, Ruofu Xiao a, b a
College of Water Resources and Civil Engineering, China Agricultural University, Beijing, 100083, China Beijing Engineering Research Center of Safety and Energy Saving Technology for Water Supply Network System, Beijing, 100083, China c Research & Test Center, Dongfang Electric Machinery Co. Ltd, Deyang, 618000, Sichuan, China b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 14 December 2018 Received in revised form 15 June 2019 Accepted 4 September 2019 Available online 6 September 2019
Pipeline turbines are designed to supply power for smart sensors installed in water pipelines. They are installed in the water supply pipe network directly and should meet the performance requirements for starting up at low flow rates and having low pressure loss at high flow rates. To this end, we designed two types of vertical axis pipeline turbines with lift-type runner and drag-type runner respectively, and conducted both experimental test and numerical simulations on the hydraulic performance of both the turbines. The results show that both the lift-type and drag-type turbines can meet the performance requirements. Under the same incoming flow conditions, the critical startup flow rate, the tip speed ratio, and the power coefficient of the lift-type turbine is larger. And the startup process of the lift-type turbine is unstable. While the drag-type turbine shows the opposite performances in these respects. In the range of incoming flow velocity studied, the pressure loss of both the turbines increases with the incoming flow velocity, however, the maximum pressure loss is controlled in an acceptable range, which will not affect the normal water supply. © 2019 Elsevier Ltd. All rights reserved.
Keywords: Water pipeline monitor Vertical axis hydro turbine Lift-type runner Drag-type runner Startup performance Numerical simulation and experimental test
1. Introduction With the development of information technology, the automation degree of the water pipeline system is getting higher and higher. Therefore, many smart instruments such as sensors with remote data transmission are installed in the water pipeline system to monitor hydraulic performances and water quality conditions along the water pipelines [1e4]. One of the challenges of such tremendous monitor systems is its limited power resources to make the smart sensors with remote data transmission operate continuously and safely due to poor site conditions. In this paper, two kinds of pipeline hydro turbines with lift-type and drag-type runner respectively have been developed to generate power for the monitor system when extra water head in the pipeline can be consumed. The characteristics of the pipeline hydro turbine are different
* Corresponding author. College of Water Resources and Civil Engineering, China Agricultural University, Beijing, 100083, China. E-mail addresses: [email protected] (W. Yang), [email protected] (Y. Hou), [email protected] (H. Jia), [email protected] (B. Liu), [email protected] (R. Xiao). https://doi.org/10.1016/j.energy.2019.116070 0360-5442/© 2019 Elsevier Ltd. All rights reserved.
from the conventional turbines and energy recovery turbines [5]. The turbine needs to meet the sensor requirements of micro instantaneous power and large cumulative power, as well as micro power output (watt level), startup ability at low flow rate, and small pressure loss at high flow rate [6]. The sensors with remote data transmission are used for the data measurement, collection, and transmission of the water pipeline system, and provide data to support the real-time control of the operation status of the water supply system, especially the monitor of leakage problems [7]. Pipeline turbine needs to be installed directly in the water pipeline. In order to guarantee normal water supply, the following requirements should be met: (1) not changing the original water flow direction; (2) simple structure, small space occupation and easy installation; (3) not affecting the water quality; (4) good startup ability at low flow rate and low pressure loss at high flow rate. In order to meet the above requirements, a vertical axis type of pipeline turbine may be the best choice. And the structure of the existing pipeline turbine is derived from the vertical axis wind turbine. There are mainly two types. One is lift-type (Darrieus type) [8]and the other is drag-type (Savonius type) [9]. These two types of runners have been successfully developed and applied on different occasions.
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W. Yang et al. / Energy 188 (2019) 116070
A lift-type hydro turbine developed by Lucid Energy Inc. is applied to pipeline with diameter more than 600 mm and a rated flow velocity more than 3 m/s. The purpose of its design and application is to extract the water energy as much as possible, which is different from the present pipeline turbine that consumes extra water head available in the water pipeline system. At the same time, the startup problem of the lift-type turbine with large diameter and large flow rate is easy to solve. However, for present pipeline turbine with lift-type runner the startup performance is a big problem. Since the flow rate is relatively low and the attack angle of the blade is dynamically changing during the working process of the lift-type pipeline turbine and the blade may be dynamically stalled [10]. The blade characteristic curve of the dynamic stall is ring-shaped, and the hysteresis phenomenon occurs [11]. The dynamic stall is a complex process accompanied by the separation of boundary layers, the generation, development and attenuation of stall vortices, and the reattachment of boundary layers [12], which have important influences on the startup performance of the turbine. A drag-type pipeline turbine with a diameter of 100 mm and a design flow rate of 1.5 m/s developed by Chen, J et al. [6] has been successfully applied by Hong Kong Water Supplies Department for supplying power to remote data transmission equipment. However, the pressure loss increases significantly with the increase of the flow velocity, and reaches 12 m when the flow velocity is 2.5 m/s. Therefore, the attention should be paid to the pressure loss problem at a large flow rate when the drag-type runner is applied in the water pipeline system. In summary, the existing lift-type pipeline turbines and the drag-type pipeline turbines are far from the performance requirements of the remote data transmission equipment installed in the water pipeline system. The existing lift-type pipeline turbines cannot meet the requirement of startup ability under low incoming flow velocity. At the same time the main goal of them is to generate electricity as large as possible which is unacceptable for water pipeline system since only the extra water head can be consumed. The pressure loss of the existing drag-type pipeline turbine increases significantly with the flow velocity, which cannot meet the requirement of low pressure loss under high flow rate. In this paper both the lift-type and drag-type turbines will be designed with improvement based on existing work to meet the specific requirements of the data remote transmission equipment listed before. The structure of the paper is as follows: in Section 2 hydraulic design of both the lift-type turbine and drag-type turbine are discussed. Then the hydraulic performances of the turbines are analyzed in Section 3 based on experimental test and numerical simulation. In Section 4 the main conclusions are drawn.
Pipe
Upper bottom
v
Shaft
Lower bottom
Blade
Fig. 1. The main structure of the lift-type turbine.
R
c
Fig. 2. The main design parameters of the lift-type turbine.
g[ ]ddBlade inclination N [-]ddNumber of blades
2. Hydraulic design of the pipeline hydro turbine 2.1. Lift-type turbine When the lift-type runner is applied in the water pipeline, the spherical shape of the runner is evolved to adapt to the circular section of the pipeline [13] as shown in Fig. 1. That is, the scanning trace of the blade is on a spherical surface and the radius of this spherical surface is the radius of the base circle of the runner. The main design parameters of the lift-type runner are shown in Fig. 2. The main design parameters of lift-type turbine are shown in Fig. 2. DN [m]ddThe diameter of the pipe R [m]ddThe radius of the base circle s[-]ddRotor solidity c [m]ddChord length of the blade airfoil
In this design, the pipe diameter DN is 200 mm. Considering the maximum utilization of the inner space of the pipe and more power generation, the radius R of the turbine is as large as possible, and sufficient clearance is retained in the pipe for consideration of high fluid permeability and low pressure loss. The concept of rotor solidity was introduced into the pipeline turbine here, which indicates the ratio of the sum of the blade chords on the equatorial plane (the section perpendicular to the turbine shaft and through the turbine centerline) of the spherical rotor to the circumference of the base circle. By consideration of the blade inclination g the rotor solidity of a lift-type turbine is defined as,
s¼
Nðc=cos gÞ R
(1)
where N is the number of blades, c is the chord length of the blade
W. Yang et al. / Energy 188 (2019) 116070
airfoil, and g is the inclination of the blade. Since the blade inclination angle is limited by the upper and lower support, the rotor solidity cannot be too large, and the influence of the number of blades and the chord length of the blade are important. The number of blades commonly used in the vertical axis wind turbines are 2 for “F“type and 3 for “H” type. The existing lift-type pipeline turbine developed by Lucid Energy Inc. has a runner with four blades [13]. Considering the smaller average flow rate of the pipeline turbines studied here, the number of blades were increased to six. And attention should be payed to the corresponding blade chord length to ensure the rotor solidity is within a reasonable range. Besides, the blade chord length should not be too small to ensure the strength of the blade. The blade inclination angle g as shown in Fig. 2 allows the blades to lie at different phase angles on different sections perpendicular to the axis, and mitigates the influence of the limited blade number. This parameter is similar to the blade helix angle that appears in the vertical axis wind turbine and the tidal current turbine. Since the trajectory of the blade is on the spherical surface rather than the cylindrical surface, the large inclination will make the effective height of the blade too small. So, the angle of the blade inclination angle should not be too large, and its influence to rotor solidity is smaller than the blade number and the chord length of the blade. Finally, the blade inclination angle was set to 15 here. And the standard airfoil NACA 0018 was used. As a result, the design parameters of the final lift-type turbine are shown in Table 1. 2.2. Drag-type turbine The drag-type pipeline turbine is evolved from the Savoniustype runner [14] which usually does not change the diameter of the runner in the axial direction. However, when the Savonius runner is applied in water pipeline, the diameter of the runner in the axial direction has to be variable to adapt to the shape of the pipe. The main structure of the present drag-type turbine is shown in Fig. 3. The main design parameters of the drag-type turbine are shown in Fig. 4. D [m]ddThe diameter of the runner N [-]ddNumber of blades s [m]ddThe diameter of inner circle s/DddOverlap H [m]ddThe height of runner W [ ]ddBlade chord angle It can be seen that for the Savonius wind turbine used in open environment, the increase of blade number does not bring an improvement in performance. When the drag-type runner is used in the water pipeline, it exhibits the opposite characteristics [6]. Whether it is a hollow (non-overlapping) or a non-hollow (overlapping) runner, the output power increases with the blade number in a certain range. And the blade number can exceed ten which is far more than that of the common drag-type wind turbines. This difference is closely related to that the drag-type runner used in the pipeline is constrained by the pipe wall, in which the fluid is not easy to bypass the runner. Instead, more fluid acts on the blade directly and the fluid permeability of the runner is decreased.
3
Referring to the existing study and taking into account the lower flow rate of the design, the blade number of the runner is set to twelve for the present drag-type turbine. The degree of overlap s/D is an important parameter affecting the performance of the drag-type wind turbine. When s/D is negative it is called non-hollow runner with overlapping blades. When s/D is positive it is called hollow runner with a void space between the runner blades as shown in Fig. 4. A certain degree of the blade overlapping can help to direct the fluid to the concave surface of the blade. However, as the number of blades increases, the flow capacity will decrease and the hydraulic loss will increase. Therefore, in the present case where the blade number is large, the hollow runner with s/D ¼ 0.5 was used to improve both the flow capacity and hydraulic loss of the turbine. When the drag-type runner is applied in the water pipeline, the pipe wall will hinder the discharge of the fluid at zone ④ as shown in Fig. 5. The high-pressure zone b is enlarged, and the adverse effect of the leeward blade is intensified. Therefore, some scholars have studied using a flow guiding device to direct the fluid to the windward blade to reduce the adverse effects of the leeward blade [6]. The purpose of the flow guiding device is to direct the flow to the windward blades and to increase the flow velocity by contracting the flow passage, which has similar effect as the inlet nozzle for drag-type hydro turbine [15]. An eye-type deflector baffle which shows a good performance [6] was used in the present case for the drag-type turbine. It should be noted that a stagnant water area can be easily generated between the baffle and the leeward blade, and large-scale vortices are easily generated in this area. The deflector baffle consists of a main baffle and a sub baffle as shown in Fig. 6(b). The baffle is used to direct the flow and increase its velocity, and the sub baffle wraps the runner to prevent the generation of large-scale vortices in the stagnant water area on the back side of the main baffle. For the deflector baffle the shrinkage rate ε is defined as the ratio of the flow path area blocked by the baffle to the cross-sectional area of the pipe. Large shrinkage rate is beneficial to guiding flow while small shrinkage rate is good for decreasing the local hydraulic loss. In this design, the deflector baffle shrinkage rate was optimized to be 0.58. Following the above analysis, the main design parameters of the drag-type turbine were determined as listed in Table 2. 3. Hydraulic performances of both the turbines 3.1. Test rig According to the test requirements, a pipeline turbine test bench was built to test the hydraulic performances of the turbine. The schematic diagram of the test bench circuit is shown in Fig. 7, which mainly includes a pipeline turbine test section, centrifugal pump, water tanks and electromagnetic flowmeters, brake, torque meter, pressure sensors, valves and so on. The brake is used to exert load for the tested turbine. Two pressure sensors were installed upstream and downstream of the turbine respectively. And the pressure test accuracy is 0.25%. The range of the torque meter is 0e2 N m with test accuracy of 0.5%. The brake is a magnetic remanence brake with rated torque of 2 N m and maximum rotational speed of 10000 r/min. The flowmeter is a three-electrode electromagnetic flowmeter with a measuring range of
Table 1 The main parameters of lift-type turbine. Number of blades
Chord length of airfoil blade
Blade inclination
Rotor solidity
Airfoil
6
26.60 mm
15
1.80
NACA 0018
4
W. Yang et al. / Energy 188 (2019) 116070
Shaft
Pipe
v
Blade Fig. 3. The main structure of the drag-type turbine.
s
Outer contour cutting line
D
Blade arc
(a) Front view
Fig. 4. The main design parameters of the drag-type turbine.
(b) Side view Fig. 6. Deflector baffle for drag-type pipeline turbine.
Fig. 5. Drag-type runner used in water pipeline.
14e282 m3/h, and its accuracy is 0.5%. The inlet and outlet valves of the test section are butterfly valves. The flow rate of the pipeline is controlled by adjusting the rotational speed of the pump. The turbine load is controlled by adjusting the current of the brake. The flow rate was directly read by the display on the flowmeter panel. The pressure, rotational speed and torque were collected by a selfdeveloped integrated data acquisition system. The test bench layout is shown in Fig. 8. The test was carried on two parts. One part is to measure the external characteristics of the turbine such as rotational speed, flow rate, pressure loss, torque and so on. Starting from no-load with the flow rate fixed, the current of the brake is gradually increased to
W. Yang et al. / Energy 188 (2019) 116070
5
Table 2 The main parameters of the drag-type turbine. Outer diameter D
Inner diameter s
Height of runner H
Blade chord angle W
Shrinkage rate of the baffle ε
12
180 mm
90 mm
104 mm
15
0.58
Pressure Sensor_in
Pressure Sensor_out DN200*100
Turbine
Butterfly valve_out Torque meter
DN200
Butterfly valve_in
Water tank_out
Brake DN200*100 Reducer pipe
Electromagnetic Electromagnetic Flowmeter valve 1
Number of blades N
Electromagnetic Centrifugal pump valve 2 Water tank_in Fig. 7. The schematic diagram of the test bench circuit.
pump Brake Torque meter
Three pressure measuring points
H¼
pi po rg
(2)
P¼
2pMn 60
(3)
h¼
P
rgQH
(4)
where r is the fluid density, g is the gravitational acceleration, and they are treated as constant in the calculation. Here the statistical uncertainty was applied to evaluate the error of the experimental results. The statistical uncertainty u was estimated through the standard deviation from the arithmetic mean value of each measurement by
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u m X 2 u 1 uðxÞ ¼ t x x mðm 1Þ j¼1 j
(5)
where m is the time of the measurement, x is the measured variable, x is the arithmetic mean value of each measurement and it is P calculated by x ¼ m j¼1 xj =m. Since the pressure loss H, the power output P, and the efficiency h are not measured directly from the experiment, transfer formula is necessary to estimate the uncertainties of these parameters. According to equation (2), equation (3) and equation (4) the uncertainty transfer formula for pressure loss H, power output P and efficiency h can be expressed as
UðHÞ ¼
Turbine test section
100%
uðHÞ ¼ H
qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi uðpi Þ2 þ uðpo Þ2
0 1 uðPÞ U @P A ¼ ¼ P
(6)
H sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi uðMÞ 2 uðnÞ 2 þ n M
(7)
Fig. 8. The test bench layout.
0 1
enlarge the load on the turbine, and the external characteristics of the turbine are measured under several operating points. The rotational speed of the turbine, the torque, the inlet pressure, and the outlet pressure were recorded during this process. Second part is to test the turbine startup performance. The flow rate is adjusted gradually from low to high under no-load condition to observe the startup performance of the turbine, and the critical flow rate of the turbine startup was recorded. The critical flow rate of the turbine stop was also recorded by decreasing the flow rate under on-load condition.
uðhÞ U @hA ¼ ¼
h
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi ffi uðPÞ 2 uðHÞ 2 uðQ Þ 2 þ þ P H Q
where u(.) represents the uncertainties of the measured variables
3.2. Uncertainties of the experimental results In the experiment the inlet pressure of the turbine pi, the outlet pressure of the turbine po, the rotational speed n (rpm), the impeller torque M and the flow rate Q are measured. Based on these measured variables the pressure loss H of the turbine, the power output P of the turbine, and the efficiency h can be calculated by
(8)
Exit part Inlet part
Rotation zone
Fig. 9. The schematic diagram of the computation domain.
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W. Yang et al. / Energy 188 (2019) 116070
(a) Lift-type turbine
(b) Drag-type turbine
Fig. 10. Mesh distributions of the rotation zone.
calculated by equation (5), U(.) represents the relative uncertainties of the measured variables, and the overbar represents the arithmetic mean value of the measured variables. By using expression (6), expression (7), and expression (8) the experimental results of the pressure loss H, power output P and efficiency h and their relative uncertainties for lift-type turbine and drag-type turbine are shown in Fig. 13 and Fig. 17 respectively. And the maximum of the relative uncertainties is no more than 5%.
Fig. 11. Mesh of the inlet section (drag-type).
Fig. 12. Mesh of the exit section.
3.3. CFD settings 3.3.1. Computation domain and grid The whole Computation domain of both the lift-type and dragtype turbines is consistent, which is a cylindrical section of water in the pipeline as shown in Fig. 9. The upstream distance which is three times of the pipe diameter from the axis was set as the inlet part of the computation domain. Considering the full development of the flow, the downstream distance which is ten times of the pipe diameter from axis was set as the exit part of the computation domain. The computation domain was divided into three parts, a spherical rotation area with a rotor, an inlet pipe section (including an upstream pipe section and a portion around the rotation zone), and an outlet pipe section. A sphere that wraps the rotor was set as the rotation zone as shown in Fig. 9. Unstructured mesh was used in the inlet section and the rotation zone, while structural mesh was used in the outlet section. In the rotation zone where the runner is located, a finer mesh is required near the surface of the blades as shown in Fig. 10. The mesh for lift-type turbine should be fine enough at both the blade leading edge and trailing edge due to the large difference between the blade chord length and the overall size of the runner. The rotation zone mesh of both the lift-type and drag-type turbine is shown in Fig. 10. Since the rotation zone is located in the inlet section, there is a small gap between the interface and the pipe wall. The mesh size of the gap was generated to ensure that there were at least five layers in the gap. The drag-type turbine has a deflector structure and the mesh near the junction of the main baffle and the sub baffle were refined. Since the downstream flow of the runner is complicated the grid with a smaller scale was applied there. The mesh of the inlet section and exit section are shown in Figs. 11 and 12 respectively. The total number of the mesh for the lift-type and the dragtype turbine are 2.17 million and 2.84 million respectively.
3.3.2. Turbulence model and boundary conditions The ANSYS CFX was used for unsteady calculation. Since the flow inside the pipeline turbine has strong rotation effects, the RNG k ε model was selected for its advantages in considering the rotation and swirl effects in the averaged flow. A high-precision difference scheme was applied for the convection term, and a second order difference scheme was used for the diffusion term. The maximum iterations in each time step were set to 20, and the residual of 104 was set as the convergence criterion for all the simulated variables. The time step was determined according to the Courant number in the calculation process, and the average
W. Yang et al. / Energy 188 (2019) 116070
(a) V=0.79m/s (Re=1.57×105)
7
(b) V=1.00m/s (Re=1.99×105)
(c) V=1.29m/s (Re=2.56×105)
Angle of attack [-]
Fig. 13. Power, efficiency and pressure loss curves of the lift-type turbine under different flow rates (P is the power, H is the pressure loss, and h is the efficiency).
Fig. 14. Velocity triangle analysis of the axial section of the runner (v is the incoming flow velocity, u is the angular velocity at which the rotor rotates, u is the peripheral velocity, and w is the relative velocity, a is the attack angle, q is the rotating angle of the runner after time t).
25 20 15 10 5 0 -5 -10 -15 -20 -25
n=200 n=250 n=300 n=348.2 n=439.9 n=480
=0°
0
60
120
180 240 Degree
300
360
Fig. 15. Angle of attack at different blade phase angle based on Eq. (3).
Courant number was kept less than unity. The average value of the y plus near the blade surface was from 30 to 50 and a standard wall function was used. It should be noted that during the transient calculation process, the rotational speed of the runner was fixed for simplicity, and the hydraulic performance of the turbine at different rotational speeds was calculated.
A velocity inlet condition was set at the inlet of the domain with the velocity direction normal to the inlet surface. A static pressure was set at the outlet of the domain. A non-slip condition was set at all the walls in the domain. And a moving mesh scheme was used for the unsteady flow simulation in the present domain including both stationary and rotation parts.
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W. Yang et al. / Energy 188 (2019) 116070
0.4 n=200 n=250 n=300 n=348.2 n=439.9 n=480
0.3
M [Nm]
0.2 0.1 0.0 =0°
-0.1 -0.2 -0.3
0
60
120
180 240 Degree
300
360
Fig. 16. Numerical results of single blade torque at different blade phase angel.
3.4. Experimental and numerical results 3.4.1. Lift-type turbine (1) External performances The power, efficiency and pressure loss curves of both the lift-
(a) v=0.79m/s (Re=1.57×105)
type and drag-type turbine under three different flow rates (Reynolds numbers) are shown in Figs. 13 and 17 respectively. These three flow rates were selected based on the work condition of the turbine. The numerical simulation was performed at the Reynolds number of 1.57 105 with corresponding flow velocity of 0.79 m/s for comparison. Obvious difference between the experimental and numerical results can be seen from Figs. 13(a) and 17(a). Three causes may contribute to the present simulation error. First one is the error from geometry simplification. For lift-type turbine the geometry simplification was performed at the pallet installed on the axis near the pipeline wall to avoid small gap between the pallet and the wall. For drag-type turbine both the axis and the pallet structure were simplified for convenience of grid generation. Second one is the error from uniform boundary conditions setup. Uniform boundary condition was set at both the inlet and outlet of the turbine which might not be the case for the experiment since flow disturbance may happen during the experiment. Third one is the error from the RNG k-epsilon turbulence model. This model is based on the eddy viscosity assumption which has limitations in the simulation of turbulence flow with rotation and curvature. However, the main characteristic of the simulated curves is consistent with the experimental results and qualitative analysis can be conducted based on the simulation results. The rotational speed of lift-type runner is in the range of 200 r/ ~ 700 r/min. The maximum efficiency of the turbine at each flow min rate is higher than 30%. At two large flow rates of V ¼ 1.00 m/s and V ¼ 1.29 m/s, the highest efficiency is more than 40%. The leftmost point of each test curve in Fig. 13 is the brake point of the turbine, and the corresponding rotational speed is the brake speed, under
(b) v=1.00m/s (Re=1.99×105)
(c) v=1.29m/s (Re=2.56×105) Fig. 17. Power, pressure loss and efficiency curve of the drag-type turbine (P is the power, H is the pressure loss, and h is the efficiency).
W. Yang et al. / Energy 188 (2019) 116070
which the turbine stops. The characteristics of the lift-type turbine that is easily braked under the condition of large load are consistent with its poor startup performance. Both the brake and startup performance are caused by the decrease of working ability when the blades work at a high angle of attack under low tip speed ratio. Both the efficiency and power of the turbine under each Reynolds number decrease with the increase of the rotational speed, while the pressure loss increases with the increase of the rotational speed. In order to further analyze the power characteristics of the lifttype turbine, it was assumed that the flow between the blades did not interfere with each other, regardless of the flow influence between the axial sections of the blade. The velocity triangle of the axial section of the runner is shown in Fig. 14, where v is the incoming flow velocity, u is the angular velocity at which the rotor rotates, u is the peripheral velocity, and w is the relative velocity. In a rotating reference frame, the absolute velocity is equal to the vector sum of the relative velocity and the convected velocity. The angle of attack a is the angle between the relative speed w and the circumferential direction. Although the direction of the flow velocity v in the pipe does not change, the angle between the peripheral velocity u and v changes constantly as the rotor rotates. In the velocity triangle, the sine theorem can be used as,
v u ¼ sin a sin½90 a ð90 qÞ
(2)
where q is the rotating angle of the runner after time t and it is defined by q ¼ ut. The formula of the attack angle a can be derived from Eq. (2),
a ¼ arctan
sin q cos q þ l
(3)
where l ¼ u=v is the tip speed ratio. It can be seen from Eq. (3) that the angle of attack is only related to the position of the blade and the tip speed ratio of the runner. The variation range of the attack angle decreases with the tip speed ratio of the runner and the stall is not likely to occur. When the tip speed ratio of the runner increases, the risk of the airfoil working at large angles of attack will be increased accordingly. The numerical results about variation of the single blade torque with the blade phase angle under the Reynolds number of 1.57 105 are shown in Fig. 16, and the corresponding theoretical angle of attack given by Eq. (3) are shown in Fig. 15. The airfoil used in the blade is a symmetrical airfoil. The change of the theoretical angle of attack is symmetrical, while the variation of the torque got by simulation is asymmetry in a rotation period of the runner. Its positive peak is located in the middle of the upstream, which corresponds to the phase angle of the maximum positive angle of attack. However, when the blade reaches the theoretical maximum negative angle of attack, there is no torque output, which may be related to the complicated flow field when the blade is located in downstream. At this time, the blade passes through the low-speed area in the middle of the runner, and the actual angle of attack of the blade is far below the theoretical angle of attack. When the blade is located between 50 150 , the output torque is large, and the theoretical angle of attack of the blade exceeds 5 . When the rotational speed exceeds 300 r/min, the phase angles of the torque peak value under different rotational speed are close. When the rotational speed decreases from 300 r/min, the maximum angle of attack is too large, and the torque peak shifts to the low phase angle, as a result the maximum torque is reduced. When the blade is located between 180 360 , the negative torque generated by the blade in the downstream region increases with the rotational
9
speed. At this time, the blade passes through the low-speed region of the runner, and the low-speed area gradually expands with the increase of the rotational speed and the flow is more chaotic, which causes generation of the negative torque. And these results are consistent with the turbine power characteristics shown in Fig. 13 obtained in experimental tests. (2) Startup characteristics The startup of the lift-type pipeline turbine is an unstable process with the vibration of the runner and flow uncertainty. Therefore, it is difficult to measure the critical startup flow rate steadily and accurately. However, it can still provide a valuable reference for the startup characteristics study of the lift-type turbine. The minimum startup flow rate v0 of the turbine is close to 0.69 m/s which is a relatively large value for the turbine application conditions in water pipeline system. The critical startup process was observed at the flow rate of v0 ¼ 0:69m=s, which can be divided into five stages: i. The runner is shaking dramatically with intermittent slow rotation. ii. Intermittent slow rotation makes the runner reach a phase angle with a large torque, then the runner is accelerated. iii. The accelerating runner does not stop before it reaches the next phase angle with a small torque. iv. The runner changes from intermittent rotation to continuous slow rotation, retaining continuous low speed rotation and slow acceleration. v. The runner passes the startup process and accelerates quickly to reach a steady operation condition. The startup condition of the turbine is that the startup torque is greater than the drag torque. Due to the dynamic change of the blade angle of attack during the rotation of the lift-type runner, the runner torque is also dynamically changed. Therefore, the startup torque will change with the different phase position of the runner, and it may be greater or less than the drag torque, which shows the above stage i and stage ii. When the runner is in the intermittent slow rotation, the rotation will make the runner rotate to a phase angle with a larger startup torque. Whether the runner can successfully reach the stage iii through the phase angle with a small startup torque is the first key for the successful startup of the turbine. After the runner passes the intermittent slow rotation, it will enter the rapid acceleration stage and before reaching the steady operation condition it is in the stage iv. At this stage, the tip speed ratio of the runner is low, the airfoil may have a stall and the generated torque is small. Whether the runner can pass the possible stall state to the stage v is another key issue in the startup of the lift-type turbine. The critical stop flow rate for lift-type turbine was also recorded by adjusting the flow rate from high level to low level, which is about 34% of the critical startup flow rate. 3.4.2. Drag-type turbine (1) External performances The power, pressure loss and efficiency test results of the dragtype turbine under three Reynolds numbers are shown in Fig. 17. The numerical simulation was performed at the Reynolds number of 1.57 105 for comparison. In each operating condition, the rotational speed is significantly less than that of the lift-type turbine. From the form of the power and efficiency curves, there is a peak value of the output power and no brake point at any rotational speed. The efficiency of the drag-type turbine is lower which does not exceed 20% under all operating conditions. The results obtained
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W. Yang et al. / Energy 188 (2019) 116070
by the simulation of the drag-type turbine are quite different from the experimental values, as shown in Fig. 17(a), which was also reported in Ref. [6]. The pressure loss curve of the drag-type turbine changes gently, and the variation trend of the power curve and the efficiency curve is similar, which is shown in both the simulation and test results. When the rotational speed is low, the output power is low because the fluid bypasses the rotor or passes through the center of the runner directly, without pushing the blade to work. When the rotational speed is large, the flow field structure is disordered, and the negative torque increases. At the same time the efficiency of the turbine and the output power are reduced. The zero-phase angular position of the drag-type turbine blade is shown in Fig. 18. The single-blade torque of the runner rotating at different speeds is shown in Fig. 19. When the rotational speed is more than 100 r/min, the blade generates the maximum torque at the phase angle of 100 200 as shown in Fig. 19. This position is the outlet of the deflector baffle at which the outflow is working directly on the blades, and the maximum torque decreases with the increase of the rotational speed. The negative moment increases when the blade rotates to the junction of the main baffle and the sub baffle. When the rotational speed is less than 100 r/min, the torque peak position appears after 280 as shown in Fig. 19, that's because the fluid fails to do work when it first passes the runner, while the fluid does work when it leaves the runner. It can be seen that although the fluctuation of the single-blade torque is large, the
generation of the negative torque of the drag-type turbine with the deflector baffle is suppressed, and there is no large-scale negative moment region compared with the lift-type turbine. (2) Startup characteristics The startup performance of the drag-type turbine was tested with a slow increase in the flow rate. The critical startup flow rate of the drag-type turbine is 0.18 m/s which is significantly lower than that of the lift-type turbine. The drag-type turbine is stable whether it is starting or stopping, and there is no strong shaking of the runner. The drag-type turbine blade has a large windward projection area, and is subjected to a large force of water flow in a stationary state, which is advantageous for the startup performance. The critical stop flow rate was also recorded, which is about 77% of the critical startup flow rate for the drag-type turbine. 3.4.3. Comparisons In general, the rotational speed of the drag-type turbine is significantly lower than that of the lift-type turbine at the same flow rate, and the lift-type turbine suddenly brakes when the load is increased to a certain value. In order to compare the characteristics of both the lift-type and drag-type turbine with different flow rates and different rotational speeds, the rotational speed of the turbine is nondimensionalized as
l¼
2pRn v
(4)
where l is the tip speed ratio, R is the runner diameter, n is the rotational speed, and v is the incoming flow velocity. The power factor is defined as
CP ¼ 1
P
(5)
3 2 rAv
where A is the cross-sectional area of the pipe, P is the output power of the runner, v is the velocity of the incoming flow, and r is the density of the water. Fig. 20 shows the variation of the power coefficient of both the lift-type and drag-type turbine with the tip speed ratio. Under the different flow rates, the power coefficient curves of the drag-type turbine tend to coincide, and they are Fig. 18. Zero-phase angular position of the drag-type turbine blade.
v=0.34 v=0.50 v=0.79 v=1.00 v=1.29 Solid: lift-type hollow: drag-type
3.5 3.0
0.8
2.5 CP [-]
M [Nm]
0.4 0.0
1.5 1.0
n=40 n=70 n=100
-0.4 -0.8
2.0
0
60
120
180 240 Degree
0.5
n=130 n=160 n=180
300
Fig. 19. Single-blade torque of the runner rotating at different speeds.
0.0 360
-0.5
0
1
2
4
[-]
5
6
7
8
Fig. 20. Power coefficient curve of both the lift-type and drag-type pipeline turbine.
W. Yang et al. / Energy 188 (2019) 116070
v=0.34 v=0.50 v=0.79 v=1.00 v=1.29 Solid: lift-type hollow: drag-type
0.8
H [m]
0.6
0.4
0.2
0.0
0
1
2
4
[-]
5
6
7
8
Fig. 21. Pressure loss curve of both the lift-type and drag-type pipeline turbine.
consistent for the lift-type turbine. The characteristics of the lifttype vertical axis pipeline turbine and the drag-type vertical axis pipeline turbine were compared with the corresponding type of wind turbine [16]. The tip speed ratio of the drag-type pipeline turbine is about unity, which is close to the tip speed ratio of the Savonius type of wind turbine. The tip speed ratio of the lift-type pipeline turbine is in the range of four to seven, which is also close to the Darrieus type of wind turbine. In addition, the multiple relationship between the power coefficient of the lift-type pipeline turbine and the drag-type pipeline turbine is similar to that of the Darrieus-type wind turbine and the Savonius-type wind turbine, which is about a triple relationship between the two turbines. Fig. 21 shows the pressure loss curves of both the lift-type and drag-type turbine with the tip speed ratio. For both turbines, the pressure loss increases significantly with the flow rate, and for the same flow rate of the same runner, the pressure loss increases with the rotational speed. However, the pressure loss of both turbines is less than 3% of the total head in the range of the incoming flow velocity studied, which has little influence on the normal water supply.
4. Conclusions Vertical axis turbines with both the lift-type and drag-type runner for micro power generation in water pipeline system are studied experimentally and numerically. The vertical axis pipeline turbine is directly installed in the water pipeline system to supply power for the data remote transmission equipment. It should meet the requirements of startup at low flow rate and low-pressure loss at high flow rate without affecting the normal water supply. In this paper, both the lift-type pipeline turbine and the dragtype pipeline turbine are designed with reference to the Darrieus wind turbine and the Savonius wind turbine, and the external and startup performances of both the turbines are experimentally measured. It is found that there are five stages during the startup process of the lift-type turbine. The five-stage startup process is unstable and is accompanied by intermittent shaking of the runner, which is caused by the dynamical change of the difference between the startup torque and the drag torque due to the dynamical variation of the attack angle. As a result, the critical incoming flow velocity of the startup process for lift-type turbine is larger than
11
that of the drag-type turbine, which can start at a lower flow rate. And the starting process of the drag-type turbine is stable. Under different incoming flow rates, the power coefficient curves with the change of the tip speed ratio of the lift-type turbine are consistent, while they tend to coincide for the drag-type turbine. The lift-type turbine has a higher speed ratio and higher output power than that of the drag-type turbine. As the turbine load increases, the lift-type turbine will brake at a large incoming flow rate until it stops running, while the drag-type turbine has a significantly lower braking flow rate. It should be noticed that the critical startup flow rate is not consistent with the critical stop flow rate for both the turbines, which indicates that there is hysteresis effect in the flow system especially for the lift-type turbine. In the range of the incoming flow velocity studied, the pressure loss of both the lift-type turbine and the drag-type turbine increases with the incoming flow rate. However, the maximum pressure loss does not exceed 3% of the total water head, which will not affect the normal water supply. In general, both the lift-type turbine and the drag-type turbine can meet the energy requirements of the data remote transmission equipment in water pipeline system, however, there is still room for improvement. For the lift-type turbine, the critical incoming flow rate of the startup can be further reduced, and the stability of the starting process can be improved. For the drag-type turbine, the range of the tip speed ratio can be further expanded, and the output power can be increased. Acknowledgements The author would like to acknowledge the financial support received from National Key Research and Development Plan (grant number 2018YFB0606103) and Beijing Waterworks Group. References [1] Yang X, Ong KG, Dreschel WR, Zeng K, Mungle CS, Grimes CA. Design of a wireless sensor network for long-term, in-situ monitoring of an aqueous environment. Sensors 2002;2(11):455e72. [2] Stoianov IA, Nachman LB, Whittle AC, Madden SD, Kling RB. Sensor networks for monitoring water supply and sewer systems: lessons from Boston. Cincinnati, OH: 8th Annual Water Distribution Systems Analysis Symposium; 2006. [3] Qureshi FU, Muhtaroglu A, Tuncay K. Near-optimal design of scalable energy harvester for underwater pipeline monitoring applications with consideration of impact to pipeline performance. IEEE Sens J 2017;17(7):1981e91. [4] Qureshi FU, Muhtaroglu A, Tuncay K. A method to integrate energy harvesters into wireless sensor nodes for embedded in-pipe monitoring applications. Cairo, EGYPT: 5th International Conference on Energy Aware Computing Systems & Applications (ICEAC); 2015. [5] Du J, Yang H, Shen Z, Chen J. Micro hydro power generation from water supply system in high rise buildings using pump as turbines. Energy 2017;137: 431e40. [6] Chen J, Yang HX, Liu CP, Lau CH, Lo M. A novel vertical axis water turbine for power generation from water pipelines. Energy 2013;54(2):184e93. [7] Mazzolani G, Berardi L, Laucelli D, Simone A, Martino R, Giustolisi O. Estimating leakages in water distribution networks based only on inlet flow data. J Water Resour Plan Manag 2017;143(040170146). [8] Islam M, Ting D, Fartaj A. Aerodynamic models for Darrieus-type straightbladed vertical axis wind turbines. Renew Sustain Energy Rev 2008;12(4): 1087e109. [9] Golecha K, Eldho TI, Prabhu SV. Influence of the deflector plate on the performance of modified Savonius water turbine. Appl Energy 2011;88(9): 3207e17. [10] Fujisawa N, Shibuya S. Observations of dynamic stall on Darrieus wind turbine blades. J Wind Eng Ind Aerodyn 2001;89(2):201e14. [11] Mcalister KW, Carr LW, Mccroskey WJ. Dynamic stall experiments on the NACA 0012 airfoil, NASA TP-1100. NASA; 1978. [12] Carr LW, Mcalister KW, Mccroskey WJ. Analysis of the development of dynamic stall based on oscillating airfoil experiments, NACA TN D-8382. NASA; 1977. [13] Schlabach RA, Cosby MR, Kurth E, Palley I, Smith G. In-pipe hydro-electric power system and turbine. 2011. p. 6e14. United States Patent, US 7959411B2. [14] Akwa JV, Vielmo HA, Petry AP. A review on the performance of Savonius wind
12
W. Yang et al. / Energy 188 (2019) 116070
turbines. Renew Sustain Energy Rev 2012;16(5):3054e64. [15] Kai S, Furukawa A, Okuma K, et al. Experimental study on simplification of Darrieus-type hydro turbine with inlet nozzle for extra-low head hydropower
utilization. Renew Energy 2012;41(2):376e82. [16] Wilson RE, Lissaman PBS. Applied aerodynamics of wind power machines. Nasa Sti/recon Technical Report N; 1974.