GB2482879A

GB2482879A - Fluid-working machine with asymmetrically profiled multi-lobe ring cam

Info

Publication number: GB2482879A
Application number: GB1013777.6A
Authority: GB
Inventors: Niall James Caldwell; Daniil Sergeevich Dumnov; William Hugh Salvin Rampen; Alasdair Ian Fletcher Robertson; Uwe Bernhard Pascal Stein; Robert George Fox
Original assignee: Artemis Intelligent Power Ltd
Current assignee: Artemis Intelligent Power Ltd
Priority date: 2010-08-17
Filing date: 2010-08-17
Publication date: 2012-02-22
Also published as: GB201013777D0

Abstract

A fluid-working machine for a renewable energy generation device e.g. a wind turbine comprises a ring cam 307 and working chambers 303. The ring cam, which may be segmented, having an annular working surface around an axis of rotation and defining a plurality of waves. Each working chamber may be canted from a radial line and has a piston engaged with the working surface to reciprocate as the ring cam and working chambers relatively rotate. The individual waves of the ring cam working surface have an asymmetric profile about a peak or a trough but the ring cam may be rotationally symmetrical. The machine may operate in both a motoring and dominant pumping modes and the cam profile may give a longer exhaust stroke than an inlet stroke in the dominant mode to reduce working face slope relative to a breathing face slope and minimize Hertzian stress and piston side loads.

Description

1 Fluid-working machine with multi-lobe ring cam

3 Field of the invention

The invention relates to multi-lobe ring cams for fluid-working machines and to fluid- 6 working machines including such ring cams. The invention is particularly applicable 7 where the fluid-working machines are large, for example, pumps or motors in 8 renewable energy extraction devices, such as wind turbines.

Background to the invention

12 Fluid-working machines include fluid-driven and/or fluid-driving machines, such as 13 pumps, motors, and machines which can function as either a pump or as a motor in 14 different operating modes.

16 When a fluid-working machine operates as a pump, a low pressure manifold typically 17 acts as a net source of a working fluid and a high pressure manifold typically acts as 18 a net sink for a working fluid. When a fluid-working machine operates as a motor, a 19 high pressure manifold typically acts as a net source of a working fluid and a low pressure manifold typically acts as a net sink for a working fluid. Within this 21 description and the appended claims, the terms "high pressure" and "low pressure" 22 are relative, and depend on the particular application. In some embodiments, low 23 pressure working fluid may be at a pressure higher than atmospheric pressure, and 24 may be several times atmospheric pressure. However, in all cases, low pressure working fluid will be at a lower pressure than high pressure working fluid. A fluid- 1 working machine may have more than one low pressure manifold and more than one 2 high pressure manifold.

4 Large displacement ring cam fluid-working machines (i.e. fluid-working machines having a large rotating annular cam driving a plurality of radial pistons arranged 6 around the cam, with each piston typically reciprocating multiple times per cam 7 revolution) are known and are proposed for use in renewable energy applications in 8 which there is a low speed rotating input but a relatively high speed electrical 9 generator (Rampen, Taylor & Riddoch, Gearless transmissions for wind turbines, DEWEK, Bremen, Dec. 2006). Ring cam fluid-working machines typically have a 11 plurality of rollers rolling on a wave shaped cam and operatively connected to pistons.

12 Each piston is slideably engaged in a cylinder, the cylinder and piston together 13 defining a working chamber containing working fluid, in communication via one or 14 more valves with high and low pressure manifolds. The pistons are each operable to undergo reciprocating motion within the cylinder so as to vary the working chamber 16 volume, when the ring cam rotates, such that a cycle of working chamber volume is 17 executed, and during which working fluid may be displaced.

19 Ring cam fluid-working machines may be configured so that the pistons and cylinders are located inside the ring cam, the ring cam having an inward facing working 21 surface, or may be configured so that the ring cam has an outward facing working 22 surface and is located inside the pistons and cylinders. Indeed, ring cam fluid-working 23 machines of either configuration are also known in which either the ring cam rotates, 24 or the pistons and cylinders rotate. It is also possible for the ring cam to have both inward and outward facing working surfaces where the ring cam is located between 26 inner and outer rings of pistons and cylinders. It is even possible for the pistons and 27 cylinders to be aligned roughly parallel with the axis of rotation, and for the ring cam 28 to have one or more axially facing working surfaces.

Multi-cylinder fluid-working machines, including ring cam fluid-working machines, 31 may be variable displacement fluid-working machines (either pumps or motors, or 32 machines operable as either pumps or motors), wherein each working chamber is 33 selectable to execute an active (or part-active) cycle of working chamber volume in 34 which there is a net displacement of working fluid, or an idle cycle in which there is substantially no net displacement of working fluid, by the working chamber during a 36 cycle of working chamber volume, for regulating the time-averaged net displacement 37 of fluid from the low pressure manifold to the high pressure manifold or vice versa.

2 Large fluid-working machines (such as those suitable for renewable energy 3 generation) are typically subject to particularly high internal forces and pressures. For 4 example, the pressure of the high (and indeed low pressure) working fluid of a large scale ring cam fluid-working machine, of a size suitable for a wind turbine, is 6 particularly high. Consequently the forces received by the ring cam from the rollers 7 are also high, and it is known for the ring cam working surfaces to degrade. It has 8 been proposed to assemble large scale ring cams from a number of segments, and it 9 is known for excessive wear to occur to the roller and to the working surface due to discontinuities which appear on the working surface under pressure of a roller at the 11 interface between segments.

13 In particular, when the operating pressure of ring cam fluid working machines is very 14 high (for example, higher than 300 Bar), the repetitive surface stress (Hertzian stress) in the ring cam and roller can exceed levels (for example, 1.5GPa) which allow a long 16 working life for the ring cam. Additionally, it is desirable to have a high number of 17 lobes on the ring cam (shortest wavelength) to increase the speed multiplication 18 factor (the factor by which the working chamber cycle frequency is increased over the 19 shaft rotation rate), but the Hertzian stress in the working surface increases with increasing slope of the ring cam surface. Thus it is not possible simply to make the 21 rollers larger for the same size of piston, because the piston would anyway only apply 22 force to the roller over the same area, nor to have more or higher amplitude waves, or 23 the machine would become larger and heavier. The curvature of the cam is also 24 important in that the curvature of the cam determines the contact area between the cam and the roller.

27 Accordingly, there remains a need for a fluid-working machine and a ring cam for a 28 radial fluid-working machine of minimum weight, maximum speed multiplication 29 factor, and having extended working lifetime.

31 Summary of the Invention

33 According to a first aspect of the present invention there is provided a fluid-working 34 machine for a renewable energy generation device, the fluid-working machine comprising a ring cam and a plurality of working chambers, the ring cam having an 36 annular working surface extending around an axis of rotation of the ring cam, the 37 annular working surface defining a plurality of waves, each working chamber having a 1 piston, each piston in operative engagement with the ring cam working surface 2 (typically, via a cam engaging element, such as a roller or piston shoe which is 3 integral to the respective piston), the ring cam and working chambers being mounted 4 to rotate relative to each other, cycles of working chamber volume being thereby coupled to rotation of the ring cam relative to the working chambers, characterised in 6 that the individual waves of the ring cam working surface have an asymmetric profile.

8 By the individual waves of the ring cam working surface having an asymmetric profile 9 we refer to the profile (i.e. the working surface) of individual waves not being symmetric under reflection through a maximum or minimum. The profile of the 11 plurality of waves are typically the same as each other and so the working surface 12 (comprising a plurality of waves) may be rotationally symmetric.

14 Thus, during use of the machine, the variation of working chamber volume during a cycle of individual working chamber volume is not symmetric in time. This contrasts 16 with known machines having working surfaces comprising waves with symmetrical 17 profiles in which the duration of exhaust and intake strokes of the working chambers 18 is the same, and where the magnitude of the rate of change of volume of the working 19 chamber is the same at corresponding times in the exhaust and intake strokes.

21 The fluid-working machine has a first operating mode. The first operating mode may 22 be pumping. The first operating mode may be motoring. The fluid-working machine 23 may have only the first operating mode. However, the fluid-working machine may 24 have a second operating mode. If the first operating mode is pumping, the second operating mode may be motoring. If the first operating mode is motoring, the second 26 operating mode may be pumping. The ring cam may rotate in the same sense 27 relative to the working chambers in both the first and the second operating modes. In 28 this case, the first operating mode is typically the dominant operating mode. Thus, 29 the fluid-working machine may be designed to operate with rotation of the ring cam relative to the working chambers in a first sense the majority of the time, but may be 31 usable with rotation of the ring cam relative to the working chambers in a second 32 sense. For example, the fluid-working machine may operate as a pump in the 33 dominant operating mode but also be usable as a motor, with the ring cam rotating in 34 the opposite sense relative to the working chambers, when motoring. Such a fluid-working machine is useful in the nacelle of a wind turbine, for example, where it may 36 be driven by the blades as a pump in normal use but occasionally be used to slowly 37 drive the blades to a desired configuration. The machine may be a machine which 1 operates more efficiently or has a longer operating lifetime in the first operating mode 2 than the second operating mode due to the profile of the waves of the ring cam.

3 Thus, the profile of the waves of the ring cam may be optimised for use in the 4 dominant operating mode but the machine may be operable, typically less efficiently, in a second operating mode in which the ring cam rotates in the opposite sense 6 relative to the working chambers.

8 The machine may be operated in the dominant mode more than 10 times, and 9 preferably more than 100 times, as much of the time as it is operated in the second mode.

12 The ring cam working surface comprises a plurality of waves having minima and 13 maxima of radius relative to the axis of rotation. The working chamber volume cycles 14 between a maximum when the cam engaging elements engage with the ring cam at bottom dead centres (BDC), and a minimum when the cam engaging elements 16 engage with the ring cam at top dead centres (TDC). Typically, for an outward facing 17 ring cam working surface, a plurality of pistons are arranged outside the ring cam and 18 the minima of radius define bottom dead centres (BDC) of cycles of working chamber 19 volume and the maxima of radius define top dead centres (TDC) of cycles of working chamber volume. Typically, for an inward facing ring cam working surface, a plurality 21 of pistons are arranged inside the ring cam and the maxima of radius define bottom 22 dead centres (BDC) of cycles of working chamber volume and the minima of radius 23 define top dead centres (TDC) of cycles of working chamber volume. A ring cam may 24 have both inwards and outward facing ring cam working surfaces. Thus, the fluid-working machine is typically a radial piston machine. However, a ring cam may have 26 a plurality of pistons arranged generally parallel to the axis of rotation of the ring cam 27 and one or more laterally facing working surfaces. The plurality of pistons are 28 typically radially arranged around the ring cam, and usually equally spaced.

Preferably, each working chamber has a volume which varies cyclically with 31 reciprocating movement of the respective piston. Preferably, each piston is slidably 32 mounted within a cylinder, such that a working chamber is defined between the 33 cylinder and piston. Typically, the fluid-working machine comprises a body and the or 34 each cylinder may be formed in the body. For example, the body may comprise or consist of a cylinder block. In some embodiments, the or each cylinder, or the or each 36 piston, may be articulated (typically via a spherical bearing). The or each piston may 37 be restrained within the body.

2 The volume of the working chamber varies cyclically with rotation of the ring cam.

3 The fluid-working machine comprises a low pressure manifold and a high-pressure 4 manifold, and a plurality of valves for regulating the flow of fluid between each working chamber and the low pressure and high-pressure manifold. Preferably the 6 plurality of valves are pressure-operated check valves, openable in one direction due 7 to pressure across said valves. Preferably the high-pressure valves (i.e. those valves 8 regulating the flow between the high pressure manifold and the working chamber) are 9 openable to allow fluid from the working chamber into the high-pressure manifold, when the pressure in said working chamber exceeds the pressure in the high 11 pressure manifold. Preferably the low-pressure valves (i.e. those valves regulating 12 the flow between the low pressure manifold and the working chamber) are openable 13 to allow fluid from the low pressure manifold into the working chamber, when the 14 pressure in said working chamber falls below the pressure in said low-pressure manifold.

17 Typically, at least one said valve associated with each working chamber is an 18 electronically controlled valve. The fluid-working machine typically comprises a 19 controller operable to control one or more said electronically controlled valves, on each cycle of working chamber volume and in phased relation to cycles working 21 chamber volume, to select the net volume of working fluid displaced by each working 22 chamber on each volume cycle. Typically, at least one said electronically controlled 23 valve associated with each working chamber is an electronically controlled poppet 24 valve. It may be that said electronically controlled valves are direct acting fast moving face-sealing poppet valves that are not openable against a substantial 26 pressure difference (e.g. not operable against a pressure of 10 bar). Thus, typically 27 the controller is operable to selectively hold open or close said electronically 28 controlled high-pressure valves, but not to open them against pressure in the high 29 pressure manifold, and typically the controller is operable to selectively close or hold closed said electronically controlled low-pressure valves, but not to open them 31 against pressure in the working chamber.

33 When the low-pressure valves are held open for a full cycle of working chamber 34 volume, the working chamber conducts an idle stroke in which there is no net displacement of fluid between the low-and high-pressure manifolds. To transfer fluid 36 from the low-pressure manifold to the high pressure manifold, the controller must 37 selectively close the low-pressure valve in a contracting stroke of the working 1 chamber, which may cause the high-pressure valve to open to exhaust fluid to the 2 high-pressure manifold in a so-called pumping stroke. The controller may then, if 3 electronically controlled high-pressure valves are employed, selectively hold open 4 said high-pressure valves during the subsequent intake stroke to accept fluid from the high-pressure manifold in a so-called motoring stroke.

7 The controller is preferably operable, in the dominant operating mode, to selectively 8 command only one of a pumping stroke or a motoring stroke (except that a motoring 9 stroke preferably starts with a very small pumping stroke). In some embodiments the controller is operable, in the alternative operating modes, to selectively command the 11 other of a pumping stroke or a motoring stroke.

13 The waves have opposite first and second faces, each face extending between a 14 maximum and an adjacent minimum. It may be that for each wave the first and second face have different arc lengths. In this case, as the ring cam is typically 16 rotated relative to the working chambers at a substantially constant angular velocity 17 the intake and exhaust strokes will have different durations.

19 Typically, all of the waves have the same profile and the arc length of each first face is the same and the arc length of each second face is the same. However, it may be 21 that some or all of the waves have first faces with different arc lengths. It may be that 22 some or all of the waves have second faces with different arc lengths.

24 Preferably, one of the first and second faces is a working face on which the cam following elements bear when the pressure in the working chamber most exceeds the 26 pressure in the low-pressure manifold and the other face is a breathing face.

28 Preferably the arc length of the working faces is larger than the arc length of the 29 breathing faces. Preferably the arc length of the working faces is more than 10%, and more preferably 20%, larger than the arc length of the breathing faces. Preferably, the 31 working faces extend for more than half of the arc of the ring cam (and typically for 32 >55% or >60% of the arc of the ring cam), and the breathing faces extend for less 33 than half of the arc of the ring cam (and typically <45% or <40% of the working 34 surface). Preferably, the working faces extend over less than two thirds of the arc of the ring cam.

1 Thus, as the arc length of the working faces is typically larger than the arc length of 2 the breathing faces, the mean slope of the working faces is typically less than the 3 slope of the breathing faces. The Hertzian stress (e.g. mean Hertzian stress or peak 4 Hertzian stress) in the working surface of ring cam fluid working machines is thus less than would be the case for known fluid-working machines in which the working and 6 breathing faces have a similar arc length. The side loads of the piston against the 7 cylinder are also reduced.

9 In machines which are pumps, or in which the dominant operating mode is pumping and which have an outward facing ring cam working surface, the working faces 11 extend between a maximum of working chamber radius and the next minimum of 12 working chamber radius around the ring cam in the direction of relative rotation (the 13 sense in which the ring cam moves relative to the working chambers if the working 14 chambers are fixed and the ring cam rotates and the opposite sense to which the working chambers rotate if the ring cam is fixed and the working chambers rotate).

16 Where the ring cam working surface faces inwards, the working faces extend 17 between a minimum of working chamber radius and the next maximum of working 18 chamber radius around the ring cam in the direction of relative rotation.

In this case, the exhaust stroke of each working chamber corresponds with the cam 21 engaging elements bearing on the working faces. Preferably, the working faces have 22 a greater arc length than the breathing faces. Thus, the exhaust stroke is preferably 23 longer than the intake stroke.

In machines which are motors, or in which the dominant operating mode is motoring 26 and which have an outward facing ring cam working surface, the working faces 27 extend between a minimum of working chamber radius and the next maximum of 28 working chamber radius around the ring cam in the direction of relative rotation 29 Where the ring cam working surface faces inwards, the working faces extend between a maximum of working chamber radius and the next minimum of working 31 chamber radius around the ring cam in the direction of relative rotation.

33 In this case, the intake stroke of each working chamber corresponds with the cam 34 engaging elements bearing on the working faces. Again, the working faces preferably have a greater arc length than the breathing faces. Thus, the intake stroke 36 is preferably longer than the exhaust stroke.

1 It may be that the pressure within a working chamber remains significantly above the 2 low pressure manifold pressure while the respective cam engaging element bears on 3 a first part (the part it first bears on) of the breathing faces. It may be that the 4 pressure within a working chamber remains close to or below the low pressure manifold while the respective cam engaging elements bears on a first part (the part it 6 first bears on) of the working faces. This can arise due to the (slight) compressibility 7 of practical working fluids. Thus, the rate of change of pressure within a working 8 chamber with time may reach zero when the respective cam engaging element has 9 passed 1.0-10.0% of the arc length of an entire wave after top dead centre or bottom dead centre.

12 The fluid-working machine may be configured such that the cam engaging elements 13 bear on the breathing faces when (or only when) the respective working chamber is 14 expanding (for example in embodiments where the fluid-working machine is a pump).

The fluid working machine may be configured such that the cam engaging elements 16 bear on the breathing faces when (or only when) the respective working chamber is 17 contracting (for example, in embodiments where the fluid-working machine is a 18 motor). The fluid-working machine may be configured such that, when rotation is in a 19 first direction, the cam engaging elements bear on the breathing faces when (or only when) the respective working chamber is expanding, and when rotation is in a 21 second direction, the cam engaging elements bear on the breathing faces when (or 22 only when) the respective working chamber is contracting (for example, in 23 embodiments where the fluid-working machine is a pump/motor operable as a 24 pumping mode in a first direction of rotation and as a motor in a second direction of rotation).

27 The variations in radius between the maxima and minima are generally small relative 28 to the diameter of the ring cam, for example, the difference between the radius at the 29 maximum and the radius at the minima is typically <5% of the mean radius of the ring cam.

32 Within this specification, we refer to the change in radius with angular position relative 33 to the axis of rotation, dr / do as the slope of the ring cam working surface.

34 Preferably, the rate of change of slope with angle, d2r / do2 is continuous. This is significant because the rate of change of slope with angle dictates the acceleration of 36 a cam engaging element which rolls or slides on the ring cam working surface. The 37 rate of change of slope with angle should never be sufficiently negative to cause a 1 cam following which rolls or slides on the ring cam working surface to disengage from 2 the working surface. Thus (d2r I do2) x (do / dt) (do / dt is the angular rotation rate) is 3 preferably less than the bias force biasing the cam engaging element against the 4 working surface divided by the combined mass of the piston and cam engaging element. There may be regions where the slope is constant, for example, lands with 6 no slope at or adjacent to minima or maxima, or regions of constant slope within the 7 first or second faces, for example, at the middle of the first and second faces.

9 Because the difference in radius between the maxima and minima is typically small relative to the radius of the ring cam, the rate of change of slope with angle is typically 11 very similar to the "curvature" of the working surface, i.e. the absolute value of the 12 second derivative of the working surface radius, Id2r I do2I. An outward facing ring 13 cam has convex portions of the working surface with d2r / do2 < 0 and concave 14 portions with d2r I do2> 0, whilst the opposite is true for an inward facing ring cam.

16 Preferably, for at least some (and typically each) wave a point or region of maximum 17 slope magnitude (typically a slope inflection point or region) of the working surface 18 intermediate a minimum and an adjacent maximum is not the same arc length from 19 the minimum and from the said maximum.

21 Thus, the rate of change of working chamber volume (and thus typically also flow 22 rate) is at a peak other than half way in time between each minimum and maximum 23 (assuming the rate of rotation of the ring cam relative to the working chambers is 24 substantially constant).

26 Preferably, each first face has a convex portion and a concave portion and the point 27 or region of maximum slope magnitude is located intermediate said portions.

28 Preferably, each second face has a convex portion and a concave portion and the 29 point or region of maximum slope magnitude is located intermediate said portions.

Preferably, each working face has a convex portion and a concave portion and the 31 point or region of maximum slope magnitude is located intermediate said portions.

32 Preferably, each breathing face has a convex portion and a concave portion and the 33 point or region of maximum slope magnitude is located intermediate said portions.

Preferably the maximum curvature of the convex portions of the working faces is less 36 than the maximum curvature of the convex portions of the breathing faces. Typically 37 the maximum curvature of the convex portions of the working faces is less than half, 1 or less than one third, of the maximum curvature of the convex portions of the 2 breathing faces.

4 Typically the maximum curvature of the concave portions of the working faces is the same or greater than the maximum curvature of the breathing faces.

7 Because the cam engaging element makes an angle to the working surface 8 compared to the working force it transfers to the piston, the Hertzian stress in the 9 working surface increases with increasing slope of the working surface. Furthermore, the curvature of the working surface is important in that the curvature of the working 11 surface determines the contact area between the working surface and a roller (being 12 an example of a cam engaging element) passing over the working surface. Thus, by 13 having a greater maximum curvature of the concave portions of the working face than 14 the convex portions, the curvature of the convex portions, where the contact area is least and the Hertzian stresses greater, can be less than would otherwise be required 16 given the constraint that the working surface has maxima and minima of a particular 17 angular separation and particular difference in radius from the axis of rotation.

19 Preferably the maximum curvature of the convex portions of the working faces is less than the maximum curvature of the concave portions of the working faces. Typically 21 the maximum curvature of the convex portions of the working faces is less than half, 22 or less than one third, of the maximum curvature of the concave portions of the 23 working faces.

Thus, the flow rate to or from the high pressure manifold via each working chamber, 26 and therefore the torque applied to the rotatable ring cam, is preferably asymmetric in 27 time and angle. This contrasts with conventional fluid-working machines using 28 eccentric cams in which, typically to achieve a smoother aggregate flow rate to or 29 from the high pressure manifold from a plurality of working chambers, the flow rate due to each working chamber is designed symmetric in time and angle.

32 Preferably, the angular separation (C) between a point or region of maximum slope 33 magnitude of the working face and the adjacent BDC is less than the angular 34 separation (D) between said point or region of maximum slope magnitude of the working face and the adjacent TDC. More preferably, C / 0 < 90%. Thus, the 36 maximum flow rate during exhaust strokes will typically occur before the respective 1 working chamber volume is at the mean of the volume of the working chamber at top 2 dead centre and at bottom dead centre.

4 Preferably, a maximum curvature of the working surface is not at a maximum or minimum of radius. Preferably, a maximum curvature of the working surface is 6 angularly spaced from a maximum or minimum of radius by 1.0-10.0% of the angular 7 separation of a wave. Preferably, a maximum curvature of the working surface is 8 angularly spaced by 1.0-10.0% of the angular separation of a wave from each 9 maximum or minimum of radius in the sense opposite to the sense of relative rotation (thus, so that the cam engaging elements bear on the curvature maxima shortly after 11 TDCorBDC).

13 It may be that the point of maximum curvature intermediate a minimum and an 14 adjacent maximum is not an angular separation half way between the said maximum and the said minimum. It may be that the maximum curvature of the working surface 16 is not at a maximum or minimum of radius. A maxima of curvature of the working 17 surface may be angularly spaced from a maximum or minimum of radius by 1.0- 18 10.0% of the angular extent of a wave, typically in a sense opposite to the direction of 19 relative rotation.

21 The fluid-working machine is typically part of a hydraulic circuit (for example it may be 22 a pump driving fluid around a hydraulic circuit or a motor driven by fluid within a 23 hydraulic circuit). The hydraulic circuit typically comprises a further fluid working 24 machine, which may also be a fluid working machine according to the present invention. The hydraulic circuit typically further comprises a fluid accumulator. The 26 fluid accumulator enables working fluid to be stored or received from the storage by 27 the fluid working machine as required. The resulting ability to vary the total volume of 28 working fluid in the remainder of the fluid circuit allows the hydraulic circuit to deal 29 with time differences between the displacement of working fluid by the fluid-working machine and the displacement of working fluid by the further fluid working machine in 31 the fluid circuit.

33 Typically, the working chambers are canted. They may for example be in the plane of 34 the ring cam but not extent directly radially outwards. Preferably, the working chambers are canted in the direction such that the axis of piston movement between 36 the points of maximum and minimum volume is closer to perpendicular to the working 37 faces than to the breathing faces.

2 Preferably the radius profile of the ring cam working surface is selected so that, at 3 least in use in the dominant operating mode, the working faces are subject to the 4 lowest peak stress, and so that the flow of fluid through the valves (in particular the low-pressure valves) caused by the operable engagement of the pistons with the 6 working surface causes the minimal energy loss.

8 Preferably, the controller is operable to control the timing of the opening or closing of 9 the electronically controlled valves to counter fluctuations in torque and flow arising from the asymmetric flow of working fluid out of the working chamber during each 11 cycle of working chamber volume. Typically the controller will receive feedback 12 values of physical properties of the renewable energy device, such as positions, 13 velocities and accelerations, and use said feedback values to select the opening and 14 closing of the valves and thus schedule in time and angle the application of torque to the ring cam (and delivery or acceptance of flow to the high-pressure manifold) by the 16 selected working chambers associated with said valves, to actively cancel the effects 17 said asymmetric flow.

19 Preferably the fluid working machine is part of a hydraulic circuit comprising a fluid compliance. Typically fluid compliance comprises one or more gas accumulators.

21 Preferably the ring cam is coupled to a large inertial load (or large inertial source), 22 such that the energy transferred fluidically by one working chamber in use is much 23 less (for example, one hundredth, or one thousandth) of the energy embodied in the 24 inertial load (or source). Typically said large inertial source is a hub and blade assembly of a wind turbine or tidal energy device.

27 Thus, the flow rate to or from the high pressure manifold via each working chamber, 28 and therefore the torque applied to the rotatable ring cam, may be asymmetric in time 29 and angle (i.e. orientation of the ring cam relative to the working chambers), and thus cause the fluid working machine to produce a varying aggregate flow and torque in 31 use, but the effect on the renewable energy device can be rendered negligible by the 32 use of the above techniques and equipment.

34 Typically, each roller, or other cam engaging element, is biased against the ring cam working surface. For example, each roller, or other cam engaging element, may be 36 biased against the working surface by an elastic member, such as a spring. Typically 37 the elastic member biases each piston against each roller, or other cam engaging 1 element, thereby biasing said roller (or other cam engaging element) against the 2 working surface. Alternatively, or in addition, each roller (or other cam engaging 3 element) and/or each piston, is biased against the working surface by fluid pressure 4 from within the respective working chamber, throughout a part or all of a cycle of working chamber volume. Typically, fluid from within the respective working chamber 6 is also in direct communication with each roller, or other cam engaging element, 7 thereby to bias said roller, or other cam engaging element, against the working 8 surface, and further to separate the roller, or other cam engaging element, from the 9 piston. For example, each said piston may define a passageway extending from the working chamber and into fluid communication with the roller and the adjacent 11 surface of the piston, so that high pressure fluid pools between the piston and the 12 roller, and functions as a self-balancing fluid bearing.

14 In practice, the force exerted on each cam engaging element in use can be substantial. This force varies periodically during cycles of working chamber volume 16 (and in some embodiments depends on the volume of fluid to be displaced by the 17 working chamber on a particular cycle of working chamber volume selected by the 18 controller).

In a second aspect, the invention extends to a kit of parts comprising a ring cam 21 having an annular working surface defining a plurality of waves, the individual waves 22 of the ring cam working surface having an asymmetric profile, and a working chamber 23 mounting chassis comprising a plurality of cylinders, or cylinder mountings, which kit 24 can be assembled to form a fluid-working machine according to the first aspect of the invention.

27 In a third aspect, there is provided a renewable energy generation device (such as a 28 wind turbine) comprising a fluid-working machine according to the first aspect of the 29 invention. The fluid-working machine may be coupled to a drive shaft driven by a renewable energy capture device, such as a shaft connected to the blades of a wind 31 turbine to receive energy from a renewable energy source (e.g. wind). Within this 32 specification and the appended claims by a renewable energy generation device we 33 include, amongst other machines, machines which generate electricity from wind, 34 such as wind turbines, or flowing water, such as tidal turbines or hydro-electric power generation turbines.

1 According to a fourth aspect of the present invention there is provided a ring cam 2 having an axis of rotation and an annular working surface defining a plurality of 3 waves, the individual waves of the ring cam working surface having an asymmetric 4 profile. Optional features of the ring cam working surface and the said waves correspond to the features discussed above in relation to the first and second aspect 6 of the invention.

8 The invention also extends in a fifth aspect to a method of operating a fluid-working 9 machine comprising providing a fluid-working machine according to the first aspect of the invention and rotating the ring cam relative to the working chambers and thereby 11 causing the volume of the working chambers to vary cyclically.

13 It may be that the working chambers remain fixed and the ring cam is rotated. It may 14 be that ring cam remains fixed and the working chambers are rotated. It may be that both the ring cam and the working chambers are rotated.

17 It may be that the duration of the intake and exhaust strokes is different.

19 Preferably, the rate of flow of working fluid during an exhaust stroke peaks before the volume of the respective working chamber reaches the mean of its volume at top 21 dead centre and bottom dead centre.

23 Optional features discussed in relation to any of the five aspects of the invention are 24 optional features of any one of the five aspects of the invention.

26 Brief Description of the Drawings

28 The invention will now be illustrated with respect to the following drawings in which: Figure 1 shows a wind turbine generator connected to an electricity network and 31 implementing the invention; 33 Figure 2 shows a section of a pump according to the invention for use in the wind 34 turbine generator of Figure 1; 36 Figure 3 shows a ring cam working surface profile for use in the pump of Figure 2; 37 and 2 Figure 4 shows the flow rate to and from a single working chamber of the pump of 3 Figure 2 when employing the ring cam profile of Figure 3.

Detailed Description

7 Figure 1 illustrates an example embodiment of the invention in the form of a Wind 8 Turbine Generator (WTG, 100), acting as the renewable energy device, and 9 connected to an electricity network (101). The WTG comprises a nacelle (103) rotatably mounted to a tower (105) and having mounted thereon a hub (107) 11 supporting three blades (109) known collectively as the rotor (110). An anemometer 12 (111) attached externally to the nacelle provides a measured wind speed signal (113) 13 to a controller (112). A rotor speed sensor (115) at the nacelle provides the controller 14 with a rotor speed signal (117). In the example system the angle of attack of each of the blades to the wind can be varied by a pitch actuator (119), which exchanges pitch 16 actuation signals and pitch sensing signals (121) with the controller. The invention 17 could be applied to a WTG without a pitch actuator.

19 The hub is connected directly to a pump (129), through a rotor shaft (125), acting as the rotatable shaft, which rotates in the direction of rotor rotation (127). The pump is 21 preferably of the type described with reference to Figure 2, and has a fluid connection 22 to a hydraulic motor (131), preferably of the type described in EP0494236. The fluid 23 connection between the pump and the hydraulic motor is through a high pressure 24 manifold (133) and a low pressure manifold (135), connected to their high pressure port and low pressure port respectively, and is direct in the sense that there are no 26 intervening valves to restrict the flow. The pump and hydraulic motor are preferably 27 mounted directly one to the other so that the high pressure manifold and low pressure 28 manifold are formed between and within them. A charge pump (137) continuously 29 draws fluid from a reservoir (139) into the low pressure manifold, which is connected to a low pressure accumulator (141). A low pressure relief valve (143) returns fluid 31 from the low pressure manifold to the reservoir through a heat exchanger (144) which 32 is operable to influence the temperature of the working fluid and is controllable by the 33 controller via a heat exchanger control line (146). A smoothing accumulator (145) is 34 connected to the high pressure manifold between the pump and the hydraulic motor.

A first high pressure accumulator (147) and a second high pressure accumulator 36 (149) (together acting as the fluid compliance) are connected to the high pressure 37 manifold through a first isolating valve (148) and a second isolating valve (150) 1 respectively. The first and second high pressure accumulators may have different 2 precharge pressures, and there may be additional high pressure accumulators with 3 an even wider spread of precharge pressures. The states of the first and second 4 isolating valves are set by the controller through first (151) and second (152) isolating valve signals respectively. Fluid pressure in the high pressure manifold is measured 6 with a pressure sensor (153), which provides the controller with a high pressure 7 manifold pressure signal (154). The pressure sensor may optionally also measure the 8 fluid temperature and provide a fluid temperature signal to the controller. A high 9 pressure relief valve (155) connects the high pressure and low pressure manifolds.

11 The hydraulic motor is connected to a generator (157), acting as the load, through a 12 generator shaft (159). The generator is connected to an electricity network through a 13 contactor (161), which receives a contactor control signal (162) from a generator and 14 contactor controller (163) and is operable to selectively connect the generator to or isolate the generator from the electricity network. The generator and contactor 16 controller receives measurements of voltage, current and frequency from electricity 17 supply signals (167) and generator output signals (169), measured by electricity 18 supply sensors (168) and generator output sensors (170) respectively, communicates 19 them to the controller (112) and controls the output of the generator by adjusting field voltage generator control signals (165) in accordance with generator and contactor 21 control signals (175) from the controller.

23 The pump and motor report the instantaneous angular position and speed of rotation 24 of their respective shafts, and the temperature and pressure of the hydraulic oil, to the controller, and the controller sets the state of their respective valves, via pump 26 actuation signals and pump shaft signals (171) and motor actuation signals and 27 motor shaft signals (173). The controller uses power amplifiers (180) to amplify the 28 pitch actuation signals, the isolating valve signals, the pump actuation signals and the 29 motor actuation signals.

31 The WTG further comprises blade sensors (185) (which might comprise one or more 32 of accelerometers, position sensors, velocity sensors or acoustic sensors) which 33 communicate blade vibrations via blade sensor signals (187) to the controller.

Figure 2 illustrates in schematic form a portion (301) of the pump (129) with 36 electronically commutated valves and a ring cam according to the invention. The 37 pump consists of a number of similar working chambers (303) in a radial 1 arrangement, of which only three are shown in the portion in Figure 3. Each working 2 chamber has a volume defined by the interior surface of a cylinder (305) and a piston 3 (306), which is driven from a ring cam (307) by way of a roller (308), and which 4 reciprocates within the cylinder to cyclically vary the volume of the working chamber.

The ring cam may be broken into segments mounted on the shaft (322), which is 6 firmly connected to the rotor shaft (125). There may be more than one bank of radially 7 arranged working chambers, arranged axially along the shaft. Fluid pressure within 8 the low pressure manifold, and thus the working chambers, greater than the pressure 9 surrounding the ring cam, or alternatively a spring (not shown), keeps the roller in contact with the ring cam. A shaft position and speed sensor (309) determines the 11 instantaneous angular position and speed of rotation of the shaft, and informs a 12 controller (112), by way of electrical connection (311, being some of the pump 13 actuation and pump shaft signals 171), which enables the controller to determine the 14 instantaneous phase of the cycles of each individual working chamber. The controller is typically a microprocessor or microcontroller, which executes a stored program in 16 use. The controller can take the form of a plurality of microprocessors or 17 microcontrollers which may be distributed and which individually carry out a subset of 18 the overall function of the controller.

There may be more than one bank of axially-spaced ring cams, the surfaces of which 21 rotate together.

23 Each working chamber comprises a low pressure valve (LPV) in the form of an 24 electronically actuated face-sealing poppet valve (313) which faces inwards toward the working chamber and is operable to selectively seal off a channel extending from 26 the working chamber to a low pressure conduit (314), which functions generally (in 27 the pumping mode) as a net source of fluid in use (or sink in the case of motoring).

28 The low pressure conduit is fluidically connected to the low pressure manifold (135).

29 The LPV is a normally open solenoid closed valve which opens passively when the pressure within the working chamber is less than the pressure within the low pressure 31 conduit, during an intake stroke, to bring the working chamber into fluid 32 communication with the low pressure manifold, but is selectively closable under the 33 active control of the controller via an electrical LPV control signal (315, being some of 34 the pump actuation and pump shaft signals 171) to bring the working chamber out of fluid communication with the low pressure manifold. Alternative electronically 36 controllable valves may be employed, such as normally closed solenoid opened 37 valves.

2 The working chamber further comprises a high pressure valve (HPV, 317) in the form 3 of a pressure actuated delivery valve. The HPV faces outwards from the working 4 chamber and is operable to seal off a channel extending from the working chamber to a high pressure conduit (319), which functions as a net source or sink of fluid in use 6 and is in fluid communication with the high pressure manifold (133). The HPV 7 functions as a normally-closed pressuring-opening check valve which opens 8 passively when the pressure within the working chamber exceeds the pressure within 9 the high pressure manifold. The HPV may also function as a normally-closed solenoid opened check valve which the controller may selectively hold open via an 11 HPV control signal (321, being some of the pump actuation and pump shaft signals 12 171) and once the HPV is opened, by pressure within the working chamber. The HPV 13 may be openable under the control of the controller when there is pressure in the high 14 pressure manifold but not in the working chamber, or may be partially openable.

16 In a normal mode of operation described in the prior art (for example, EP 0 361 927, 17 EP 0 494 236, and EP 1 537 333), the controller selects the net rate of displacement 18 of fluid to the high pressure manifold by the hydraulic pump by actively closing one or 19 more of the LPVs typically near the point of maximum volume in the associated working chamber's cycle, closing the path to the low pressure manifold and thereby 21 directing fluid out through the associated HPV on the subsequent contraction stroke.

22 The controller selects the number and sequence of LPV closures to produce a flow or 23 apply a torque to the shaft (322) to satisfy a selected net rate of displacement. As well 24 as determining whether or not to close or hold open the LPVs on a cycle by cycle basis, the controller is operable to vary the precise phasing of the closure of the LPVs 26 with respect to the varying working chamber volume and thereby to select the net rate 27 of displacement of fluid from the low pressure manifold to the high pressure manifold.

29 The pump has a dominant operating mode of selectively actuated pumping strokes whilst the ring cam rotates in the clockwise direction as illustrated in Figure 2 (note 31 that Figure 2 illustrates the pump viewed from the opposite direction to its illustration 32 in Figure 1). In some embodiments it has alternative operating modes which include 33 pumping whilst the cam rotates in the opposite direction, and motoring whilst rotating 34 in either direction.

36 The controller is operable to use blade sensor signals (187) to select the timing of the 37 opening and closing of the valves and thus schedule in time and angle the application 1 of torque to the ring cam (and delivery or acceptance of flow to the high-pressure 2 manifold) by the working chambers. One possible technique for this is disclosed in 3 GB 1003000.5, which is hereby incorporated by reference.

The working chambers of the pump are canted. They do not directly radially 6 outwards. The working chambers are canted about 10° in the clockwise direction in 7 Figure 2.

9 Figure 3 shows one section of the cam profile (200) of the pump of Figure 2, on which the roller (308) rolls in use. The profile is repeated typically 15-25 times on each ring 11 cam bank, and forms an effectively continuous working surface around the ring cam.

12 The cam profile is defined by the radius in mm (202) from the centre of rotation of the 13 ring cam and the angle (204) from an arbitrary reference point (206) through one 14 cycle of working chamber volume of the pump. The relative scale of the axes of Figures 3 has been selected for clarity and so does not accurately portray the depth 16 of the profile compared to the pitch (A) between the maxima (TDC) and the next 17 adjacent minimum (BDC), nor the pitch (B) between minima (BDC) and the next 18 maximum (TDC). Figure 3 also shows a reference (sine wave) cam profile (208) to 19 illustrate the difference between the ring cam according to the invention and a conventional ring cam.

22 The profile comprises a convex section (210) and a concave section (212), which 23 meet at working (214) and breathing (216) points of inflexion. The working surface 24 comprises a working region (218) extending over those areas of the working surface on which the roller rolls at any time when, in the dominant operating mode, the 26 pressure in the working chamber significantly exceeds the pressure in the low- 27 pressure manifold, for example it is over 100 Bar, due to the selective opening and 28 closing of low and high pressure valves (313,317) under the control of the controller.

29 The working surface also comprises a breathing region (220) that extends over those areas of the working surface which are not subjected to forces from significant 31 working chamber pressure, in the dominant operating mode.

33 The working regions form the majority of the face of the working surface extending 34 over region B (which functions as the working face), and the breathing regions comprise the majority of the face of the working surface extending over region A 36 (which functions as the breathing face). However, they do not align perfectly because 37 of the compressibility of typical working fluids, for example hydraulic oil. At the 1 beginning of a selected pumping stroke, the working chamber pressure rises 2 monotonically over a small angle after the low pressure valve closes (which typically 3 occurs at BDC but potentially occurs anywhere when the roller bears on the region 4 B), meaning that the working region starts a little beyond BDC. After a selected pumping stroke, the working chamber is still pressurised for a small angle after the 6 high pressure valve closes at TDC, extending the working region a little beyond TDC.

7 This allows the commutation of working chambers alternately to the high and low 8 pressure manifolds to occur with no significant pressure across the valves, increasing 9 the efficiency of the fluid working machine/pump and decreasing the operating noise.

11 In the example embodiment, the pitch (angular separation A) between the maxima 12 (TDC) and the next adjacent minimum (BDC) is less than the pitch (angular 13 separation B) between minima (BDC) and the next maximum (Too).

Thus, in the dominant operating mode, the exhaust stroke is longer than the intake 16 stroke. Further, in use in the dominant operation mode, the slope of the working faces 17 is less than the slope of the breathing faces. The Hertzian stress in the working 18 surface is thus reduced, in comparison to machines of the prior art in which the 19 working and breathing faces have a similar length. The side loads of the piston against the cylinder (305) are also reduced by the reduced slope of the working 21 faces.

23 The angular separation (C) between the working face point of inflexion (functioning as 24 the point of maximum slope magnitude) and the adjacent BOO is less than the angular separation (0) between working point of inflexion and the adjacent TDC, and 26 in this embodiment C I 0 < 90%. Thus, the maximum flow rate during exhaust 27 (pumping) strokes occurs before the working chamber volume is at the mean of its 28 volume at top dead centre and at bottom dead centre.

In the example embodiment, the maximum curvature of the convex portions of the 31 working faces is less than one half of the maximum curvature of the convex portions 32 of the breathing faces. Also the maximum curvature of the convex portions of the 33 working faces is less than one half of the maximum curvature of the concave portions 34 of the working faces.

36 The relatively flat, convex, region on the working face has a low curvature which 37 reduces the Hertzian stress increase due to said convexity. In contrast, the relatively 1 steep, concave, region has a high curvature which reduces the Hertzian stress 2 increase due to said steepness. Due to the canting of the working chambers, the 3 Hertzian stress in the region of the working point of inflection is minimised by the 4 sliding axis of the working chambers being close to perpendicular to the working surface in near the working point of inflection.

7 The low curvature convex surface is necessarily larger (in terms of the arc covered, 8 and in terms of working surface area) than the high curvature concave surface, and 9 thus, the flow rate to or from the high pressure manifold via each working chamber, and therefore the torque applied to the rotatable ring cam, is asymmetric in time and 11 angle. This contrasts with conventional fluid-working machines using eccentric cams 12 in which, typically to achieve a smoother aggregate flow rate to or from the high 13 pressure manifold from a plurality of working chambers, the flow rate due to each 14 working chamber is designed symmetric in time and angle.

16 The ring cam and fluid working machine according to the invention has a lower stress 17 in its working surface, and therefore longer lifetime, than ring cams according to the

18 prior art.

Figure 4 illustrates the theoretical flow rate (400) to and from a single working 21 chamber of the pump of Figure 2 when employing the ring cam profile of Figure 3.

22 The vertical axis (402) measures flow (in L/min), against the angle (404) from BDC in 23 Figure 3. Positive values represent the contraction strokes in the dominant operating 24 mode (i.e. flows to the low pressure manifold when not pumping, and flows to the high pressure manifold when pumping), while negative values represent the intake 26 strokes.

28 In the example embodiment shown in Figure 4, the rate of flow of working fluid during 29 an exhaust stroke peaks before the volume of the respective working chamber reaches the mean of its volume at TDC and BDC. The flow of working fluid during an 31 exhaust stroke is asymmetric about the angular midpoint of the working stroke 32 because of the different curvatures of the convex and concave regions of the working 33 face, and the flow is also asymmetric about the TDC and BDC points illustrated.

Also, the ring cam slope during the intake stroke is constant, which causes a 36 relatively constant flow period (410). Thus the mean pressure drop (generally related 37 to the square of the flow) through the low pressure valve (in the dominant operating 1 mode) is lower than if a more peaked profile were chosen, and the energy consumed, 2 by moving fluid into and out of the working chamber through the low pressure valve, 3 is reduced.

The theoretical flow rate illustrated in Figure 4 varies from the actual flow rate in use 6 due to fluid compressibility, fluid leakage from the working chamber, and the 7 dynamics of the high and low pressure valves. Further, the theoretical flow profile of 8 Figure 4 does not exactly match the slope of the cam profile of Figure 3 due to the 9 finite size of the roller, and the consequently varying contact angle.

11 The radius profile of the ring cam working surface of Figure 3 has been selected so 12 that, at least in use in the dominant operating mode of pumping and clockwise 13 rotation, the working faces are subject to the lowest peak or mean stress, and so that 14 the flow of fluid through the valves (in particular the low-pressure valves) caused by the operable engagement of the pistons with the working surface causes the minimal 16 energy loss. The optimisation of the tradeoff between these and other competing 17 criteria (such as aggregate flow smoothness) is within the skill of the competent 18 designer.

The first and second high pressure accumulators (147,149, acting as the fluid 21 compliance) and the ring cam being coupled to the hub and blade assembly (acting 22 as a large inertial source) render the effect of the varying aggregate flow and torque, 23 due to the asymmetric ring cam waves, on the WTG negligible. Furthermore, the 24 controller (112) is operable to control the timing of the opening or closing of the electronically controlled valves of at least the pump to counter fluctuations in torque 26 and flow arising from the asymmetric flow of working fluid out of the working 27 chambers of the pump in use, using the blade sensor signal (187) to select the timing 28 of the opening and closing of the valves and thus schedule in time and angle the 29 application of torque to the ring cam (and delivery or acceptance of flow to the high-pressure manifold) to further, actively, cancel the effects of the asymmetric flow.

32 In some embodiments the working and breathing points of inflexion may be extended 33 -that is to say, sections of the profile may be non-curved. Non-curved sections 34 typically lie between the concave and convex portions of the profile.

1 In some embodiments the cam comprises a series of, typically identical, segments, 2 abutting to form an effectively continuous working surface or surfaces. The working 3 surface is typically treated for hardness, for example using nitriding techniques.

Further modifications and variations may be made within the scope of the invention 6 herein described.