GB2550349B

GB2550349B - Wind Turbine

Info

Publication number: GB2550349B
Application number: GB1608516.9A
Authority: GB
Inventors: Robert Murcott Benjamin
Original assignee: Individual
Current assignee: Individual
Priority date: 2016-05-16
Filing date: 2016-05-16
Publication date: 2019-06-12
Anticipated expiration: 2036-05-16
Also published as: GB201608516D0; GB2550349A

Description

WIND TURBINE

The present invention relates to a wind turbine with an automatic transmission.

There is a relationship between the efficiency of a wind turbine (in converting the power of the wind passing through the turbine into shaft power) and the ratio of rotational speed of the turbine and the wind speed. The ratio of rotational speed of the turbine and the wind speed is often expressed as a non-dimensional ratio, λ, of turbine tip speed and wind speed.

Different designs of rotor tend to be optimised for different ratios of tip speed and wind speed. For example, a Savonius type (vertical axis) rotor may produce maximum shaft power at relatively low ratio (e.g. λ-l) of rotor tip speed and wind speed, and a two bladed horizontal axis rotor may produce maximum shaft power at a relatively high ratio of rotor tip speed and wind speed (e.g. λ~10). The ratio of the wind power passing through the turbine with the shaft power of the rotor may be termed the power factor or power coefficient. The theoretical maximum power factor (according to Betz’s law) is 0.593.

Gearboxes have been used in existing wind turbines to match the optimum rotational speed of the rotor (with respect to the expected wind speed), with a more appropriate rotational speed for a generator to produce power. For utility scale (i.e. turbines with a rated output power of 1MW or greater), a fixed ratio gearbox is typically used. The use of a continuously variable transmission (CVT) has been proposed (Mangialardi, L., and G. Mantriota. "Dynamic behaviour of wind power systems equipped with automatically regulated continuously variable transmission." Renewable Energy 7.2 (1996): 185-203). A CVT equipped turbine is able to operate at a more ideal tip speed ratio Λ in a variable speed wind environment by varying the drive ratio to follow large fluctuations in wind speed. CVTs tend to be complex, and are typically less efficient than a conventional (e.g. fixed speed) gearbox.

An improved arrangement for improving performance of a wind turbine is desirable.

According to an aspect of the invention, there is provided a wind turbine in accordance with claim 1 or claim 2.

The centrifugal clutch may be configured to engage at a first threshold rotational rate, the first threshold rate being 20, 40, 60, or 100 rotations per minute or less.

The wind turbine may comprise a one-way bearing or overrunning clutch, configured to allow elements of the transmission to overrun each other in a first ratio, and to transmit power from the drive shaft to the driven shaft in a second ratio.

The centrifugal clutch may comprise; a first friction surface; a second friction surface that is moveable along an axis of rotation of the clutch to frictionally engage with the first friction surface; an eccentric mass, offset from the axis of rotation; a linkage, connected to the eccentric mass and the second friction surface and configured to convert radial inertial forces arising from rotation of the eccentric mass into axial forces for actuating the second friction surface.

The centrifugal clutch may further comprise a resilient element coupled to the linkage, arranged to urge the second friction surface away from the first friction surface.

The eccentric mass may be one of a plurality of eccentric masses, arranged in a dynamically balanced configuration about the axis of rotation. The linkage may be connected to the plurality of eccentric masses and configured to convert radial inertial forces arising from rotation of the plurality of eccentric masses into axial forces for actuating the second friction surface.

The driven shaft may be coaxial with the drive shaft. The transmission may further comprise an output one-way bearing or overrunning clutch coupled between the drive shaft and driven shaft, the output overrunning clutch or one-way bearing configured to allow the drive shaft to drive the driven shaft with a 1:1 transmission ratio, and to allow the driven shaft to overrun the drive shaft.

At least one of the first pair of transmission elements may be coupled to the drive shaft or further shaft by a one-way bearing or overrunning clutch.

The wind turbine may comprise a third transmission stage, the third transmission stage comprising: a third pair of coupled transmission elements, a first one of the third pair of coupled transmission elements being on the drive shaft and a second one of the third pair of coupled transmission elements being on the further shaft, a second clutch operable to engage and disengage the second transmission stage to change the transmission ratio.

The rotor may be configured to rotate about a vertical axis.

Embodiments of the invention will now be described, purely by way of example, with reference to the accompanying drawings, in which:

Figure 1 is a schematic diagram of a wind turbine with a horizontal axis rotor;

Figure 2 is a schematic diagram of a wind turbine with a vertical axis rotor;

Figure 3 is a schematic diagram of a two-stage, one-clutch, automatic gearbox;

Figure 4 is a schematic diagram of an alternative two-stage, one-clutch, automatic gearbox;

Figure 5 is a front view of an example clutch arrangement;

Figure 6 is a side view of the clutch arrangement of Figure 5;

Figure 7 is a diagram illustrating the operation of a linkage clutch like that shown in Figures 5 and 6;

Figure 8 is a schematic diagram of a three-stage, two-clutch automatic gearbox in a first state;

Figure 9 is a schematic diagram of the gearbox of Figure 8 in a second state

Figure 10 is a schematic diagram of the gearbox of Figures 8 and 9 in a third state;

Figure 11 is a view of a wind turbine according to an embodiment;

Figure 12 is a further view of the wind turbine of Figure 11;

Figure 13 is a view showing a clutch of the wind turbine of Figure 11;

Figure 14 is a graph of power co-efficient vs tip speed ratio for the rotor of the embodiment of Figure 11; and

Figure 15 is a graph comparing voltages generated by a wind turbine according to an embodiment, compared to a fixed drive ratio (but otherwise identical) wind turbine.

Figure 1 is a schematic view of a horizontal wind turbine 100 according to an embodiment, comprising a rotor 101, drive shaft 110, gearbox 150, generator 105, housing 103 and mast 102.

The rotor 101 may of any type, and is configured to convert the energy of the wind passing through the rotor 101 (in a direction substantially parallel to the rotational axis of the rotor 101) into rotational power by turning the rotor 101. The rotor 101 turns the drive shaft 110, which is connected to the gearbox 150. The gearbox 150 provides a multi-speed transmission between the rotor 101 and a driven shaft (not shown) of the generator 105, automatically selecting a transmission ratio based on the rotational speed of the rotor 101. As will be explained in more detail below, the gearbox 150 includes at least one centrifugal clutch which engages at a threshold rotation rate to alter the transmission ratio of the gearbox 150.

Altering the transmission ratio of the gearbox 150 may improve the efficiency with which the wind turbine 100 converts the energy in the wind passing through the rotor 101 into rotational power (for input to the generator 105). For example, when the wind is relatively low speed, the transmission of the gearbox 150 may be changed to reduce the load on the rotor 101, so that it rotates at a more optimal rate that improves the power factor of the wind turbine (by operating more close to the optimal tip speed to wind speed ratio).

Figure 2 shows an alternative wind turbine 100 according to an embodiment, in which the rotor 101 is a vertical axis rotor, configured to produce rotational power from wind passing through the rotor 101 in a direction substantially perpendicular to the rotational axis of the rotor 101. This embodiment may lack a mast 102, but may otherwise share the same elements as the wind turbine 100 of Figure 1. Similar features in Figure 2 are given the same reference numerals.

Figure 3 is a schematic view of a gearbox 150 for use in a wind turbine according to an embodiment. The gearbox 150 comprises a first (driven) shaft 120, second (drive) shaft 130, first transmission stage TS1 and second transmission stage TS2. The drive shaft 110 from a rotor 101 is coupled (e.g. directly) to the first shaft 120, and a generator 105 is coupled (e.g. directly) to the second drive shaft 130.

The first transmission stage TS1 comprises a first drive element AI coupled to the first shaft 120 by a first bearing 122, a second drive element A2 coupled to the second shaft 130 by a second bearing 132, a centrifugal clutch 180 and a first transfer element 141. The first transfer element 141 couples the first and second drive element AI, A2 so that their tangential velocities match.

The second transmission stage TS2 comprises a third drive element BI coupled to the first shaft 120 by a third bearing 124, a fourth drive element B2 coupled to the second shaft 130 by a fourth bearing 134, and a second transfer element 142. The second transfer element 142 couples the third and fourth drive element BI, B2 so that their tangential velocities match.

The first bearing 122 is a two-way bearing that allows the first drive element AI to rotate with respect to the first shaft 120, in either direction, unless the clutch 180 is engaged with the first drive element AI. When the clutch 180 is engaged with the first drive element AI, the first drive element AI is constrained to rotate with the first shaft 120, in both directions. The second bearing 132 fixes the second drive element A2 to rotate with the first shaft 120, in both directions. The third bearing 124 fixes the third drive element BI to rotate with the second shaft, in both directions. The fourth bearing 134 is a one-way bearing, which allows the second shaft 130 to overrun (i.e. rotate faster than) the fourth drive element B2, but does not allow the fourth drive element B2 to overrun the second shaft 130. The fourth drive element B2 can therefore drive the second shaft 130, but cannot be driven by it.

The one-way bearings (or overrunning clutches) used herein may be any suitable one way bearing mechanism, such as a sprag clutch or pawl based freewheel mechanism.

The drive elements AI, A2, BI, B2 may comprise any suitable power transmission elements, such as drive sheaves, sprockets, or gears. The transfer elements 141, 142 may comprise chains or belts (or a combination of chains and belts). Where the drive elements comprise gears, the transfer elements 141, 142 may be dispensed with (the drive elements may directly engage each other).

When the clutch 180 is not engaged the transmission ratio (rotation speed at generator 105 / rotation speed at drive shaft 110) of the gearbox 150 is that of the second transmission stage TS2, which is defined by the ratio of radii of the third and fourth drive elements rBl/rB2. For example, rBl/rB2 in the present example may be less than 1, defining a relatively low transmission ratio. The one-way fourth bearing 134 means that the fourth drive element B2 drives the second shaft 130.

When the clutch 180 is engaged, the transmission ratio of the gearbox 150 is that of the first transmission stage TS1, which is defined by the ratio of radii of the first and second drive elements rAl/rA2. For example, rAl/rA2 in the present example may be greater than 1, defining a relatively high transmission ratio. The second shaft 130 overruns the fourth drive element B2 when the clutch 180 is engaged.

The clutch 180 is a centrifugal clutch, and is configured to engage when the first shaft 120 rotates faster than a threshold speed (and may disengage when the first shaft rotates slower than the threshold speed). As the rotor 101 turns faster, the gearbox 150 therefore responds by increasing the transmission ratio.

There may be momentary slipping between the friction surfaces before the surfaces lock/engage to transfer torque. The margin of rotational speed over which slipping occurs could be reduced or eliminated by including a bistable mechanism, for instance employing buckling. The bistable mechanism could be connected to the linkage 184.

In the example of Figure 3, the clutch 180 is adjacent to the first drive element along the first shaft 120. In other embodiments, the clutch 180 may be concentric with the first drive element Ai and/or the bearing 122 by which the first drive element Ai is mounted on the first drive shaft 120. Figure 4 illustrates such an example arrangement. The gearbox 150 of Figure 4 is otherwise the same as that of Figure 3.

In general, a centrifugal clutch is any device in which an apparent centrifugal force resulting from rotation of an eccentric mass is used to actuate a friction plate. The centrifugal clutch 180 may comprise any suitable arrangement.

Figures 5 and 6 illustrate an example centrifugal clutch 180, comprising two eccentric masses 183, linkage mechanism 184, clutch drive shaft 186, and friction plates 181, 182. The eccentric masses 183 are offset from the rotational axis of the drive shaft 130, and are attached to the linkage mechanism 184. The linkage mechanism 184 comprises a plurality of substantially rigid links, connected together at pivots 185, and is arranged to convert the apparent centrifugal force resulting from rotation of the eccentric masses 183 about the rotational axis of the shaft 130 into an axial thrust force urging the friction plate 182 in the direction of an adjacent friction plate 181, mounted on an adjacent drive element 131. A resilient element (or spring) 187a, 187b, not shown in Figure 6, is preferably used to bias the friction plate 181 away from the adjacent friction plate attached to the drive element 131. The threshold rotation speed, at which the clutch friction plates 182, 181 engage, is set by a combination of the design of the mechanical linkage 184, the mass and offset (from the rotational axis) of the eccentric masses 183 and the spring constant of the resilient element (if present). In this example a compression spring 187a is used with a tension spring 187b. These springs both act to resist engagement of the clutch. The compression spring 187a may be stiffer than the tension spring 187b.

Figure 7 illustrates the relationship between the forces acting within the mechanical linkage 184 for an example design. The linkage is substantially symmetric about the shaft axis, and on each side comprises a first bar 184a and second bar 184b. The first bar 184a is pivotally connected at a first end to the clutch drive shaft 186, and at the second end to the second bar 184b. The second bar 184b is pivotally connected at a first end to the first bar 184a, and at a second end to a resilient member (or spring) 187. The resilient member 187 is connected at either end to the second end of the second bar 184 of each respective side of the linkage 184. The resilient member 187 is configured to resist movement of the linkage 184, and may be arranged to pre-load the linkage (for example, to urge friction plate 182 away from the other friction plate 181). The second bar 184b is further pivotally connected to a constraint that prevents movement transverse to the shaft axis. The constraint is illustrated schematically in Figure 7, but may, for instance, take the form of a bar connecting the second bar 184b of each side of the linkage together.

The threshold rotation rate at which the friction plate 182 engages with the friction plate of the corresponding drive element is determined by a number of factors. The apparent centripetal force Cf resulting from rotation at an angular velocity, ω of each eccentric mass m with centre of mass at offset r is given by:

Cf = m.r. ω2

The resulting axial thrust force F„ on the friction plate 182, as a result of the force from two such eccentric masses is given by:

Fn = 2-Cftan(0)

The level of torque Tc that can be transmitted between the engaged friction plates 182 is a function of the axial thrust force F„, the coefficient of friction μ between the clutch plates, and the outer and inner radius of the clutch plate (r0, r,): <3 3% 2 ro - ri T = -μ-F · - e 3 11 2 2 r — r· <0 h /

The force as a result of the resilient member or spring Fs is given by:

Fs = k-> where k is the spring constant and x is the change in length of the resilient member or spring.

The geometry of the linkage, the mass and offset of the eccentric mass, and the spring constant can all be varied to change the threshold at which the clutch 180 engages.

Figures 8 to 10 show an alternative example gearbox 150, having three transmission stages (A, B C). The gearbox comprises a first (jack) shaft 120, second (drive) shaft 130, first drive element Al, second drive element A2, third drive element Bi, fourth drive element B2, fifth drive element Cl, sixth drive element C2, first clutch 180a, and second clutch 180b.

The first, third and fifth drive elements Al, Bi, Cl, are respectively mounted to the first shaft 120 by first, third and fifth bearings 122, 124, 126. The second, fourth and sixth drive elements A2, B2, C2 are respectively mounted to the second shaft 130 by second, fourth and sixth bearings 132, 134, 136. The second drive element A2 is coupled directly to the driven shaft (not shown) of a generator 105. The first drive element Al has a larger radius than the second drive element A2.

The fourth and sixth bearings 134, 136 are two way bearings, so that the fourth and sixth drive elements B2, C2 can rotate in either direction relative to the second drive shaft 130, unless they are engaged by first or second clutch 180a, 180b.

The third and fifth bearing 124, 126 are each one-way bearings, which allow the third and fifth drive elements Bi, Cl to drive the first shaft 120, but which also allow the first shaft 120 to overrun the drive elements B2, Cl. The second bearing 132 is a oneway bearing, which allows the second drive element A2 to overrun the second drive shaft 130, but which allows the second drive shaft 130 to drive the second drive element A2. The first bearing 122 is a fixed bearing, which prevents relative rotation between the first drive element Ai and the first drive shaft 120. A first transfer element 141 (e.g. chain/belt) couples the first drive element Ai and second drive element A2 to rotate at the same tangential velocity. A second transfer element 142 (e.g. chain) couples the third drive element Bi and fourth drive element B2 to rotate at the same tangential velocity. A third transfer element 143 (e.g. chain) couples the fifth drive element Cl and sixth drive element C2 to rotate at the same tangential velocity.

The rotor drive shaft (not shown) may be coupled (e.g. directly) to a first end of the second (drive) shaft 130 with the generator 105 positioned at the second end. The driven shaft of the generator 105 is not directly coupled to the second shaft 130, but is instead coupled directly to the second drive element A2.

Figure 8 illustrates the operation of the gearbox 150 when the second drive shaft 130 (and rotor) is rotating relatively slowly, and both the first and second clutches 180a, 180b are disengaged. The second drive shaft 130 is coupled to the generator via the one-way second bearing 132 and the second drive element A2. The drive ratio (ratioj) is therefore 1:1, with the driven shaft of the generator 105 being effectively coupled directly to the rotor via an overrunning bearing that allows the generator to overrun the rotor. Since the first and second clutches 180a, 180b are disengaged, the fourth and sixth drive elements B2, C2 do not transmit torque, and are free to rotate relative to the second drive shaft 130.

Figure 9 illustrates the operation of the gearbox 150 when the second drive shaft 130 is rotating above the first threshold speed which causes engagement of the first clutch 180a, but below the second threshold speed that causes engagement of the second clutch 180b. The fourth drive element B2 is engaged with the second shaft 130 by the first clutch 180a, with the result that the first shaft 120 is driven by the second transfer element 142 and third drive element BI, via the one-way bearing 124. The first shaft 120 in turn drives the generator 105 via the first drive element AI, first transfer element 141 and second drive element A2 (which is coupled directly to the driven shaft of the generator 105). The second drive element A2 overruns the second drive shaft 130 on the second one-way bearing 132. The drive ratio in this configuration (ratio 2) is defined by the relative radiuses rAl, rA2, rBl, rB2 of the first to fourth drive elements respectively:

In this example, ratio2 > 1, because rB2/rBl = l, and rAl>rA2. More specifically rBl=rB2, and rAl=2.rA2, so the ratio2 in this example is 2.

Figure 10 illustrates the operation of the gearbox 150 when the second (driven) shaft 130 is rotating above the second threshold speed, which leads to engagement of the second clutch 180b (the first clutch 180a may remain engaged). The sixth drive element C2 is thereby constrained to rotate with the second drive shaft 130. The fifth drive element Cl is driven via the third transfer element 143, and the first shaft 120 is driven by the fifth drive element Cl via the one-way fifth bearing 126. The third drive element B2 remains engaged with the second shaft 130 via the first clutch 180a, and the third drive element BI therefore continues to be driven. However, since rC2/rCl > rB2/rBl, the fifth drive element Cl will rotate faster than the third drive element BI, so the first shaft 120 will overrun the third drive element BI. As in Figure 9, the drive is transferred from the first shaft 120 to the driven shaft of the generator 105 via the first and second drive elements AI, A2. The drive ratio in this configuration (ratio 3) is given by:

In this example, ratio3 > ratio2 because rC2/rCl > rB2/rBl. More specifically, rC2=2.rCl, so ratio3 in this example is 4.

The number of transmission stages and ratios may be arbitrarily large. Further drive stages (e.g. 1, 2, 3, 4 or more) may be added to produce a gearbox with a larger number of drive ratios. The number of drive shafts may be two, with each additional stage being added along the length of the shafts, or further drive shafts may be added to include additional drive stages.

In some embodiments, at least one electronically controlled clutch may be included. Although a centrifugal clutch is a low cost and reliable actuation method, electronic clutches may increase the flexibility of the gearbox, since they can be programmed to engage and disengage under arbitrary conditions, allowing for more complicated modes of operation.

Figures 11 to 13 show views of an example wind turbine 100. The wind turbine 100 comprises three bladed vertical axis rotor 101, frame 200 and gearbox 150. The gearbox 150 in this example is similar in operation to that shown in Figures 8 to 10, but in other embodiments an arrangement more similar to Figures 3 or 4 may be used, or features may be combined from either embodiment.

Figure 14 illustrates the power coefficient vs tip speed ratio for the rotor 101 shown in Figure 11. The maximum power coefficient is achieved when the tip speed ratio is around 0.8. A significant drop in the power ratio occurs when the tip speed ratio is not near this operating speed. A wind turbine with a gearbox according to an embodiment can maintain operation in a regime that is closer to the optimal tip speed ratio than a conventional turbine.

Figure 15 illustrates the real benefits that are obtainable with a turbine according to an embodiment. A turbine similar to the embodiment of Figure 11 was operated with and without the automatic gearbox during variable wind conditions, and the voltage at the generator monitored. In some test runs, the generator was given an initial speed of rotation, sufficient to generate a predetermined voltage.

Figure 15 shows: 401 a first run using the automatic gearbox, 402 a second run using the automatic gearbox, 403 a run without the automatic gearbox, with the generator starting from rest, 404 a run without the automatic gearbox, with the generator starting from a speed that generates 2V, 405 a run without the automatic gearbox, with the generator starting from a speed that generates IV, 406 a run without the automatic gearbox, with the generator starting from a speed that generates 2V, 407 a run without the automatic gearbox, with the generator starting from a speed that generates 2.25V.

The best results are obtained from run 402, in which a gearbox according to the invention was used. Although this test is rather anecdotal, it demonstrates the potential improvements in power that can be obtained.

The foregoing are merely examples that are not intended to limit the scope of the invention, which is limited only by the appended claims.

Claims

1. A wind turbine, comprising: a rotor, for driving in rotation by wind; a generator; and a multi-speed transmission between the rotor and the generator, the transmission comprising: a drive shaft on the rotor side of the transmission; a driven shaft on the generator side of the transmission; a centrifugal clutch, operable to automatically alter a transmission ratio between the drive shaft and driven shaft in response to the speed of the drive shaft; a first transmission stage comprising a first pair of coupled transmission elements, a first one of the first pair of coupled transmission elements being on the drive shaft and a second one of the first pair of coupled transmission elements being on the driven shaft, wherein the centrifugal clutch is operable to engage and disengage the first transmission stage to change the transmission ratio; and a second transmission stage comprising a second pair of coupled transmission elements, one of the second pair of coupled transmission elements on the drive shaft and the other on the driven shaft.

2. A wind turbine, comprising: a rotor, for driving in rotation by wind; a generator; and a multi-speed transmission between the rotor and the generator, the transmission comprising: a drive shaft on the rotor side of the transmission; a driven shaft on the generator side of the transmission; a centrifugal clutch, operable to automatically alter a transmission ratio between the drive shaft and driven shaft in response to the speed of the drive shaft; a further shaft; a first transmission stage comprising a first pair of coupled transmission elements, a first one of the first pair of coupled transmission elements being on the drive shaft and a second one of the first pair of coupled transmission elements being on the further shaft, wherein the centrifugal clutch is operable to engage and disengage the first transmission stage to change the transmission ratio; and a second transmission stage comprising a second pair of coupled transmission elements, one of the second pair of coupled transmission elements on the further shaft and the other on the driven shaft.

3. The wind turbine of claim 1 or 2, wherein the centrifugal clutch is configured to engage at a first threshold rotational rate, the first threshold rate being 100 rotations per minute or less.

4. The wind turbine of any preceding claim, further comprising an overrunning clutch, configured to allow elements of the transmission to overrun each other in a first ratio, and to transmit power from the drive shaft to the driven shaft in a second ratio.

5. The wind turbine of any preceding claim, wherein the centrifugal clutch comprises; a first friction surface; a second friction surface that is moveable along an axis of rotation of the clutch to frictionally engage with the first friction surface; an eccentric mass, offset from the axis of rotation; a linkage, connected to the eccentric mass and the second friction surface and configured to convert radial inertial forces arising from rotation of the eccentric mass into axial forces for actuating the second friction surface.

6. The wind turbine of claim 5, wherein the centrifugal clutch further comprises a resilient element coupled to the linkage, arranged to urge the second friction surface away from the first friction surface.

7. The wind turbine of claim 5 or 6, wherein the eccentric mass is one of a plurality of eccentric masses, arranged in a dynamically balanced configuration about the axis of rotation, and the linkage is connected to the plurality of eccentric masses and configured to convert radial inertial forces arising from rotation of the plurality of eccentric masses into axial forces for actuating the second friction surface.

8. The wind turbine of any preceding claim that depends from claim 2, wherein the driven shaft is coaxial with the drive shaft; the transmission further comprising an output overrunning clutch coupled between the drive shaft and driven shaft, the output overrunning clutch configured to allow the drive shaft to drive the driven shaft with a 1:1 transmission ratio, and to allow the driven shaft to overrun the drive shaft.

9. The wind turbine of any preceding claim, wherein at least one of the first pair of transmission elements is coupled to the drive shaft or further shaft by a one-way bearing.

10. The wind turbine of any preceding claim that depends from claim 2, further comprising a third transmission stage, the third transmission stage comprising: a third pair of coupled transmission elements, a first one of the third pair of coupled transmission elements being on the drive shaft and a second one of the third pair of coupled transmission elements being on the further shaft, a second clutch operable to engage and disengage the second transmission stage to change the transmission ratio.

11. The wind turbine of claim 10, wherein at least one of the third pair of transmission elements is coupled to the drive shaft or further shaft by a one-way bearing.

12. The wind turbine of any preceding claim, wherein the rotor is configured to rotate about a vertical axis.