NZ626075B2

NZ626075B2 - Power generating system and hydraulic control system

Info

Publication number: NZ626075B2
Application number: NZ626075A
Authority: NZ
Inventors: Geoffrey Morgan Henderson; Peter John Scott; Matthew James Silvester; Richard James Trudgian
Original assignee: Windflow Technology Limited
Priority date: 2011-12-20
Filing date: 2012-12-11
Publication date: 2016-03-01

Abstract

Disclosed is a hydraulic control system for controlling a variable ratio transmission of a power generating system. The power generating system includes a rotor or turbine (101) adapted to be driven by a fluid at variable speed and an electrical generator (103) adapted to be driven at a constant or variable speed, the transmission coupling the rotor or turbine (101) to the electrical generator (103). The hydraulic control system has a hydraulic motor/pump unit (104) adapted to be operably connected to a superposition gear. The hydraulic motor/pump unit (104) is connected to a hydraulic circuit with an orifice and/or a relief valve configured to open at a predetermined hydraulic pressure. The hydraulic circuit switches between a variable low-speed operating mode and a torque limiting high-speed operating mode. In, the torque limiting high-speed operating mode, the hydraulic motor/pump unit (104) is adapted to be driven by the superposition gear and operates as a pump to drive hydraulic fluid through the orifice and/or relief valve to provide a passive torque limiting function. In the variable low-speed operating mode the hydraulic motor/pump unit (104) is adapted to be driven as a motor to drive the superposition gear. The hydraulic control system is configured to provide a desired rotor speed by controlling hydraulic fluid flow rate through the hydraulic motor/pump unit (104). variable speed, the transmission coupling the rotor or turbine (101) to the electrical generator (103). The hydraulic control system has a hydraulic motor/pump unit (104) adapted to be operably connected to a superposition gear. The hydraulic motor/pump unit (104) is connected to a hydraulic circuit with an orifice and/or a relief valve configured to open at a predetermined hydraulic pressure. The hydraulic circuit switches between a variable low-speed operating mode and a torque limiting high-speed operating mode. In, the torque limiting high-speed operating mode, the hydraulic motor/pump unit (104) is adapted to be driven by the superposition gear and operates as a pump to drive hydraulic fluid through the orifice and/or relief valve to provide a passive torque limiting function. In the variable low-speed operating mode the hydraulic motor/pump unit (104) is adapted to be driven as a motor to drive the superposition gear. The hydraulic control system is configured to provide a desired rotor speed by controlling hydraulic fluid flow rate through the hydraulic motor/pump unit (104).

Description

POWER GENERATING SYSTEM AND HYDRAULIC CONTROL SYSTEM FIELD OF THE INVENTION This invention s to a power generating system of the type in which a fluid driven turbine or rotor drives an alternating current tor. The invention is more specifically concerned with such a system wherein the rotational torque of the turbine or rotor may vary due to changes in the speed of the fluid driving the turbine. An example of such a power generating system is a wind-driven turbine or ll. The invention also relates to a hydraulic control system for use in a transmission of such a power generating system.

BACKGROUND Variable rotor speed in wind turbines is desirable for several reasons. For a given rotor, the optimum rotor speed for maximum aerodynamic efficiency is a function of the wind speed. lly, the optimum rotor speed increases with increasing wind speed. Therefore, le rotor speed in wind turbines is desirable to maximise energy capture. Variable rotor speed also enables power generation at lower wind speeds (which correspond to lower rotor speeds), sing the range of wind conditions that a wind turbine can operate in and thereby increasing the annual number of ‘run-hours’. Operation at lower speeds also has noise advantages because the rotor sound levels reduce when the turbine runs at lower rpm. Finally, variable rotor speed enables the torque on the ission to be limited and smoothed at the turbine’s rated power, thus reducing, to a greater or lesser extent depending on where in the transmission and how the torque is limited, the torque duty of the gearbox and transmission in general.

Since the 1990s, most commercial wind turbines have adopted some means to enable variable speed of the wind turbine rotor. Such means may be electrical or mechanical ‘variable speed’ systems. Wind turbines prior to the late 1990s typically used a conventional asynchronous ion generator with a fixed-ratio gearbox which allowed only small rotor speed variations.

Electrical le speed systems lly enable the generator speed to vary significantly and thereby allow variable rotor speed with a fixed ratio transmission. Power electronic rectifiers and inverters are generally necessary to enable the wind e generator (which may be of synchronous or asynchronous design but in either case is not synchronised with the grid in the manner of a tional synchronous generator) to be electrically connected to an alternating current electrical grid of constant frequency. This is a high cost approach as power electronic iers and inverters at the rating of utility sized commercial wind turbines are expensive.

Wind turbines having electrical variable speed s also undergo greater drive-train torque fluctuations than those having mechanical systems because the inertia of the generator rotor needs to be accelerated during gusts, giving rise to non-trivial fluctuations in the drive-train torque even though the electrical system may control the ic torque at the rotor-stator interface very effectively. Excessive drive-train torque fluctuations can be ng to the wind turbine transmission and this damage mode is a continuing problem in the wind energy ry.

Mechanical variable speed systems enable the use of a directly grid connected synchronous generator, which runs at constant speed set by and synchronised with the grid. The use of a directly grid connected synchronous generator has cost advantages over electrical variable speed systems because a low cost ‘utility grade’ onous generator can be sourced and there is no need for expensive power onic rectifiers and inverters. However, a directly grid connected synchronous generator must operate at a constant speed dictated by the electrical grid frequency and the number of poles of the synchronous generator. Therefore, either the wind turbine rotor speed must be constant for a fixed ratio transmission, or a variable ratio transmission must be employed to enable variable rotor speed to se the energy capture. Rotors in wind turbines undergo turbulence-induced torque fluctuations. Since grid-connected synchronous tors have no compliance to absorb such torque fluctuations, this means that constant rotor speed is not a practical option for wind turbines and the ations must be smoothed by some other form of compliance (provided by a mechanical variable speed system) to prevent damage to the wind turbine transmission.

Mechanical variable speed systems can manage turbulence-induced torque fluctuations in wind turbines and similar power generation sources such as tidal stream generators.

Most attempts to enable variable rotor speed in wind turbines with nt speed synchronous generators have been variations of the invention described in WO 81/01444. In that document, a variable ratio transmission is achieved using one or more epicyclic gear stages connecting the main mechanical power transmission path n the wind turbine rotor and generator, to a , parallel bypass power transmission path. The bypass power transmission path may be hydraulic or electrical with two hydraulic or electrical machines, one or both of which may be variable speed, and both of which may be variable power. The two hydraulic or electrical machines are interconnected by a means of transmitting power which may be a closed loop lic or electrical circuit, as appropriate. At low rotor speed bypass power is itted from the generator side to the rotor side of the variable ratio transmission to increase the transmission ratio. At high rotor speed bypass power is transmitted from the rotor side to the generator side of the variable ratio transmission to decrease the transmission ratio. This type of variable ratio transmission adds significant complexity and cost to the wind turbine train.

For a typical commercial wind turbine having a variable ratio transmission as described above, the maximum bypass power required to enable le rotor speed mode in all wind speeds up to the rated wind speed, may be 20-50% of the rated wind turbine power, depending on the variable speed range and the type of transmission. Figure 1 shows a graph of rotor efficiency against tip speed ratio (TSR), which is the ratio of the tangential speed of the tip of the rotor blade to the actual velocity of the wind. The peak of this efficiency-TSR curve for a typical rotor is relatively broad and flat, which means the incremental gains in energy capture diminish to the point of being negligible as the TSR is moved closer to the ‘optimum TSR’, where the ‘optimum TSR’ is that which gives maximum rotor efficiency (at the maximum in Figure 1), although the cost of the bypass power system increases the wider the speed and power range over which it is active. Therefore, the benefit to cost also diminishes the wider the speed and power range over which one attempts to in the TSR at optimum and a practical limit is determined depending on the cost of the particular bypass power system utilised.

The above type of variable ratio ission may also limit excessive drive-train torque fluctuations by means of an active controller as described in and WO 49109. These active controllers change the transmission ratio in a complementary way, ideally at the same rate as the ent change in turbine speed, to maintain a constant generator speed and torque while allowing the wind turbine rotor to rate and absorb the transient excess kinetic energy. This method of limiting excessive drive-train torque fluctuations may not always be effective due to the response time of the active controller.

US patent 5140170 to Geoffrey M. Henderson describes a wind turbine transmission in which ng drive-train torque fluctuations are substantially eliminated using a variable ratio transmission with a passive hydraulic torque limiting function, and an active blade pitch l system which allows small ions in wind turbine rotor speed. In the system described in that patent, a grid connected synchronous generator is driven at a constant speed and the wind turbine rotor speed is near constant until the design rated rotor torque level is exceeded due to transient aerodynamic torque fluctuations. When the design rated rotor torque is exceeded, the transmission ratio decreases rapidly due to e lic slip, the transient excess energy being stored as kinetic energy as the wind turbine rotor accelerates, and also being dissipated as heat by the hydraulic .

The active blade pitch control in US 5140170 ts rotor over-speed at high wind speeds when the aerodynamic rotor power available is sufficient to operate at or above the design rated rotor torque level. This torque ng system is simple and cost effective, however the narrow band of variable rotor speed achieved (typically with up to 5% rotor speed variation above the minimum rotor speed) enables the torque limiting system but does not se energy capture other than by recovering kinetic energy stored in the rotor during transient fluctuations in wind speed at the rated power level. Such a system is generally ured so that the optimum TSR, and hence peak aerodynamic rotor efficiency, is achieved at a n wind speed. At lower wind speeds the rotor speed is faster than the ideal rotor speed for the optimum TSR and at higher wind speeds the rotor speed is slower than the ideal rotor speed for the optimum TSR. Having a faster than m rotor speed at low wind speeds also means that the wind turbine will have a relatively high cut-in wind speed and relatively high turbine sound levels in light winds.

In this specification where reference has been made to patent specifications, other external documents, or other sources of information, this is generally for the purpose of providing a context for discussing the es of the invention. Unless specifically stated otherwise, reference to such external documents or such sources of information is not to be construed as an admission that such documents or such sources of information, in any jurisdiction, are prior art or form part of the common l knowledge in the art.

It is an object of at least preferred embodiments of the t invention to provide a simple and cost effective control system for a torque limiting variable ratio transmission to provide variable rotor speed at low rotor speeds, to enable a wind turbine with a grid connected synchronous tor to operate with continuously varying rotor speed, and thereby to reduce cut-in wind speed, reduce turbine sound levels in light winds, and increase energy capture across a broader wind speed range compared to a near constant rotor speed wind turbine, and/or to provide the public with a useful alternative to electrical variable speed systems.

SUMMARY OF THE INVENTION In a first aspect, the invention may be said to broadly consist in a hydraulic control system for controlling a variable ratio transmission of a power generating system comprising a rotor or turbine adapted to be driven by a fluid at variable speed and an electrical generator adapted to be driven at a constant or variable speed, the transmission coupling the rotor or turbine to the electrical generator, the hydraulic control system comprising: a hydraulic motor/pump unit adapted to be operably connected to a osition gear; said hydraulic motor/pump unit connected to a hydraulic circuit that comprises an orifice and/or a relief valve configured to open at a predetermined hydraulic pressure, the hydraulic t configured to switch between a variable low-speed ing mode and a torque limiting high-speed operating mode, wherein in the torque limiting high-speed operating mode the hydraulic pump unit is adapted to be driven by the superposition gear and operates as a pump to drive hydraulic fluid through the orifice and/or relief valve to provide a passive torque limiting function, and wherein in the variable low-speed operating mode the hydraulic motor/pump unit is adapted to be driven as a motor to drive the superposition gear and the hydraulic control system is configured to provide a desired rotor speed by controlling hydraulic fluid flow rate through the hydraulic motor/pump unit.

In an embodiment of the t ion, the hydraulic motor/pump unit is of a ve cement type. Alternatively, the hydraulic motor/pump unit may be a dynamic type.

In an embodiment of the present invention, the hydraulic circuit is configured to switch n the variable low-speed operating mode and the torque limiting high-speed operating mode in response to measured hydraulic pressure in the hydraulic circuit or otherwise in response to measured torque on the transmission.

In an embodiment of the present invention, the hydraulic fluid flow rate through the hydraulic motor/pump unit is controlled in the variable low-speed operating mode in se to measured hydraulic pressure in the hydraulic t or otherwise in response to measured torque on the ission.

In an embodiment of the present invention, the hydraulic control system further comprises a variable displacement supply pump, wherein in the variable eed operating mode the le displacement supply pump is configured to supply fluid to drive the hydraulic motor/pump unit as a motor. The variable displacement supply pump may be ically driven. Alternatively, the variable displacement supply pump may be ured to be directly driven by the transmission.

In an embodiment of the present invention, the hydraulic control system r comprises a pressure measuring device arranged to measure the pressure in a high pressure region of the hydraulic circuit. The re measuring device may comprise a transducer.

In an embodiment of the present invention, the hydraulic circuit defines a first fluid path and a second fluid path which respectively define fluid flow in the le low-speed operating mode and the torque limiting peed operating mode.

Preferably, the first and second fluid paths are closed loop paths.

In an embodiment of the present invention, the lic control system comprises a controller configured to switch the hydraulic circuit between the first and second fluid paths, at a predetermined hydraulic pressure.

In an embodiment of the present ion, the hydraulic control system further ses at least one solenoid valve for switching between the first and second fluid paths.

In a second aspect, the invention may be said to broadly consist in a variable ratio transmission for a power generating system, the transmission comprising: a transmission input adapted to be driven at a variable speed; a transmission output adapted to be driven at a nt or variable speed; and a hydraulic control system as outlined in relation to the first aspect above.

In an embodiment of the present invention, the transmission comprises a planetary gear arrangement between ission input and the transmission output, the planetary gear arrangement comprising a ary gear input operably connected to the transmission input, a planetary gear output operably connected to the transmission output, and the osition gear.

In a third aspect, the invention may be said to broadly consist in a power generating system comprising: a rotor or turbine d to be driven at a variable speed; a generator adapted to be driven at a constant or variable speed; and a variable ratio transmission as outlined in relation to the second aspect above that couples the rotor or turbine and the generator.

In an ment of the present invention, the power generating system has a rated power, and the power output from the power generating system when in the variable low-speed operating mode is limited to be less than or equal to about 25% of the rated power.

In an embodiment of the present ion, the rotor or turbine has a zero slip speed when the angular velocity of the superposition gear is zero, and the minimum rotor speed in the variable low-speed operating mode is greater than or equal to about 75% of the zero slip speed.

In an embodiment of the present invention, the power generating system has a rated power, and the maximum power required by the hydraulic control system ents to achieve the variable low-speed operating mode is less than or equal to about 5% of the power generating system rated power.

In an embodiment of the present invention, the generator is a synchronous generator operating at a constant speed set by a larger generating system or grid. Alternatively, the generator may be an asynchronous generator operating at variable speed in combination with an electrical variable speed system.

The following summarises an ment of the present invention, which is described in further detail in the ‘detailed description’ section. The transmission assembly es a mechanical train connecting the rotor to the tor, and a superposition gear either driving or being driven by a positive displacement hydraulic pump unit (which preferably has a linear relationship between fluid pressure and output torque). A hydraulic circuit is provided which is switchable n two closed loop fluid paths that correspond with two distinct operating modes. The hydraulic t and positive displacement motor/pump unit together make up the hydraulic control system.

In the first operating mode (variable low-speed operating mode), the positive displacement hydraulic motor/pump unit es as a motor and is driven by fluid flow ed by an electrically driven variable displacement hydraulic pump, which enables the rotor to operate at a variable low speed.

In the second operating mode (torque limiting high-speed operating mode), the positive displacement hydraulic motor/pump unit operates as a pump and drives fluid through an orifice resulting in substantially zero slip in the transmission, until the rotor torque s a predetermined level at which a relief valve opens (passive operation) and allows increased fluid flow-rate through the positive displacement hydraulic motor/pump unit and hence the transmission assembly is allowed to slip so as to permit the rotor speed to increase in such a manner as to in a substantially constant torque at the generator.

The variable ratio transmission s a variable rotor speed that increases rotor power and hence energy capture at low wind speeds, specifically at wind speeds below the point at which the rotor reaches the zero slip rotor speed which is the rotor speed at which the positive displacement lic motor/pump unit is stationary. At wind speeds above this point the variable ratio transmission is ily intended as a passive torque limiting transmission and at high wind speeds the rotor speed is less than the optimum rotor speed for maximum energy capture.

An active hydraulic control loop is provided to control the transmission in se to the torque derived by ement of the hydraulic pressure. The control loop is configured to provide an output for switching n the variable low-speed and torque limiting high-speed operating modes, and also to control the hydraulic fluid flow rate when operating in the variable low-speed operating mode which sets the speed of the hydraulic motor/pump unit, and therefore sets the rotor speed when in the variable low-speed operating mode.

The term ising" as used in this specification and claims means "consisting at least in part of". When reting statements in this specification and claims which include the term "comprising", other features besides the features prefaced by this term in each statement can also be present. Related terms such as "comprise" and "comprised" are to be interpreted in a similar manner.

It is intended that reference to a range of numbers sed herein (for example, 1 to 10) also orates reference to all rational numbers within that range (for example, 1, 1.1, 2, 3, 3.9, 4, , 6, 6.5, 7, 8, 9 and 10) and also any range of rational numbers within that range (for example, 2 to 8, 1.5 to 5.5 and 3.1 to 4.7) and, therefore, all sub-ranges of all ranges expressly disclosed herein are hereby expressly disclosed. These are only examples of what is specifically intended and all possible combinations of numerical values between the lowest value and the highest value enumerated are to be considered to be sly stated in this application in a similar manner.

As used herein the term “(s)” following a noun means the plural and/or singular form of that noun.

As used herein the term “and/or” means “and” or “or”, or where the context allows both.

The invention ts in the foregoing and also envisages constructions of which the following gives examples only.

BRIEF DESCRIPTION OF THE DRAWINGS The present invention will now be described by way of example only and with nce to the accompanying drawings in which: Figure 1 is a graph of rotor aerodynamic efficiency versus tip speed ratio (TSR) for a known wind turbine rotor; Figure 2 is a schematic diagram of a power ting system comprising a hydraulic control system HCS and transmission according to one embodiment of the present invention; Figure 3 is a schematic diagram of the hydraulic circuit for the hydraulic control system for a transmission according to one ment of the present invention; Figure 4 is a graph of rotor and hydraulic unit power versus wind speed for one embodiment of the present invention; Figure 5 is a graph of rotor and hydraulic unit speeds versus wind speed for one embodiment of the present invention; Figure 6 is a graph of rotor and hydraulic unit torques versus wind speed for one embodiment of the t invention; Figure 7 is a graph of rotor and hydraulic unit s versus rotor speed for one embodiment of the present invention; Figure 8 is a graph of hydraulic unit power, torque and speed versus wind speed for one embodiment of the present invention; Figure 9 is a graph of hydraulic unit flow rate and wind speed versus hydraulic unit differential pressure for one embodiment of the present invention; Figure 10 is a simplified l logic diagram for a hydraulic control system for a transmission according to one embodiment of the present ion; Figure 11 is a schematic diagram of the hydraulic control system for a transmission according to a second ment of the present invention; Figure 12 is a schematic diagram of the hydraulic control system for a transmission according to a third ment of the present invention; and Figure 13 is a graph showing detail D13 of a part of the graph shown in Figure 4, provided for clarity.

DETAILED DESCRIPTION OF A RED EMBODIMENT Referring to Figure 2, in a power generating system such as a wind turbine, the shaft 1 of a wind driven rotor 101 provides the input to a planet carrier of a fixed ratio planetary gear stage PGS of a wind turbine transmission, which includes a plurality of planet gears 5, a stationary s gear 6 mounted to the transmission g 16, and a sun gear 7. The sun gear 7 provides an input to a fixed ratio parallel gear stage FRPGS of the wind turbine transmission, which includes a wheel gear 8 and a pinion gear 9 mounted on an output shaft 2. The fixed ratio planetary gear stage PGS and the fixed ratio parallel gear stage FRPGS together form a two-stage fixed ratio transmission for a wind turbine, which may provide a step-up ratio, for example, of 10 or 12 to 1. A brake 13 is ly connected to the output shaft 2 from the parallel gear stage, to provide a means for braking the rotor 101.

The output shaft 2 from the parallel gear stage provides the input to the planet r of a variable ratio epicyclic gear stage VREGS of a wind turbine transmission. The variable ratio epicyclic gear stage includes a ity of planet gears 10, an annulus gear 11, and a sun gear 12.

The sun gear 12 drives an output shaft 3 that is operably coupled to a grid connected onous generator 103 through a one-way clutch 17. The speed of the generator 103 is maintained constant due to synchronism with the electrical grid to which the tor is coupled.

Alternative gear ements are possible which may include a different number of gear stages and the components identified may be connected to different gears in some or all of the gear stages. The input shaft 1 has an angular velocity ω1, and the output shaft has an angular velocity ω3. The embodiment of the gear arrangement described is advantageous in providing a high overall variable step-up ratio ω3/ω 1, however the present invention does not preclude alternative gear arrangements.

The output shaft 3 is driven at a constant speed, for example 1200, 1500 or 1800 rpm depending upon the electrical grid frequency and the number of poles in the generator 103, except when the one-way clutch 17 over-runs. The one-way clutch may over-run, for example, during transient fluctuations in wind speed, or upon shut-down of the wind turbine if the rotor 101 is braked to a rapid halt by the brake 13 but the generator 103 shaft takes a period of time to come to rest after being electrically disconnected from the grid, when generator deceleration is effected only by onal and windage losses in the generator itself.

The annulus gear 11 of the lic gear stage VREGS is operably connected to a wheel gear 14 which may drive or be driven by a pinion gear 15 mounted on a shaft 4. The pinion shaft 4 may drive or be driven at a variable speed by a reversible hydraulic unit 104 that is operable as either a lic pump or motor. The lic unit 104 forms part of the hydraulic control system HCS of an embodiment of the t ion.

The overall step-up ratio from the input shaft 1 of angular velocity ω1, to the main output shaft 3 of angular velocity ω3, is infinitely variable within a design range. The variable transmission ratio is achieved by the annulus gear 11 of the epicyclic gear stage VREGS acting as a speed osition gear either driving or being driven at variable speed by the hydraulic motor/pump unit 104. A superposition gear in this context is one member of the epicyclic gear stage VREGS that comprises three rotating members, and wherein one member is operably connected to the main output shaft 3 of known angular velocity ω3 and another member is operably connected to the input shaft 2 of variable angular velocity ω2 such that the angular velocity of the superposition gear or that of the input shaft 2 are calculable by superposition provided that the angular velocity of either is known. In alternative ments the input shaft, output shaft and superposition gear can be operably connected to different members of the lic gear stage VREGS. A second parallel gear stage SPGS between the annulus 11 and the pinion shaft 4 comprising a wheel gear 14 and the pinion gear 15 enables the ratio of this gear stage to be tuned such that an optimum speed and torque output can be obtained using a readily available hydraulic motor/pump unit. The design range of the overall variable step-up ratio ω3 to ω1 may be for example between 30 to 1 and 40 to 1.

The red embodiment transmission bed above has three slip states: a negative slip state, a zero slip state, and a positive slip state. At zero slip, the rotational velocity of the hydraulic motor/pump unit 104 is zero and the rotor has a zero slip rotor speed ω1zero . The ission has a zero slip rotor speed ω1zero so long as sufficient aerodynamic rotor power is available to maintain the main output shaft 3 at the constant speed of the grid connected onous generator 103 and the one-way clutch 17 is therefore not over-running. Negative slip occurs when the actual rotor speed ω1 is slower than the zero slip rotor speed ω1zero , and positive slip occurs when the actual rotor speed ω1 is faster than the zero slip rotor speed. When operating with negative slip, the hydraulic motor/pump unit 104 operates as a motor to make up the speed deficit and provide reaction torque to enable power transfer from the input shaft 1 to the main output shaft 3. When operating with positive slip, the hydraulic motor/pump unit 104 operates as a pump to accommodate a lower ω3/ω 1 ratio and provide reaction torque to enable power transfer from the input shaft 1 to the main output shaft 3.

Figure 3 shows a first preferred embodiment of a hydraulic circuit for the hydraulic motor/pump unit 104, and that forms part of the hydraulic control system HCS. The hydraulic circuit has a high pressure side HPS and a low pressure side LPS. As sed above, the transmission has two operating modes, a first mode in which the hydraulic motor/pump unit 104 operates as a pump, and a second mode in which the unit 104 operates as a motor. When the lic motor/pump unit 104 operates as a pump, solenoid valves 23, 24 are in the de-energised ons shown in Figure 3. A hydraulic charge pump 33 and reservoir 34 provide hydraulic fluid and system pressure. A pressure reducing/relieving valve 32 s fluid is provided at a constant charge pressure to the low pressure side of the hydraulic motor/pump unit 104. A low pressure lator 27 is provided to damp dynamic effects on the low pressure side of the hydraulic t.

When the hydraulic motor/pump unit 104 is operating as a pump, the ission provides the input power to pump fluid from the low pressure side to the high pressure side of the hydraulic motor/pump unit 104. The high pressure fluid is then relieved through a fixed orifice 28 and also through a relief valve 29 above a set pressure before being returned to the low pressure side of the hydraulic motor/pump unit 104 through a check valve 35, oil cooler 30, filter 31 and solenoid valve 24. The relief valve 29 is set to open at a predetermined differential pressure that corresponds with an upper limit torque on the hydraulic motor/pump unit 104 which, in turn, corresponds to an upper limit torque on the main output shaft 3. ably the upper limit torque on the output shaft 3 is the torque at the rated power of the wind turbine.

When the relief valve 29 is closed, the transmission is substantially rigid other than some small compliance provided by fluid flow through the orifice 28 and the high pressure accumulator 26.

Flow through the orifice 28 allows a small positive slip, and the high pressure accumulator 26 damps dynamic s on the high pressure side of the lic circuit. The relief valve 29 opens in response to increased re as the fluid flow rate through the orifice 28 ses due to increased speed of the motor/pump unit 104, increasing positive slip, and thereby allowing the speed of the rotor 101 to increase while the speed of the generator 103 remains constant.

When the motor/pump unit 104 is acting as a pump, a variable displacement supply pump 21 driven by an electric motor 22 provides cooling for the hydraulic fluid in the reservoir 34 - which may be heated in the course of performing other wind turbine hydraulic functions such as blade pitch and e yaw functions. When the variable displacement pump 21 is activated, additional fluid from the oir 34 is drawn into the hydraulic circuit through a check valve 36, the supply pump 21 and solenoid valve 23. This fluid is circulated through the oil cooler 30, filter 31 and second solenoid valve 24. Any additional fluid is then relieved from the hydraulic circuit through the pressure reducing/relieving valve 32 and returned to the system reservoir 34, such that the pressure on the low pressure side of the hydraulic circuit is maintained at the charge pressure set by the pressure reducing/relieving valve 32.

When the hydraulic motor/pump unit 104 operates as a motor, the solenoid valves 23 and 24 are in their energised positions, contrary to the ons drawn in Figure 3. The supply pump 21 is driven by the electric motor 22 to pump fluid from the low pressure side to the high pressure side of the supply pump 21. The fluid flow path from the high pressure side of the supply pump 21 is through valve 23, hydraulic motor/pump unit 104, valve 24, check valve 35, oil cooler 30, filter 31 and valve 24 to provide a closed loop circuit. The pressure reducing/relieving valve 32 provides the same function in this operating mode to maintain the set charge re on the low pressure side of the circuit. There is a small leakage flow from the high pressure side to the low pressure side of the circuit through the orifice 28, which does not contribute to useful work as it es the hydraulic motor/pump unit 104. This leakage flow is small and the consequent power loss is negligible, as the differential pressure is low when operating the transmission in this mode.

A re transducer 25 is positioned to provide a measure of the hydraulic fluid pressure on the high pressure side of the hydraulic circuit. The differential pressure across the hydraulic motor/pump unit 104 can be derived from the pressure at the transducer 25 at all times, because the low pressure side of the circuit is maintained at the set charge pressure by the pressure reducing/relieving valve 32. This differential re across the hydraulic motor/pump unit 104 is used as the measured l input parameter to set the position of the solenoid valves 23 and 24 and ore to select the operating mode of the transmission.

When the hydraulic motor/pump unit 104 operates as a motor, the transmission is operating with negative slip which is herein referred to as the variable eed operating mode. When the hydraulic motor/pump unit 104 operates as a pump, the transmission is operating with positive slip which is herein referred to as the torque limiting peed operating mode.

It will be appreciated that power input is required from the variable displacement supply pump 21 in the form of electrical power drawn by the motor 22 when operating in the variable ed ing mode, and that this power input (neglecting losses) is recovered by the generator 103. It will also be appreciated that there is a waste of power dissipated as heat when operating in the torque limiting high-speed operating mode. ting ciencies in the generator and mechanical and hydraulic systems of the transmission, the equations governing the wind e power relationships can be set out as follows: Rotor power: P101 = T1 · ω1 (1) Hydraulic motor/pump unit power: P104 = T1 · (ω1 – ω1zero ) = T4 · ω4 (2) Generator power: P103 = P101 – P104 = T3 · ω3 (3) Where T1 is the torque at component 1, T3 is the torque at component 3, and T4 is the torque at component 4, ω3 is the angular velocity of component 3, and ω4 is the angular velocity of component 4. The operating characteristics of the preferred embodiment transmission will now be described with reference to the figures. The transmission has four operating regions through the operating wind speed range. These regions are indicated by letters A to E in the figures.

Figure 4 shows a typical rotor power verses wind speed curve for an example wind turbine rotor with variable speed to maintain the most efficient TSR (solid thin line). A rotor power versus wind speed curve for the same wind turbine rotor with a nt rotor speed is also shown n dot line). In this example, the constant rotor speed shown is sed for approximately 9 ms-1 wind speed, and this constant rotor speed is taken as the zero slip rotor speed ω1zero for demonstration of operating characteristics of the transmission according to the present invention. Alternatively, the zero slip rotor speed may be optimised for a different wind speed and a ent rotor design, depending upon the particular wind turbine application.

In the first operating region A-B, the wind turbine rotor 101 of the red embodiment is controlled to run at its minimum operating speed, which may be determined considering machine inefficiencies, rotor dynamic stability and the hydraulic motor/pump unit 104 speed limit. In the second operating region B-C, the wind turbine rotor 101 of the preferred embodiment is lled to run at uously variable speed such that the rotor speed is dependent on wind speed and is optimised for maximum energy capture, subject to the maximum power rating of the variable displacement supply pump 21 and its motor 22. In regions A-B-C the hydraulic motor/pump unit 104 is operating as a motor (variable low-speed operating mode). Point C is the zero slip operating condition which may or may not be on the power curve for optimum TSR.

At higher wind speeds through the third and fourth regions C-D-E, the hydraulic motor/pump unit 104 is operating as a pump (torque limiting high-speed ing mode). In regions C-D-E, the wind turbine rotor 101 is passively controlled with narrow band speed control. ent rotor over-speed TOS above the wind turbine rotor 101 rated speed set point in the region D-E may occur due to wind turbulence; r, the magnitude of such over-speed events is ined within acceptable limits by the rotor pitch control .

As wind speed increases and point D is reached, the relief valve 29 will open and high pressure fluid will be relieved h the fixed orifice 28 and also through the relief valve 29.

The hydraulic motor/pump unit 104 must extract or supply power from and to the transmission.

The power extracted (or supplied) is a on of the rotor torque and speed as described by Equation (2). The net power shown in Figure 4 (large dashed line) is the useful rotor power which is calculated from the generator 103 power, less the hydraulic motor/pump unit 104 input power in variable low-speed operating mode, neglecting inefficiencies. This shows the improved energy capture at wind speeds below point C, and a lower cut-in wind speed, relative to the same wind turbine rotor at constant speed. Figure 13 shows an enlargement of detail D13 of Figure 4.

The hydraulic motor/pump unit 104 is preferably a constant displacement type and therefore exhibits a constant torque per unit differential pressure. The variable ratio epicyclic gear stage VREGS dictates that the speed of the hydraulic motor/pump unit 104 is proportional to rotor slip speed (ω1 – ω1zero ). Therefore, speed, torque and lic pressure/flow characteristics relating the rotor 101 and the hydraulic motor/pump unit 104 can be derived: Dimensionless rotor slip speed: %slip = (ω1 – ω1zero ) · 100 / ω1zero (4) Hydraulic pump unit speed: ω4 = K · %slip (5) Hydraulic motor/pump unit differential pressure: p104 = T4 / C (6) Hydraulic motor/pump unit volumetric flow rate: q104 = W · ω4 (7) Where the above constants are defined: K = Hydraulic motor/pump unit 104 angular velocity per percent rotor slip C = Hydraulic motor/pump unit 104 torque per unit ential pressure W = Hydraulic motor/pump unit 104 volumetric displacement per unit angular displacement The desired value for constant K can be achieved through the gear design of the transmission using known gear ratio relationships. The constants C and W are properties of the selected hydraulic motor/pump unit 104. For a particular wind turbine application, the constant K and a particular hydraulic pump unit are selected depending on the speed and power requirements for the motor/pump unit, limitations relating to m hydraulic system pressure, the range of variable rotor speed and the ratio required of the transmission, and the selection of a zero slip rotor speed.

Figures 5, 6 and 7 show speed and torque relationships according to ons (1) to (7), for a preferred embodiment transmission. In the variable low speed operating mode in regions A to C, the maximum power supplied to the hydraulic motor/pump unit is assumed to be d to 3% of the rated wind turbine electrical power output. This indicates a practical limitation on the power that can be supplied by the supply pump 21 and its motor 22 given the cost of the pump 21 and motor 22 relative to the aforementioned increasingly incremental improvement in energy capture due to improved rotor efficiency as the optimum TSR is approached (see Figure 1). In the e shown, this power limitation has the effect of requiring a slightly faster rotor 101 speed than for the m TSR speed for a part of the region B-C.

The present invention can achieve a more cost-effective solution by applying this power limitation to save more value in capital cost and hydraulic system losses than is sacrificed by having incrementally sub-optimal rotor energy capture. ing to Figures 8 and 9, the transmission control ter will now be described. In the variable low-speed operating mode (regions A-B-C), the wind turbine rotor 101 speed is controlled by controlling the fluid flow rate supplied by the le displacement supply pump 21 to the hydraulic pump unit 104. In this ing mode, the wind turbine control system uses the measured hydraulic motor/pump unit 104 differential pressure as a direct e of the unit 104 reaction torque, and this reaction torque or pressure value is then used to set the hydraulic motor/pump unit 104 speed for the desired rotor 101 speed. In the torque limiting high-speed operating mode (regions C-D-E), the wind turbine is passively controlled by the hydraulic circuit. The ission control parameter is therefore hydraulic motor/pump unit 104 torque (or pressure).

Figure 9 shows the hydraulic fluid flow control relationship which may for instance be implemented by the wind turbine controller by means of a look-up table or function to set the supply pump flow rate as a function of hydraulic motor/pump unit ential pressure which is measured by the high pressure transducer 25. The small difference in flow rates between the hydraulic motor/pump unit 104 and supply pump 21 is due to e flow through the orifice 28 when operating in variable low-speed operating mode.

The wind turbine control system also employs blade pitch control with a rotor speed set point corresponding to the desired rotor speed at rated power (point D). The blades will therefore normally be pitched fully to fine h regions A-B-C and most of region C-D. Blade pitch activity normally only occurs at and above rated power where the blades feather to limit the maximum rotor speed.

This control system, utilising pressure from the transducer 25, eliminates the need to measure wind speed or shaft speeds for the purposes of controlling the le ratio transmission in the le low-speed operating mode. Wind speed fluctuations and hence rotor power and torque fluctuations that occur at a faster rate than that at which the hydraulic fluid flow control is set to d are benign, as these will only cause the rotor to arily operate at a torque and speed combination that departs ly from the design characteristic. In practice, the se speed of the hydraulic fluid flow control system is set considering the rate of wind speed fluctuations, the achievable rate of rotor acceleration and deceleration, and practical limitations on the rate at which hydraulic fluid flow can be changed through the variable displacement hydraulic supply pump 21.

Simplified control logic is shown in Figure 10 for the transmission controlled by the hydraulic circuit of Figure 3. Letters ‘B’ and ‘C’ referred to in the flow chart of Figure 10 refer to points B and C in Figure 9. The control logic will be embodied in any le type of ller, such as a programmable logic controller or a programmable microprocessor for example. The two operating modes (variable low-speed and torque limiting high-speed) are switched between using the solenoid valves 23 and 24. When the variable low-speed operating mode is activated, the rotor speed and hence power is controlled by setting the variable displacement supply pump 21 flow rate in response to the lic motor/pump unit 104 differential pressure derived from the measured hydraulic pressure at the high pressure transducer 25.

Referring to Figure 10, the l logic shall now be described. When the wind turbine run flag 201 is enabled, wind speed is measured and if the measured wind speed 203 is greater than the programmed cut-in wind speed, the up routine 205 is entered. In the start-up routine the variable low-speed operating mode 207 is ted, the rotor speed demand input for the wind turbine pitch control algorithm is set to the minimum operating rotor speed 209, the supply pump 21 flow rate is set to its maximum hydraulic flow rate 211 corresponding to the minimum operating rotor speed, and when the rotor speed is sufficiently stabilised and controlled such that the ed error values between the generator and utility grid electrical phase angle and frequency are within programmed limits 213, the generator to grid contactor is closed to synchronise 215 the tor with the grid.

The generating mode 217 is entered upon completion of the start-up routine 205. In the ting mode, the rotor speed demand for the wind turbine pitch control algorithm is set to the rated (maximum nominal) rotor speed 219. Either the torque limiting high-speed operating mode 227 or the variable low-speed operating mode 229 is selected according to the measured hydraulic pressure compared with the programmed pressure set point C. The hysteresis loop 221 is provided in order to prevent excessive switching between the torque limiting high-speed operating mode and the variable low-speed operating mode when the measured lic pressure is fluctuating about the pressure set point C. Within the hysteresis loop 221, if the measured hydraulic pressure is less than the pressure set point C minus the hysteresis allowance 223 then the variable low-speed operating mode 229 is entered, and if the measured hydraulic pressure is r than the re set point C plus the hysteresis allowance 225, then the torque limiting high-speed operating mode 227 is entered. The order of conditions 223 and 225 means that if the measured hydraulic pressure is within the hysteresis band about the pressure set point C, then provided that the turbine run flag 247 s enabled, the operating mode (torque limiting high-speed or variable low-speed) will remain unchanged and the hysteresis condition 221 will be re-tested at the programmed cycle frequency.

Within the torque limiting high-speed operating mode 227, if the le low-speed operating mode was active 231 on the previous programme cycle then the variable low-speed ing mode is de-activated 233 by switching solenoid valves 23 and 24, and if the variable eed operating mode was not active on the previous programme cycle 231 then this remains true.

Within the variable low-speed ing mode 229, if the variable low-speed operating mode was active on the previous programme cycle 235 then the variable low-speed control mode 239 is entered directly, otherwise the variable low-speed operating mode is ted 237 by switching solenoid valves 23 and 24 and the variable low-speed control mode 239 is then entered. Within the variable low-speed control mode 239, if the measured hydraulic pressure is less than the programmed re set point B (condition 241) then the supply pump 21 hydraulic flow rate is set to its maximum flow rate 245 corresponding to the m operating rotor speed.

Otherwise the supply pump 21 hydraulic flow rate is varied 243 as a function of the measured hydraulic pressure, for variable rotor speed.

The ting mode 217 control logic process described in steps 219 through 247 of Figure 10 will normally be repeated at the programmed cycle frequency, provided that the wind turbine run flag 247 remains enabled, and other conditions not described in Figure 10 continue to be met, such as the measured wind speed ing above the cut-in wind speed and provided that no fault or alarm conditions are active. If the turbine run flag 247 is disabled then the wind turbine is shut down by the shut down routine 249.

Figure 3 illustrates the hydraulic circuit for one preferred embodiment of the lic control system for a transmission according to the present invention. Figures 11 and 12 show hydraulic circuits for alternative embodiments, and like numerals indicate like parts. Unless described below, the operation of the circuits is as described above, and control of those embodiments will be as per Figure 10.

In the embodiment shown in Figure 11, the igured hydraulic circuit allows the two solenoid valves 23 and 24 of the first ment to be replaced by a single solenoid valve 38 and onal check valves 37 and 39 to prevent reverse flows. Figure 11 shows solenoid valve 38 in the de-energised position, which corresponds to the torque limiting high-speed operating mode. In this operating mode, the operation of the hydraulic circuit is as described for Figure 3.

When the variable low-speed operating mode is selected by switching the solenoid valve 38 to the position ry to that shown in Figure 11, the circuit varies from the operation of that in Figure 3 in that the oil cooler 30 and filter 31 are located on the inlet side of the hydraulic motor/pump unit 104 which is the high re side of the hydraulic circuit.

Practical pressure limitations on hydraulic cooler and filter ents mean that the transmission with a hydraulic circuit configured as shown in Figure 11 may have a maximum pressure tion when operating in the variable low-speed operating mode. Such a pressure tion would be that the transmission should be ed from variable low-speed operating mode to torque limiting high-speed operating mode at a lower pressure, and therefore a lower rotor torque and lower wind speed than may otherwise be selected as the switch over point. The effect on the variable low-speed operating characteristics is easily calculable using the Equations 1 to 7, and it can be seen that point C on Figure 4 will be at a lower wind speed and on a power curve corresponding to the zero slip rotor speed but not necessarily on the power curve for optimum TSR. r ative embodiment of the hydraulic circuit is shown in Figure 12. In that embodiment, the two solenoid valves 23 and 24 of the first embodiment shown in Figure 3 have been eliminated, and no solenoid valve switching is required to change between variable low- speed and torque limiting high-speed operating modes. In this ative embodiment, the variable cement hydraulic supply pump 21 is controlled in the same way as described for the us embodiments, and check valves 39 and 40 prevent reverse flow so that the transmission will operate in variable low-speed operating mode when the flow delivered by the supply pump 21 is non-zero. The torque limiting high-speed operating mode is passively selected when the supply pump 21 flow rate is set to zero.

In the embodiment of Figure 12, the supply pump 21 cannot perform the hydraulic cooling pump function in the torque limiting high-speed ing mode as for the ment of Figure 3. Therefore, an additional pump may be supplied to provide this function. An additional oil cooler 41 is necessary in this embodiment because there is no other hydraulic cooling provided in the closed loop (the upper loop of Figure 12) that forms the main lic fluid flow path when operating in variable low-speed operating mode. Heat generation in this part of the hydraulic circuit is only due to losses in the hydraulic motor/pump unit 104 and the variable displacement lic supply pump 21, therefore the oil cooler 41 may be r than the oil cooler 30 for the torque limiting circuit (the lower loop of Figure 12).

Preferred embodiments of the invention offer a variable low-speed operating mode that can be configured to offer the most desirable features of a variable speed transmission for a wind turbine or similar application: increased energy e in light winds due to lower rotor speed and a lower cut-in wind speed, and correspondingly reduced noise levels in light winds. Preferred ments of the invention can be configured for a particular application such that the net benefit in terms of me energy capture per unit cost is maximised by selecting a suitable zero slip rotor speed and power and rotor speed limits in the variable low-speed operating mode.

Typically the variable low-speed operating mode may be active up to about 25% of the rated power and the minimum rotor speed in the variable low-speed operating mode may be about 75% of the zero slip rotor speed. The maximum hydraulic power supplied by the hydraulic motor / pump unit when in the le low-speed operating mode may lly be d to about 5% of the rated power, allowing the rotor TSR to depart slightly from the m TSR when this limit is enforced. The selection of these parameters can be made considering the quantifiable lifetime energy capture per unit cost given any particular known wind speed distribution.

Preferred embodiments of the invention have been bed by way of example only and modifications may be made o without departing from the scope of the ion.

For example, the le displacement hydraulic supply pump 21 in Figure 3, 11 or 12 may be directly driven ically (for example by a driveshaft or belt drive) from the transmission assembly instead of an by electric motor 22. This may be preferable for a particular application if the additional drive-train complexity required to drive the pump 21 can be incorporated at a lower cost than the electric motor and pump configuration shown.

While the hydraulic control system HCS has been described as forming an al part of a wind e, the hydraulic control system could instead be retrofitted to the transmission of an existing wind e to improve its performance.

As another example, while the hydraulic motor/pump unit is described as being of a positive displacement type, it could instead be a dynamic type.

As another example, while ing between operating modes and controlling the hydraulic fluid flow rate through the hydraulic motor/pump unit in the variable low-speed operating mode is described as occurring in response to measured hydraulic pressure, either could be done based on a different measured parameter such as measured rotor speed or wind speed for example.

However, automatic mode switching and variable low-speed control based on measured hydraulic pressure is preferred, because hydraulic pressure is in effect a direct measure of torque which responds faster than rotor speed to changes in aerodynamic rotor power. Additionally, wind speed measurements are usually provided by an anemometer at one location, lly on the wind turbine nacelle, which provides a measure of wind speed not always exactly representative of the equivalent wind speed for the net power at the rotor. Accurate and low cost hydraulic pressure transducers are readily available, making the control based on the measurement of hydraulic pressure practical and cost effective.

While the invention has been described using an example of a wind turbine application, the power generating system and hydraulic control system HCS could be applied to any other rotor or turbine driven power generating source wherein turbulence-induced torque fluctuations must be odated and/or ical variable speed is desirable for ed energy capture.

Tidal stream generators are a further example of such power generation sources.

To those skilled in the art to which the invention s, many changes in construction and widely differing embodiments and applications of the invention will suggest themselves without departing from the scope of the invention as defined in the appended claims. The disclosures and the descriptions herein are purely illustrative and are not intended to be in any sense limiting. Where specific integers are mentioned herein which have known equivalents in the art to which this invention s, such known equivalents are deemed to be incorporated herein as if individually set forth.

Claims

1.

A hydraulic control system for lling a variable ratio transmission of a power generating system comprising a rotor or turbine adapted to be driven by a fluid at variable 5 speed and an electrical generator adapted to be driven at a nt or variable speed, the transmission coupling the rotor or turbine to the ical generator, the hydraulic l system comprising: a hydraulic motor/pump unit adapted to be operably connected to a superposition gear; said hydraulic pump unit connected to a hydraulic circuit that comprises an e and/or a relief valve configured to open at a ermined hydraulic 10 pressure, the hydraulic circuit configured to switch between a variable low-speed operating mode and a torque ng high-speed operating mode, wherein in the torque ng high-speed operating mode the hydraulic motor/pump unit is adapted to be driven by the superposition gear and operates as a pump to drive hydraulic fluid through the orifice and/or relief valve to provide a passive torque limiting function, and wherein in the 15 variable low-speed operating mode the hydraulic motor/pump unit is adapted to be driven as a motor to drive the superposition gear and the hydraulic l system is configured to provide a desired rotor speed by controlling hydraulic fluid flow rate through the hydraulic motor/pump unit. 20 2. A hydraulic control system as claimed in claim 1, wherein the hydraulic motor/pump unit is of a positive displacement type.

3. A hydraulic control system as claimed in claim 1 or 2, wherein the hydraulic circuit is configured to switch between the variable low-speed operating mode and the torque 25 limiting high-speed operating mode in response to measured lic pressure in the hydraulic circuit or otherwise in response to measured torque on the transmission.

4. A lic control system as claimed in claim 1, n the hydraulic fluid flow rate through the hydraulic motor/pump unit is controlled in the variable low-speed operating 30 mode in response to measured hydraulic pressure in the hydraulic circuit or otherwise in response to measured torque on the transmission.

5. A hydraulic control system as claimed in any one of claims 1 to 4, further comprising a le displacement supply pump, wherein in the variable low-speed operating mode the 35 variable displacement supply pump is configured to supply fluid to drive the hydraulic motor/pump unit as a motor.

6. A hydraulic control system as claimed in claim 5, wherein the variable displacement supply pump is electrically driven. 5

7. A hydraulic control system as claimed in claim 5, wherein the variable displacement supply pump is ured to be directly driven by the transmission.

8. A hydraulic control system as claimed in any one of the preceding claims, further comprising a pressure measuring device arranged to measure the pressure in a high 10 pressure region of the hydraulic circuit.

9. A hydraulic control system as claimed in claim 8, wherein the pressure measuring device comprises a transducer. 15

10. A hydraulic control system as claimed in any one of the preceding claims, wherein the hydraulic circuit s a first fluid path and a second fluid path which respectively define fluid flow in the variable low-speed operating mode and the torque limiting high-speed operating mode. 20

11. A hydraulic control system as claimed in claim 10, wherein the first and second fluid paths are closed loop paths.

12. A lic l system as d in claim 10 or 11, further comprising a controller ured to switch the hydraulic circuit between the first and second fluid paths, at a 25 ermined hydraulic pressure.

13. A hydraulic control system as claimed in any one of claims 10 to 12, further comprising at least one solenoid valve for switching between the first and second paths. 30

14. A le ratio transmission for a power generating system, the transmission comprising: a transmission input adapted to be driven at a variable speed; a transmission output adapted to be driven at a constant or le speed; and a hydraulic control system as claimed in any one of claims 1 to 10. 35

15. A variable ratio transmission as claimed in claim 14, wherein the variable ratio transmission comprises a planetary gear arrangement n transmission input and the transmission output, the planetary gear arrangement comprising a planetary gear input operably ted to the transmission input, a planetary gear output operably connected to the transmission output, and the superposition gear.

16. A power generating system comprising: 5 a rotor or e adapted to be driven at a variable speed; a generator adapted to be driven at a constant or variable speed; and a variable ratio transmission as claimed in claim 14 or 15 that couples the rotor or turbine and the generator. 10

17. A power generating system as claimed in claim 16, wherein the power generating system has a rated power, and the power output from the power generating system when in the variable low-speed operating mode is limited to be less than or equal to about 25% of the rated power. 15

18. A power generating system as claimed in claims 16 or 17, wherein the rotor or turbine has a zero slip speed when the angular velocity of the superposition gear is zero, and the minimum rotor speed in the variable low-speed operating mode is greater than or equal to about 75% of the zero slip speed. 20

19. A power generating system as claimed in any one of claims 16 to 18, wherein the power generating system has a rated power, and the m power required by the hydraulic control system components to achieve the le low-speed operating mode is less than or equal to about 5% of the power generating system rated power. 25

20. A power generating system as claimed in any one of claims 16 to 19, n the tor is a synchronous generator operating at constant speed set by a larger generating system or grid. WO 95162 Rotor Efficiency 10 12 14 16 Tip Speed Ratio