Predicted and measured performance of a vertical axis wind turbine with passive variable pitch compared to fixed pitch

Abstract

The performance of a 5-m diameter Darrieus vertical axis wind turbine was predicted using both a double multiple streamtube model and a two-dimensional unsteady Reynolds-averaged Navier–Stokes computational fluid dynamics simulation with constant rotational speed for a series of operational points. The actual performance was measured in both fixed and variable pitch modes. The aims were (1) to compare starting torque and peak efficiency in fixed and variable pitch modes and (2) to test an overspeed control mechanism. Starting torque was approximately three times higher in variable pitch mode and the maximum efficiency on some runs was significantly higher. The overspeed control mechanism functioned consistently as designed. Thus, variable pitch was shown to overcome two major disadvantages of normal fixed pitch vertical axis wind turbines, self-starting and overspeed control. Discrepancies between the predicted and measured results showed the importance of accurately assessing parasitic drag losses and the need for three-dimensional simulation to give reliable performance predictions.

Keywords

Darrieus wind turbine variable pitch double multiple streamtube computational fluid dynamics test data

Introduction

In recent years, there has been a proliferation of small (less than 20 kW), fixed pitch vertical axis wind turbines (VAWTs) on the market (see, for example, Alibaba, 2016), ranging from the relatively conventional with three or more straight or helical fixed pitch blades as shown in Figure 1, to some designs which show little regard for aerodynamic principles.

Figure 1.

Fixed pitch VAWTs: left, 3 m diameter with three helical blades; right, 6 m diameter with six straight blades.

Unsubstantiated and/or unbelievable claims are usually made about their performance. For example, one website that has since disappeared claimed in 2014 that their VAWT with a swept area of 2.8 m² produces 13.1 W in a 2 m/s wind. This would require a coefficient of performance of 0.974, well above the theoretical maximum (Betz limit) of 0.593, even assuming this is shaft power, not electrical power.

It is very difficult to find the actual believable performance data, possibly because small, high solidity fixed pitch VAWTs generally perform very poorly, and also because it is difficult to measure performance. Very few wind tunnels are big enough to accommodate a realistic sized wind turbine, although Armstrong et al. (2012) describe one exception, the 8 × 5 m working section at the University of Waterloo, Canada, and small-scale models in small wind tunnels operating at low blade chord Reynolds numbers do not accurately represent the performance of full-scale turbines. Short-term fluctuations in wind speed make field data difficult to interpret (Kirke, 1998). Models of fixed pitch VAWTs predict very low starting torque and low efficiency (Kirke and Lazauskas, 1991), and observations of the actual VAWTs in service tend to support these predictions. Performance data for hydrokinetic turbines (HKTs) that are driven by flowing water in the same way that wind turbines are driven by moving air also tend to support these predictions (Kirke, 2011).

Variable pitch can provide adequate starting torque and overspeed control and improve the efficiency of high solidity turbines which do not reach high enough tipspeed ratios (TSRs) to avoid blade stall without blade pitching. However, active variable pitch requires relatively sophisticated instrumentation, microcontrollers and stepper motors which are too expensive for small turbines. Several variable pitch mechanisms have been patented (Drees, 1979; Herter Rotor GMBH, 1986; Nemec, 2001; Pryor, 1983), but few of these have been proven and none has reached the market as far as the authors are aware.

B.K.K. has designed and constructed several prototype passive variable pitch VAWTs in the past (Kirke, 1998), but although starting torque was adequate, efficiency was poor. However, more recent modelling of HKTs and VAWTs with a simple sinusoidal pitch regime (Lazauskas, 2008; Paillard, 2014) predicts a big improvement over fixed pitch, and promising performance has been measured on a 1-m diameter HKT with sinusoidal pitch.

Computational fluid dynamics simulation

The open-source solver OpenFOAM was used for this study. The simulations were performed for fixed pitch and two variable pitch regimes, with constant rotational speed for each operational point. Additional corrections were used to estimate the performance of the corresponding three-dimensional (3D) turbine. The main purpose of these simulations was to predict the performance of a proposed new pitch control mechanism which has not yet been built, to assess whether likely performance improvements would justify building and testing a prototype. In this article, comparisons between the simulations and actual measured performance are used to assess whether the simulations would provide reliable predictions. Comparisons are shown in Appendix 3.

Simulation strategy

An unsteady two-dimensional (2D) Reynolds-averaged Navier–Stokes (RANS) simulation was carried out. The rotating motion of the blades was obtained by Arbitrary Mesh Interface, which enables a fluid to travel through two sliding interfaces even if the nodes do not correspond, as shown in Figure 2. A main interface was used for the global turbine rotation, inside of which three interfaces were used for the pitch variation of the blades. The resulting domain consists of five regions.

Figure 2.

CFD domain showing regions and interfaces.

A constant rotational velocity was imposed on the rotor, according to the desired TSR. The blade pitch angle was then controlled by a look-up table which was updated on the fly in order to maintain a constant angle of attack at quarter chord. This angle was simply computed by probing the flow on the trajectory of the blades at each revolution and adding the rotational velocity components.

Spatial discretization

The turbine modelled has three blades of NACA 0018 section, chord length 0.3 m, turbine radius 3 m, that is, solidity nc/r = 3 × 0.3/3 = 0.3. The simulations were done for 4 and 10 m/s wind. The computer-aided design (CAD) and the mesh of the fluid domain were built with the open-source software Salome. The turbine was placed in a rectangular domain with 10 diameters in the upwind and transverse directions and 20 diameters in the downwind direction. The global rotation domain size was 3 radii wide, and the blade rotating domains were 2/3 radius wide, in order to reduce the velocity gradient convected through the interfaces. The configuration is shown in Figure 2.

The mesh was hybrid, with a structured mesh around the aerofoil for boundary layer resolution, and an unstructured mesh was used elsewhere, with triangle elements. The 2D simulation was carried out on a 3D mesh with a single-cell extrusion; hence, hexahedra are obtained in the boundary layer and prisms are obtained in the unstructured mesh. The ratio of prism to hexahedra is approximately 5:1. A total of 100,000 points were used for these runs. Approximately 24,000 points are used for the rotating blade domains, 17,000 points were used for the rotor domain and 11,000 for the outer domain. At each rotating blade interface, 200 points were used, and at the rotor interface 600 points were used. Along each blade chord, 340 points were used. The cell size along the blade surfaces was 0.3 mm, which gives y+ values along the blade chord between 0.5 and 3. This y+ value was checked with steady NACA0012 RANSE (Reynolds-Averaged Navier–Stokes Equations) results against experimental data and various wall functions, as shown in Figure 3. Further details on this validation process can be found in Paillard (2015). The expansion ratio in the boundary layer was 1:2. The results were found to be insensitive to further mesh refinements.

Figure 3.

Lift and drag coefficients of a steady 2D NACA0012 at 8° and Re = 6e6, as a function of maximum y+. Various wall functions implemented in OpenFOAM were used.

The mesh quality characteristics were the following. The maximum aspect ratio (AR) was 350, the maximum non-orthogonality was 30 and the maximum skewness was 0.35.

Temporal discretization/wake convergence

The azimuthal step was fixed at a value of 0.25° for TSR = 0.5, 0.5° for TSR = 1, 2 and 3 and 1° for TSR above 4. The resulting maximum Courant number was kept below 50. Again, this was checked with pitching NACA0012 RANSE results against experimental data, as shown in Figure 4. Further details on this validation process can be found in Paillard (2015). The results were found to become insensitive to time step change below this value. The wake convergence was obtained after 5–10 revolutions, depending on the TSR value. The total number of time steps is thus between 5000 and 7000.

Figure 4.

Unsteady drag coefficients of a pitching 2D NACA0012 at a reduced frequency of 0.1 and Re = 2.5e6, for various pitching amplitudes.

Boundary and flow conditions

The fluid considered is incompressible, with a kinematic viscosity of 1.511e−5 m²/s and a turbulence intensity of 0.5%. The inlet boundary condition was imposed through a zero gradient condition on pressure and a fixed value condition on velocity. The outlet boundary condition was imposed through zero relative pressure and zero gradient on velocity. The top and bottom boundary conditions are slip conditions. At all sliding interfaces, an AMI (Arbitrary Mesh Interface) formulation was used.The solid boundaries on all blades have a relative velocity fixed at zero and a zero pressure gradient condition.

Turbulence model

A Spalart–Allmaras turbulence model was used with no wall function at solid boundaries (Spalart and Allmaras, 1992). According to the inlet turbulence characteristics, the modified turbulent viscosity ν~ is fixed at the inlet at 3.57e−05.

Schemes and convergence criterion

A backward, second-order scheme is used for temporal equations. A second-order Euler scheme is used for spatial equations. The hybrid PISO-SIMPLE algorithm is used for resolution. The PISO (Pressure Implicit Split Operator) algorithm is designed for unsteady cases while the SIMPLE (Semi Implicit Methods Pressure Linked Equations) algorithm is designed for steady-state cases. The convergence criterion is 10e−5 for pressure and 10e−6 for all other quantities.

3D corrections

Corrections are applied to the results in order to account for the 3D effects encountered in the real turbine. These include the following:

AR effect on lift coefficient;

Arm influence: (1) tangential drag force on arms and (2) junction drag. The arm–blade junctions increase the effective solidity locally at the tips of the blades and are a source of vortex generation in the tip area.

Lift correction

The lift coefficient is modified according to the following equation, after Katz and Plotkin (2004)

C_{L, 3 D} = \frac{C_{L, 2 D}}{(1 + 2 / AR)}

(1)

where AR is the aspect ratio of the blades (span/chord length), and AR = 10.

This correction is applied, in spite of the end-plate effect provided by the arms, for two reasons: First, the arms increase the solidity locally at the tips of the blades and are source of vortices’ generation in the tip area. This is similar to the behaviour encountered with a finite wing, and this influence decreases when the AR increases. Second, the influence of the arms on the transverse flow and on the downwash associated with a finite wing is not clear, and it will not reduce the 3D effect to zero. This is why it was decided to use this correction, even though connecting the arms to the blade ends aims at reducing this 3D effect.

Arm and arm–blade junction drag loss

Two radial arms are used for each blade. They are connected to the blade near the tip, and they are oriented 30° to horizontal, up for the top arm and down for the bottom arm. Their aerodynamic behaviour has an impact on the torque of the turbine. This impact is obtained through three corrections: (1) equivalent tangential force, (2) arm drag and (3) junction drag.

Arm drag

The equivalent tangential force is obtained by the integration of the tangential force along the blades

{Δ C}_{P, arms} = (2 \tan (30) / R \cdot L) \int_{0}^{R} C_{P} (r) \cdot r \cdot dr

(2)

where C_P(r) is the power coefficient for the fixed pitch turbine, as a function of the radius, thus as a function of TSR.

The arm influence on the tangential torque is a function of the fixed pitch turbine performance; thus, the values will be presented in the next section. The power consumption from the arm drag is obtained by integration along the arms

{Δ P}_{drag} = 6 \cdot \frac{1}{2} ρ \cdot (\frac{c}{\cos (30)}) \cdot C_{d, section} \cdot ω^{3} \cdot \frac{r^{4}}{4}

(3)

where C_d = 0.015 at the Reynolds number considered (Abbott and Von Doenhoff, 2012).

This power reduction is not a function of the turbine performance. The values as a function of the TSR are presented in Figure 5. The values given are absolute: ΔC_P stands for the non-dimensional power coefficient variation.

Figure 5.

Arm parasitic drag loss as a function of tipspeed ratio TSR.

Junction drag

This is obtained by the use of the following thickness drag coefficient after Hoerner (1965)

C_{dt, junction} = 17 \cdot {(\frac{t}{c})}^{2} - 0.05

(4)

{Δ D}_{junction} = C_{dt, junction} \cdot Q \cdot t^{2}

(5)

{Δ P}_{junction} = 6 \cdot Δ D \cdot rω

(6)

where t is the thickness, c is the chord, ΔD_junction is the drag increment, Q is the dynamic pressure, Q = 1/2ρV_rel² and ΔP_junction is the power consumed by the junctions. This junction drag coefficient can be used for a plain, straight connection between the arms and the blades. Any other geometry requires another coefficient. Like the arm drag, this power reduction is not a function of the turbine performance. The values as a function of the TSR are shown in Figure 5. It will be seen that both arm and junction losses increase steeply as TSR increases.

Arm wake interference with inflow on downwind pass

The arm wake is likely to have a negative effect on the flow over the blades on their downwind pass, but this is difficult to assess and was not attempted.

Prototype variable pitch VAWT design and construction

A 5-m diameter VAWT was built with three straight fibreglass blades of NACA0018 profile, 3 m long, tapering from 400 mm chord at midspan to 200 mm chord at the rounded tips to minimize tip losses. These are mounted at 21% span from each end to minimize blade bending moments, on 2.5-m tapered steel box section arms with sheet aluminium fairings. As shown in Figure 6, the arms radiate outwards from a single hub to minimize bending moments on the shaft. Taking the average chord length as 0.3 m gives solidity nc/r = 3 × 0.3/2.5 = 0.36.

Figure 6.

A 5-m diameter wind turbine used for tests, shown on a 4-m high tower in an industrial area surrounded by factory buildings.

The blades are pivoted on rod end bearings forward of their aerodynamic centre, with control cables connecting the trailing edges to a ring which rolls around the turbine shaft with enough clearance to provide the desired pitch amplitude, as shown in Figure 7.

Figure 7.

Schematic representation of pitch control mechanism.

As shown in Figure 7, the blades are pivoted forward of their aerodynamic centre, which is located at the quarter chord point (c/4). Therefore, aerodynamic force acts at all times to pivot the trailing edge downwind, thereby reducing the angle of attack. But because the centre of mass of the blades is well to the rear of the pivots, typically at about 0.45 c, the centrifugal force tends to pivot the trailing edges outwards. On the downwind pass, the aerodynamic and centrifugal forces on a blade act outwards and keep its control cable taut. But on the upwind pass, the aerodynamic force acts inwards while the centrifugal force acts outwards as shown in Figure 7 on the blade at left. For a VAWT of this size, the centrifugal force is typically several times larger than the aerodynamic force, and the moment arm of the centrifugal force is much larger than that of the aerodynamic force, so the net moment on the blades is always outward and the pitch control cables shown in Figure 6 and 7 are always in tension.

For example, the blades of the present turbine have a mass m approximately 10 kg, so at an angular velocity ω of 10 rad/s (95.5 r/min), the governing speed for this turbine, the blade tangential velocity = 25 m/s

Centrifugal force = {mw}^{2} r = 10 \times {(10)}^{2} \times 2.5 = 2500 N

(7)

The centre of mass of the blade is about 0.14 m to the rear of the pivot so the pitching moment M_c of the centrifugal force is given by

M_{c} = 2500 \times 0.14 = 350 N m

(8)

In contrast, the maximum lift force F_L when C_L = 1 and the blade is travelling almost directly upwind in a 10 m/s wind is given by

F_{L} = C_{L} \cdot \frac{1}{2} \cdot ρ \cdot A \cdot {V_{rel}}^{2} = 1 \times \frac{1}{2} \times 1.2 \times 1 \times 35^{2} = 735 N

(9)

where C_L is the lift coefficient, ρ is the air density of approximately 1.2 kg/m³, A is the blade planform area of approximately 1 m² and V_rel is the relative velocity between the blade and air of approximately 25 + 10 = 35 m/s.

The centre of lift is about 0.06 m from the pivot so the moment M_a of the aerodynamic force is given by

M_{a} = 735 \times 0.06 = 44 N m

(10)

Thus, the moment M_a of the aerodynamic force is about 1/8 of the centrifugal force moment so the control cable will be in tension. This can be shown to be the case except when the turbine is starting up, when additional pitching on the upwind pass will occur and will aid starting torque.

The moments on the three blades due to the centrifugal force are balanced by the pitch control cables so the net moment on each blade is provided by the aerodynamic force only, and the blades tend to pitch trailing edge downwind, limited only by the pitch amplitude ring, giving a sinusoidal motion which has been shown by Lazauskas (2008) to give performance close to that of optimized pitch (see Appendix 2).

Overspeed protection was initially provided by preloaded pneumatic cylinders as shown in Figure 6, and later by preloaded hydraulic cylinders which allow the blades to swing outwards when the angular velocity reaches a desired governing speed and the preload is overcome, thereby increasing drag and acting as fail-safe aerodynamic brakes regardless of wind speed.

Governing will occur when the preload is overcome and the blade starts to swing trailing edge outwards on the downwind pass before this occurs on the upwind pass, which will lead to some imbalance. But because the centrifugal moments on all the blades are balanced, the imbalance is caused by the aerodynamic moments only, and these are much less than the centrifugal moment so the imbalance will not be large. All straight blade VAWTs must be designed to withstand some inherent imbalance in the aerodynamic force, so this should not be a problem. And any pitching outside the sinusoidal regime will increase drag and limit power and angular velocity as required.

Governing speed can be adjusted by increasing or decreasing the preload pressure in a reservoir connected to the cylinders and the turbine can be slowed almost to a standstill by opening a valve at the bottom of the tower to release the pressure in the cylinders. This mechanism is protected by Australian and US patents. The turbine has been running for more than a year on a 4-m high tower in an industrial area surrounded by factory buildings as shown in Figure 6 and has been observed to start easily in light winds and accelerate up to speed. But to test its performance, a steady, controllable wind speed was needed.

Performance testing

To test its performance, the turbine was mounted on a 4-m high tower on a truck as shown in Figure 8 and driven at a controlled speed on a deserted road in open country to simulate a steady wind.

Figure 8.

The turbine on the truck.

To check whether the air flow over the cab affected readings, two Wittich and Visser PA2 cup anemometers were mounted in front of the turbine, one at hub height and one level with the bottom of the blades, as shown in Figure 8. The turbine was held stationary by means of a brake until the truck reached the desired speed as indicated by the speedometer, then the brake was released and the turbine was able to accelerate against its own known rotational inertia. Relative wind speed and turbine angular velocity were logged on a Graphtec GL220 midi logger. Runs were done at approximately 20, 30, 40 and 50 km/h (5.5, 8.3, 11 and 13.9 m/s) in both the directions, and the overspeed control worked consistently at all speeds.

It was not possible to maintain an exactly steady speed and there was a cross wind fluctuating between approximately 3 and 5 m/s during some tests, but by logging instantaneous relative wind speed and the corresponding angular velocity it was possible to calculate instantaneous angular accelerations, hence torque and shaft power corresponding to each wind speed reading. Hence, power (C_P − λ) and torque (C_Q − λ) curves could be plotted.

To measure revolutions per minute (RPM), 30 holes were drilled in a circle in a 360-mm-diameter steel disc, which was mounted at the bottom of the shaft. A Moeller LSI-R12M-F2-LD Inductive Proximity Switch was mounted above the disc so it would detect the holes and give 30 pulses per revolution, and the number of pulses per 2 s would give RPM. This worked reliably when tested in a lathe in the factory, but proved unreliable in the field, probably because the disc also functioned as a disc brake and was displaced slightly when the brake was applied and released. The Proximity Switch was only able to detect holes up to 3 mm away. If mounted too far away from the disc, it failed to detect holes reliably, and if too close it rubbed against the disc and could have been damaged. As a result, most of the data were clearly unrealistic and had to be discarded. However, a few runs at about 8 m/s gave believable results and these are presented below.

Test results

Figure 9 shows the comparison between the readings recorded on the two anemometers. It will be seen that there is no consistent difference in readings, indicating that the air flow over the cab did not affect readings significantly. Given that the number of pulses per second corresponds to the relative wind speed in metres/second, and that the nearest whole number of pulses per second is recorded, the absolute error in recorded wind speed is ±0.5 m/s and the relative error is 0.5/8 = 6.25% in an 8 m/s wind. But the torque varies with wind speed squared and power with wind speed cubed, so the relative error on torque is 1/2(1 − (7.5/8.5)²) = 11% and in power it is 1/2(1 − (7.5/8.5)³) = 15.6%.

Figure 9.

Comparison between readings recorded on the two anemometers, one at hub height and other level with the bottom of the blades.

Figure 10 shows the performance coefficient C_P (efficiency) plotted against TSR for the turbine in fixed and variable pitch modes in an approximately 8 m/s wind, and Figure 11 shows the corresponding torque coefficient for same tests.

Figure 10.

C_P (efficiency) versus λ for the turbine in fixed and variable pitch modes.

Figure 11.

Torque coefficient C_Q versus tipspeed ratio λ for the turbine in fixed and variable pitch modes.

Importantly, it will be seen that the starting torque is about three times higher in variable pitch mode and the peak C_P is about 0.28–0.34, with a large amount of scatter, compared to about 0.29 in fixed pitch mode. Figure 12 compares angular velocity versus time for fixed and variable pitch modes, showing that the turbine in variable pitch mode accelerates from the rest about three times as fast as it does in fixed pitch mode, reaching its governed maximum 95 r/min in about 40 s, while in fixed pitch mode it starts very slowly from the rest and takes about 80 s to reach 40 r/min (TSR about 1.3) after which it accelerates rapidly, corresponding to the higher torque coefficient shown in Figure 11. Equally important is the fact that the passive, fail-safe overspeed control worked reliably in variable pitch mode. Such a simple and effective passive overspeed control could not be achieved with fixed pitch, and in the present tests, the disc brake had to be applied to prevent the turbine overspeeding. Aside from blade pitching for overspeed control and braking, which is commonly used on large wind turbines, some designs use flaps designed to open at a given speed, but in the authors’ view these are less direct and reliable than blade pitching. Electromagnetic control is commonly used, but in this case it was not possible as there was no generator, and if a turbine is driving a mechanical load, for example, a pump directly without an electrical system, electromagnetic control would not be possible.

Figure 12.

Angular velocity versus time for the two modes.

Comparison between predicted and measured performance

Figures 13 and 14 compare the predicted and measured C_P in fixed and variable pitch modes, and Figures 15 and 16 compare the predicted and measured C_Q in fixed and variable pitch modes. It will be seen that there are considerable discrepancies. In particular, the measured performance drops away more steeply than the simulations at high λ.

Figure 13.

Predicted and measured C_P in fixed pitch mode.

Figure 14.

Predicted and measured C_P in variable pitch mode.

Figure 15.

Predicted and measured C_Q in fixed pitch mode.

Figure 16.

Predicted and measured C_Q in variable pitch mode.

Discussion

The possible reasons for the discrepancies between the simulated and measured performance shown in Figures 13 to 16 include the following:

The pitch control mechanism may not have been working perfectly, so the experimental pitch regime may not have been quite sinusoidal.

The computational fluid dynamics (CFD) simulation made allowances for arm drag and additional drag generated by the junction of arms and blades, but these allowances were only estimates and may not have been as large as the actual drag losses. Figure 17 compares the CFD simulation allowance for these losses with the actual losses measured by Kirke (1998) by removing the blades from a 2-m-diameter × 1-m-high VAWT with two arms per blade, motoring up to speed and measuring the deceleration. The geometry of this turbine was similar but not identical to that of the turbine reported in this work, and arm tip losses in the absence of blades are not necessarily the same as arm–blade junction losses, so these curves are not strictly comparable. However, they illustrate the importance of accurately accounting for these losses.

Figure 17.

Parasitic drag loss due to arms and arm–blade junctions predicted by CFD, compared with measured arm drag after Kirke (1998).

Further evidence of the importance of arm drag is provided by Figures 18 to 20. Although all these refer to HKTs, the principles are the same as for wind turbines. Figure 18 after Marsh et al. (2013) shows a full 3D CFD simulation vortex visualization and illustrates the 3D effects. Figure 19 after Marsh et al. (2013) compares (1) CFD predictions with and without strut (arm) loss, (2) double multiple streamtube (DMS) predictions and (3) measured performance of a fixed pitch turbine. It will be apparent that all the predictions greatly over-predict C_P at λ > 2, especially those without arm drag corrections (Appendix 3).

Figure 18.

Full 3D CFD simulation vortex visualization after Marsh et al. (2013).

Figure 19.

Comparisons of C_P for (1) streamtube, (2) 2D CFD, (3) streamtube allowing for radial arm drag (S), (4) 3D CFD with ‘appendages’, that is, allowing for arm and junction drag (CFD-S) and (5) experimental results (EFD), after Marsh et al. (2013).

Figure 20.

Comparison of Ck (=C_P) versus TSR (=λ) for a turbine with four different radial arm arrangements, after Rawlings (2008).

Figure 20 after Rawlings (2008) compares C_P against TSR (λ) for the same turbine with four different radial arm arrangements under otherwise identical conditions. ‘Arm A’ comprised two 31% thick aerofoil arms attached to the blades at the quarter span points, and ‘Arm B’ comprised two 21% thick approximate aerofoil arms at quarter span points. ‘Arm C’ (arms at ends and middle, green curve) comprised three thin (NACA0012) aerofoil arms at blade ends and midspan and Arm C (ends only, purple curve) comprised two NACA0012 aerofoil arms at blade ends only. It will be apparent that both the arm section and location make a very large difference. For efficiency, it is clearly preferable to place arms at the ends of the blades only, but from a structural point of view this leads to much higher bending moments, stresses and deflections.

Conclusion

The performance of a straight blade VAWT has been modelled and measured experimentally in both fixed pitch and sinusoidal variable pitch modes. It is found that

Variable pitch improves starting and low-speed torque by a factor of about 3, allowing the variable pitch turbine to self-start, respond to gusts and accelerate up to operating speed much more quickly than a fixed pitch turbine, thus capturing more energy.

Variable pitch can provide simple and reliable overspeed control, unlike fixed pitch.

2D modelling ignoring parasitic drag losses over-predicts peak performance coefficient C_P and runaway TSRs.

Depending on the arrangement of the arms supporting the blades, parasitic drag losses due to arms and blade–arm junctions can be very large, and the reported experimental C_P of small models rarely exceed about 0.32.

Recommendations

It is recommended that the actual parasitic losses be assessed by measuring (1) the drag force on arm–blade junctions in a wind tunnel and (2) the angular deceleration of a complete fixed pitch turbine in still air.

Footnotes

Appendix 1

Appendix 2

Appendix 3 Acknowledgements

The authors would like to thank the following organizations and people: Tecalemit Aust P/L for the provision of material and workshop facilities; staff at Tecalemit who helped to build the turbine; Aussie Kanck, Jason Williams and Rob Kirke who helped to test it; the University of South Australia for the provision of a data logger; Rob Manson of I Want Energy, who financed the CFD work; Dr Leo Lazauskas whose DMS work provided the incentive for developing the variable pitch system and above all former General Manager at Tecalemit John Thomas, without whose enthusiasm and effort the turbine would not have been built.

Declaration of conflicting interests

The author(s) declared no potential conflicts of interest with respect to the research, authorship and/or publication of this article.

Funding

The author(s) received no financial support for the research, authorship and/or publication of this article.

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