AXIAL PISTON DEVICE Field of Invention

This invention relates to rotary valved axial piston engines and pumps. In particular, although not solely, this invention relates to two or four-stroke axial piston internal combustion engines and pumps and to the gearing used to index the relative rotation of the crankshaft to the cylinder porting mechanism in such an engine or pump. Background

Axial piston machines are machines in which a plurality of axially extending cylinders, together comprising the cylinder cluster, are arranged in a rotationally symmetrical layout around a central axis coincident with the rotational axis of the crankshaft. Each cylinder may be parallel to the others or slightly inclined to each other.

Amongst the methods that can be used to convert the reciprocating motion of the pistons to the rotational motion of this crankshaft is a reciprocator. A reciprocator is an intermediate body free to rotate around a section of the crankshaft inclined to, and intersecting with, the crankshaft's rotational axis at an acute angle. The reciprocator is restrained against rotation with respect to the cylinder cluster so that the rotation of the inclined section of the crankshaft causes the reciprocator to nutate. This creates a predominantly axial oscillatory motion in the frame of reference of the cylinder cluster at points on the body of the reciprocator in a plane perpendicular to the axis of the inclined section of the crankshaft at its point of intersection with the crankshaft axis. The connection between the reciprocator and pistons can take many forms but generally connecting rods having two or more rotational degrees of freedom at either end are utilised.

The rotationally symmetrical layout of the cylinders in a two-stroke or four-stroke axial piston engine or pump makes it possible to use a single common rotary valve porting mechanism in which an opening in each cylinder head passes in succession past every port in the rotary valve head. This is for example

described in the patent specifications of US Patent 3,654,906 (Airas), US Patent 6,494,171 (Duke) both of which describe engines in which the rotary valve head is held stationary while both the crankshaft and cylinder cluster rotate either in the same direction with respect to the rotary valve head (co-rotating) or in opposite directions with respect to the cylinder head (counter-rotating). For both co-rotating and counter-rotating four-strokes the ratio between crankshaft and cylinder cluster relative rotational speed and cylinder cluster and rotary valve head relative rotational speed will generally be equal to the number of cylinders, with the torque transmitted through the indexing drive from the cylinder cluster to the crankshaft being a fraction of the total engine torque output.

It is generally desirable to make multi-cylinder internal combustion engines as compact as possible in order to save mass, space and hopefully cost. To achieve this end some production inline-cylinder engines have cylinder pitch spacings of as little as 1.1 times the cylinder bore. Whilst axial piston engines can have inherent size and mass benefits over inline combustion engines, there are still aspects which can be addressed other than space and size. The known axial piston engines should they have a reasonably large diameter, can have potential mass imbalance and bearing friction issues as a result of a relatively large mass rotating about the crankshaft axis. At a larger diameter of the cylinder cluster sliding speeds of the rotary valve seals at any given engine speed can be high resulting in high wear of the rotary valve seals and frictional losses. Furthermore large diameter cylinder cluster arrangements may also create high or undesirable friction on the piston side walls and seals and scraper rings as a result of high lateral forces on the pistons and the connecting rods created by centrifugal forces as the cylinder cluster spins about its axis.

It is therefore an object of the invention to provide an axial piston engine or pump which uses a compact port indexing drive as well as other engine components that allow the cylinder pitch spacing and overall engine to be kept

compact to address the immediately abovementioned disadvantages or to at least provide the public with a useful choice.

It is also desirable to make whatever gearing system is employed to index the crankshaft with the rotary valve head with the minimum number of gearing stages in order to improve engine efficiency while saving size and mass.

There are many ways that the necessary indexing of the cylinder cluster and the crankshaft can be implemented. US Patent 3,654,906 (Airas) describes three basic arrangements of cylinder cluster drive gearing for a four-stroke axial piston engine utilising a reciprocator and having a spinning cylinder cluster and planar rotary valve head. These three gearing arrangements are all positioned on the far side of the reciprocator from the cylinder cluster, two of the options incorporate epicyclic gear sets, and the other utilises a parallel layshaft. The epicyclic gear sets both have a sun gear on the crankshaft driving planet gears that in turn drive an annular gear. In one instance the annular gear is mounted rigidly off of the stationary head (US Patent 3,654,906 Figl) resulting in a co- rotating engine, and in the other instance the annular gear is mounted off of the spinning cylinder cluster (US Patent 3,654,906 Fig4) resulting in a counter- rotating engine. In both of these detailed epicyclic configurations the sun gear is mounted on a crankshaft that is transmitting torques much greater than the sun gear is transmitting, potentially making the sun gear and epicyclic gear set greater in diameter than would otherwise be desirable.

The sun and planet/s are very much smaller than the annulus, thus . imposing higher numbers of loading cycles on the teeth of the sun and planet gears. There are also smaller numbers of teeth engaged in the sun-planet mesh compared to the planet-annulus mesh. These are all factors that will tend to lead to an annulus that is stronger and heavier than otherwise necessary in order to match the necessary size of the more highly loaded sun and planet gears.

The third arrangement detailed in US Patent 3,654,906 (Airas) has a parallel layshaft (US Patent 3,654,906 Fig5). Unlike the epicyclic port indexing gear sets the engine is limited to being only of the co-rotating type with

crankshaft and cylinder cluster rotating in the same direction. The teeth of the spur gear on the crankshaft will probably need to be bigger and stronger than for an equivalent epicyclic gear set because there are not multiple planet gears to share the transmitted torque. The parallel layshaft gearing arrangement is also likely to take up far more space both radially and axially than an equivalent epicyclic gear set.

US Patent 6,494,171 (Duke) describes a double ended four-stroke planar rotary valve axial piston engine having three cylinders at each end and with an epicyclic gear set in one of the stationary rotary valve heads of the engine controlling the indexing of the cylinder porting. The sun gear is mounted on the crankshaft and the annulus gear is mounted off of the cylinder cluster. The planet gear carrier is mounted off of the stationary valve head that in conjunction with the rotating cylinder ports forms the rotary valve so that the cylinder cluster spins in the opposite direction to the crankshaft at one third of the crankshaft's speed. The relatively large gear ratios that are required for the port indexing gearing can lead to port indexing epicyclic gear sets being much larger diameter than would otherwise be desirable.

These prior art port indexing gear sets all utilise two sets of meshing gears in the path of the power flow, with losses consequently accruing in each stage. The symmetry of the crankshaft mounted gears about the crankshaft axis of rotation in these prior art port indexing gear sets also means that they make little or no contribution to the dynamic balancing of the axial piston machine, and in general additional dedicated balance masses that serve no other useful function will need to be added to the crankshaft assembly in order to dynamically balance the machine.

Accordingly it is another object of the present invention to provide an engine or pump with a gearing means to index the relative rotation of the cylinder porting to the cylinder cluster to address these abovementioned disadvantages or to at least provide the public with a useful choice.

Gears in the valve trains of any internal combustion engine can be noisy, due to the fluctuating torque output of the engine caused by the cyclic internal combustion process. This impulsively loads any gearing subjected to the engine torque output, and in extreme cases can even result in impact noise due to alternating torque flow through gear sets possessing backlash. This problem is made worse in engines with smaller numbers of cylinders and is exacerbated at low speed due to the lower stored rotational kinetic energy. It is known that intervening mass can attenuate noise transmission, and it would be of benefit if existing engine components in an axial piston engine could be utilised to help provide this gear noise attenuating intervening mass.

Accordingly it is an object of the present invention to provide an axial piston combustion engine or pump which utilises necessary engine components in a configuration which creates an intervention to the transmission of noise from the or at least some of the gearing utilised in such an engine, or to at least provide the public with a useful choice.

The angular momentum of the rotating components of IC engines can potentially influence vehicle handling in asymmetrical ways. For most engines possessing rotating components; as the orientation of the engine is changed a gyroscopic reaction torque will in most cases be applied by the engine to the vehicle, and if engine speed is changed a resultant torque will also in most cases be applied by the engine to the vehicle. By way of example of this; boat outboard motors having generally vertically orientated engine crankshafts will tend to pitch the boat down when the boat rolls in one direction and pitch the boat up when the boat rolls in the opposite direction. Any changes in the outboard motor's speed will typically have to be reacted by an extra torque applied to the engine by the steering mechanism. In a motorbike with an engine having a longitudinally aligned crankshaft, accelerating the engine will tend to tip the motorbike over in one direction, while decelerating the engine will tend to tip the motorbike over in the opposite direction. Similar effects can also be found in aircraft. It would be advantageous if such effects could be reduced.

Accordingly it is a further object of the present invention to provide an axial piston engine or pump with components placed so as to reduce gyroscopic reaction torque or to at least provide the public with a useful choice.

Rather than attaching a separate pump for the pumping of lubricating and cooling oil through an engine it would be advantageous to use some of the components of an axial piston engine for the purposes of creating a pump or pump like mechanism. The utilisation of existing components in an axial piston engine will allow for the engine to be reduced in size since no or less ancillary equipment such as a pump needs to be attached or work in conjunction with the engine. There are hence efficiencies to be gained by utilising some of the engine components for establishing a pump.

It is accordingly an object of the present invention to provide an axial piston engine which incorporates a means to create a pump or pump like mechanism from the gearing components which provide the indexing between the cylinder cluster and the ports, or to at least provide the public with a useful choice.

It is a characteristic of some meshing gear sets that they can create low frequency harmonic noise and uneven tooth wear due to the repeated meshing of one or more slightly irregularly formed or damaged teeth with a small subset of the total number of teeth on the meshing gear. This low frequency harmonic gear noise can be more noticeable and harder to attenuate than higher frequency noise created by the same meshing gears. Such noise is of particular concern for applications in which the gear sets must operate in close proximity to humans, which includes a large proportion of all internal combustion engines and pump installations. Implementing sound attenuating systems that could mitigate such noise is generally undesirable owing to the added weight, complexity and expense, if feasible it would be better to instead limit the generation of such noise. The uneven wear of a select few gear teeth on one gear or other of a meshing set of gears can also become a limiting factor for the life of the gear set. It would thus be advantageous if a way could be found to reduce the amount of

uneven tooth wear and low frequency harmonic generation within gear sets employed in axial piston machines. It is desirable in designing a gear set to have each tooth on a gear come into cyclic contact with every tooth on the meshing gear, a condition that is satisfied only when the numbers of teeth on each of the meshing gears are "relatively prime", in other words when the number of teeth on each of the meshing gears shares no common factors with the number of teeth on the gear with which it meshes. This is sometimes termed a "hunting tooth ratio" and helps to ensure even wear of all gear teeth, it can also serve to reduce the harmonic noise produced when a damaged or malformed tooth comes into cyclic contact with only a small proportion of the teeth on the meshing gear.

Accordingly it is an object of the present invention to provide gearing in an axial piston combustion engine or pump which addresses the abovementioned desiderata or to at least provide the public with a useful choice. Brief description of the invention In a first aspect the present invention consists in an axial piston device acting as an internal combustion engine or pump comprising; a crankshaft having a crankshaft axis and carrying an oblique crank journal having an oblique crank axis which is oblique to the crankshaft axis but aligned to intersect therewith at a point, a cylinder cluster body of at least two cylinders, each cylinder containing a complementary piston to each reciprocate along the reciprocating axis of a respective cylinder and each having a cross section matched to the cross section of the cylinder, said cylinder cluster body and pistons mounted to rotate relative to said crankshaft and about said crankshaft axis, each said cylinder in fluid connection with at least one inlet/outlet port therefor, a reciprocator mounted to rotate relative to the crankshaft about said oblique crank axis, said reciprocator in mechanical engagement with each piston to allow the requisite reciprocating displacement of each piston within its respective cylinder between top dead centre (TDC) and bottom dead centre

(BDC) upon the cylinder cluster body rotating relative to and about the said crankshaft axis, rotational constraint means to ensure that said cylinder cluster body and said reciprocator rotate at the same angular rate with respect to said crankshaft and oblique crank journal respectively, port providing means relative to which said cylinder cluster body rotates and each at least one inlet/outlet port of each said cylinder sealably moves save for when said inlet/outlet ports are sequentially presented to cylinder communication means said port providing means being selected from one or more of a plurality of fluid transfer ports and spark plugs rotational indexing means to index at some rate, the rotation of the cylinder cluster body with the port providing means to present the inlet/outlet ports to said cylinder communication means to time such presentation with the reciprocating motion of each piston in its cylinder so as to cycle through the stages of an internal combustion engine or pump (such as the 4 stages of a four stroke engine) wherein said indexing means is a planocentric gear.

Preferably said cylinder cluster body includes two or more cylinders in the case of a two-stroke internal combustion engine or pump. Preferably said cylinder cluster body includes an odd number of three or more cylinders in the case of a four-stroke internal combustion engine.

Preferably said reciprocating axis of each said cylinder is parallel to the crankshaft axis.

Preferably said mechanical engagement of said reciprocator with each piston is by a connection rod as an extension from or part of said reciprocator and offering sufficient degrees of freedom to allow the linear reciprocating motion of the pistons relative to a respective cylinder and the oscillation motion of the reciprocator.

Preferably said planocentric gearing comprises a first gear means mounted from said port providing means concentrically to said crankshaft axis and relative

to which said crankshaft rotates, said first gear means including gear _. teeth to operatively mesh with gear teeth of a first means rotatable, said first means rotatable journaled for rotation about an eccentric axis which is eccentric to the crankshaft axis and parallel to said crankshaft axis, wherein rotational constraint means are provided to and intermediate of said cylinder cluster body and said first means rotatable to constrain the rotation of said cylinder cluster body relative to said first means rotatable.

Preferably said rotational constraint means comprises of two or more intermediate bodies, intermediate of said cylinder cluster body and said first means rotatable, each having two rotational axes parallel to said crankshaft axis, the intermediate body rotational axes having the same eccentricity relative to each other as said eccentric axis to said crankshaft.

Preferably for each intermediate body, the first of the rotational axes is defined by a bearing held by said first means rotatable such that the intermediate body is able to rotate with respect to the first means rotatable about its said first axis, the second of the said two intermediate body rotational axes of each intermediate body is mounted in a bearing held by said cylinder cluster body.

Preferably said planocentric gear comprises a first gear means mounted from said port providing means concentrically to said crankshaft axis and relative to which said crankshaft rotates, said first gear means including gear teeth to operatively mesh with gear teeth of a first means rotatable, said first means rotatable journaled for rotation about an eccentric axis which is eccentric to the crankshaft axis and parallel to said crankshaft axis, wherein rotational constraint means are provided to and intermediate of said port providing means and said first means rotatable to constrain the rotation of said port providing means relative to said first means rotatable.

Preferably said rotational constraint means comprises of two or more intermediate bodies intermediate of said port providing means and said first means rotatable, each having two rotational axes parallel to said crankshaft axis,

the intermediate body rotational axes having the same eccentricity relative to each other as said eccentric axis to said crankshaft.

Preferably for each intermediate body, the first of the rotational axes is mounted in a bearing held by said first means rotatable such that the intermediate body is able to rotate with respect to the first means rotatable about its said first axis, the second of the said two intermediate body rotational axes of each intermediate body is mounted in a bearing held by said port providing means.

Preferably the first rotational axis of each said intermediate body is at the same distance from the axis of said first means rotatable as the second rotational axis of each intermediate body is from the crankshaft axis.

Preferably, in operation each intermediate body of said rotational constraint means rotates synchronously with but separately to said crankshaft.

Preferably each said intermediate body includes a first and second rotational axis crank portion which extends substantially lateral to said crankshaft axis.

Preferably an odd ^J number of equi-spaced intermediate bodies are provided.

Preferably said rotational constraint means comprises of at least three contact rollers mounted from said port providing means and each with their rotational axis parallel to and to orbit about the crankshaft axis, each said contact roller engaging in a respective cylindrical or elliptical camming surface of said first means rotatable, said camming surface being of a diameter equal to the diameter of the complementary contact roller plus two times the eccentricity of the first means rotatable from the crankshaft axis, wherein each said contact roller orbits within its respective cylindrical camming surface.

Preferably said rotational constraint means comprises of at least three contact rollers mounted from said first means rotatable and each with their rotational axis parallel to and to orbit about the eccentric axis, each said contact roller engaging in a respective cylindrical camming surface of said port providing means, said camming surface being of a diameter equal to the diameter of the

complementary contact roller plus two times the eccentricity of the first means rotatable from the crankshaft axis, wherein each said contact roller orbits within its respective cylindrical camming surface.

Preferably the contact rollers are equi-spaced. Preferably said contact rollers are not equi-spaced and the angle between any two adjacent rollers about the crank axis or eccentric axis is not greater than 180 degrees.

Preferably said crankshaft or extensions to said crankshaft may extend axially to both sides of said first means rotatable. Preferably said planocentric gear is located partially or wholly between the cylinders of said cylinder cluster body.

Preferably said planocentric gear is located partially or wholly between said reciprocator and said cylinder cluster body.

Preferably the angular positioning of said eccentric axis to said crankshaft axis, relative to said the angular positioning of said oblique crank axis to said crankshaft axis, results in the mass of said first means rotatable contributing to the dynamic balancing of the inertial force created by the motion of said reciprocator and/or said pistons to said crankshaft (preferably in general, the planet gear carrying eccentric journal or planet gear assembly carrying eccentric journal will be most advantageously situated for balancing the inertial couple created by the reciprocating components and oblique crank journal of the axial piston machine if its centre of mass is situated as far as possible from the oblique crank journal's axis in the plane in which both the oblique crank axis and the crankshaft axis lie). Preferably at least one of each of the bearing mounts allows a small amount of radial motion with respect to either the eccentric axis in the case of the bearing mounted to the first means rotatable or the crankshaft axis in the case of the bearing mounted to the port providing means, to which the first means rotatable is constrained so that small geometrical distortion in the first means

rotatable or the intermediate bodies or in the port providing means to which the first means rotatable rotation is constrained, can be absorbed.

Preferably the contact rollers are each mounted so as to allow a small amount of radial motion with respect to the axis of the object that they are mounted from (whether it be the first means rotatable or the port providing means) to which the first means rotatable is held irrotational, or the complementary cylindrical camming surfaces in which the contact rollers orbit are made very slightly ^' elliptical with the major axis of the ellipse being radial to the axis of the object that the cylindrical camming surface is provide to, whether it be the first means rotatable or the port providing means to which the first means rotatable is held irrotational, so that small geometrical distortions in the first means rotatable or the port providing means to which the first means rotatable is rotationally constrained, can be absorbed.

In a second aspect the present invention consists in or for an axial piston device acting as an internal combustion engine or pump, the engine or pump comprising; a crankshaft having a crankshaft axis and carrying an oblique crank journal having an oblique crank axis which is oblique to the crankshaft axis but aligned to intersect therewith at a point, a cylinder cluster body of at least two cylinders, each cylinder containing a complementary piston to each reciprocate along the reciprocating axis of a respective cylinder and each having a cross section matched to the cross section of the cylinder, said cylinder cluster body and pistons mounted to rotate relative to said crankshaft and about said crankshaft axis, each said cylinder in fluid connection with at least one inlet/outlet port therefor, a reciprocator mounted to rotate relative to the crankshaft about said oblique crank axis, said reciprocator in mechanical engagement with each piston to allow the requisite reciprocating displacement of each piston within its respective cylinder between top dead centre (TDC) and

bottom dead centre (BDC) upon the cylinder cluster body rotating relative to and about the said crankshaft axis, rotational constraint means to ensure that said cylinder cluster body and said reciprocator rotate at the same angular rate with respect to said crankshaft and oblique crank journal respectively, port providing means relative to which said cylinder cluster body rotates and each at least one inlet/outlet port of each said cylinder sealably moves save for when said inlet/outlet ports are sequentially presented to cylinder communication means being selected from one or more of a plurality of fluid transfer ports and spark plugs of said port providing means, rotational indexing means to index at some rate, the rotation of the cylinder cluster body with the port providing means to present the inlet/outlet ports to said cylinder communication means to time such presentation with the reciprocating motion of each piston in its cylinder so as to cycle through the stages of an internal combustion engine or pump (such as a four stroke cycle of a four stroke combustion engine), wherein said indexing means is a planocentric gearing comprising of a first gear means mounted from said port providing means concentrically to said crankshaft axis and relative to which said crankshaft rotates, said first gear means including gear teeth to operatively mesh with gear teeth of a first means rotatable, said first means rotatable journaled for rotation about an eccentric axis which is eccentric to the crankshaft axis and parallel to said crankshaft axis, wherein rotational constraint means are provided intermediate of (i) one of (a) said cylinder cluster body and (b) said port providing means, and (ii) said first means rotatable, to constrain the rotation of said cylinder cluster body or said port providing means with said first means rotatable, wherein said first means rotatable has external gear teeth and said first gear means has internal gear

teeth, said first gear means being of a PCD greater than said second means rotatable, a fluid pump defined in part by a piston, rotatable about said eccentric journal and with said eccentric journal, to be positioned intermediate of said first gear means and first means rotatable at a cavity region created away from the gear contact zone of said external and internal gear teeth, wherein said cavity region is sealed or confined between an upper and lower sealing surfaces, save

(and preferably provided through at least one of said upper and lower sealing surfaces) for an opening or openings for fluid communication into and out of said cavity.

In a further aspect the present invention consists in a planocentric gearing with an integral fluid pump comprising; an annular gear having internal gear teeth to operatively mesh with external gear teeth of a planetary gear, said planetary gear rotatably journaled for rotation about an eccentric to the axis of said annular gear,axle which has an axis parallel to said axis of said annular gear, wherein said eccentric axle is controlled to rotate about its axis of rotation at the same velocity as the orbiting rotation of said planetary gear with said first means rotatable, said annular gear means being of a PCD greater than said planetary, a piston, rotatable with said eccentric axle, positioned intermediate of said annular gear and said planetary gear at a cavity region created away from the mutual gear contact zone of said planetary gear and said annular gear, wherein said ^' cavity region is sealed or confined between an upper and lower sealing surfaces, save (and preferably provided through at least one of said upper and lower sealing surfaces) for an opening or openings for fluid communication into and out of said cavity.

Preferably there is provided a fluid pump comprising; a piston, rotatable about said eccentric journal and with said eccentric journal, positioned intermediate of said first gear means and first means rotatable at a cavity region created away from the gear contact zone of said external and

internal gear teeth, wherein said cavity region is sealed between an upper and lower sealing surfaces, save (and preferably provided through at least one of said upper and lower sealing surfaces) for an opening or openings for fluid communication into and out of said cavity. In a further aspect the present invention consists in an axial piston device acting as an internal combustion engine or pump comprising; a crankshaft having a crankshaft axis and carrying an oblique crank journal having an oblique crank axis which is oblique to the crankshaft axis but aligned to intersect therewith at a point, a cylinder cluster body of at least two cylinders, each cylinder containing a complementary piston to each reciprocate along the reciprocating axis of a respective cylinder and each having a cross section matched to the cross section of the cylinder, said cylinder cluster body and pistons mounted to rotate relative to said crankshaft and about said crankshaft axis, each said cylinder in fluid connection with at least one inlet/outlet port therefor, a reciprocator mounted to rotate relative to the crankshaft about said oblique crank axis, said reciprocator in mechanical engagement with each piston to allow the requisite reciprocating displacement of each piston within its respective cylinder between top dead centre (TDC) and bottom dead centre (BDC) upon the cylinder cluster body rotating relative to and about the said crankshaft axis, rotational constraint means to ensure that said cylinder cluster body and said reciprocator rotate at the same angular rate with respect to said crankshaft and oblique crank journal respectively, port providing means relative to which said cylinder cluster body rotates and each at least one inlet/outlet port of each said cylinder sealably moves save for when said inlet/outlet ports are sequentially presented to cylinder communication means being selected from one or more of a plurality of fluid transfer ports and spark plugs of said port providing means,

rotational indexing means to index at some rate, the rotation of the cylinder cluster body with the port providing means to present the inlet/outlet ports to said cylinder communication means to time such presentation with the reciprocating motion of each piston in its cylinder so as to cycle through the stages of an internal combustion engine or pump (such as the 4 stages of a 4 stroke internal combustion engine), wherein said indexing means is a fixed differential gear. Preferably said cylinder cluster body includes two or more cylinders in the case of a two-stroke internal combustion engine or pump. Preferably said cylinder cluster comprises of an odd number of three or more cylinders in the case of a four-stroke internal combustion engine.

Preferably said reciprocating axis of each said cylinder is parallel to the crankshaft axis.

Preferably said mechanical engagement of said reciprocator with each piston is provided by a connection rod as an extension from or part of said reciprocator and offering sufficient degrees of freedom to allow transfer of the linear reciprocating motion of the pistons relative to a respective cylinder to the oscillation motion of the reciprocator and visa versa.

Preferably said fixed differential gear comprises a first gear means mounted from said port providing means concentrically to said crankshaft axis and relative to which said crankshaft rotates, said first gear means including gear teeth to operatively mesh with a first gear of a first means rotatable, said first means rotatable is journaled for the concentric rotation of said first gear about an eccentric axis which is eccentric to the crankshaft axis and parallel to said crankshaft axis, said first means rotatable including a second gear to rotate concentrically about said eccentric axis and to operatively engage with gear teeth of a gear means of said cylinder cluster body which rotates concentrically about said crankshaft axis, wherein said first and second gear of said first means rotatable are gears selected from one of a pairing of (a) internal and internal gears respectively,

(b) internal and external gears respectively,

(c) external and internal gears respectively, and

(d) external and external gears respectively the gears of said pairing to mesh with said gear teeth of said first gear means and said gear means of said cylinder cluster body respectively, each such gear teeth of said first gear means and of said gear means of said cylinder cluster body selected from (i) an internal and (ii) and external configuration, being opposite to the first and second gear of said first means rotatable.

Preferably said first gear of said first means rotatable is on a different PCD than the gear teeth of said first gear means.

Preferably said second gear of said first means rotatable is on a different PCD than the gear teeth of said gear means of said cylinder cluster body.

Preferably said first means rotatable is mounted on a bearing to rotate about an eccentric axle carried by said crankshaft. Preferably said crankshaft or extensions to said crankshaft extend axially to both sides of said first means rotatable.

Preferably said fixed differential gearing is located partially or wholly between the cylinders of said cylinder cluster body.

Preferably said fixed differential gearing is located partially or wholly between said reciprocator and said cylinder cluster body.

Preferably the angular positioning of said first means rotatable relative to said crankshaft is, in relation to said oblique crank axis angular positioning relative to the crankshaft, resultant in the mass of said first means rotatable contributing to the dynamic balancing of the inertial forces created by the motion of said reciprocator and/or said pistons to said crankshaft (preferably in general, the planet gear carrying eccentric journal or planet gear assembly carrying eccentric journal will be most advantageously situated for balancing the inertial couple created by the reciprocating components and oblique crank journal of the axial piston machine if its centre of mass is situated as far as possible from the

oblique crank journal's axis in the plane in which both the oblique crank axis and the crankshaft axis lie).

In a further aspect the present invention consists in a fixed differential gearing with an integral fluid pump comprising; an annular gear having internal gear teeth to operatively mesh with external gear teeth of a planetary gear, said planetary gear rotatably journaled for rotation about an eccentric to the axis of said annular gear axle which has an axis parallel to said axis of said annular gear, wherein said eccentric axle is controlled to rotate about it axis of rotation at the same velocity as the orbiting rotation of said planetary gear with said annular gear, said annular gear being of a PCD greater than said planetary gear, a piston, rotatable with said eccentric axle, positioned intermediate of said annular gear and said planetary gear at a cavity region created away from the mutual gear contact zone of said planetary gear and said annular gear, wherein said cavity region is sealed or confined between upper and lower sealing surfaces, save (and preferably provided through at least one of said upper and lower sealing surfaces)for an opening or openings for fluid communication into and out of said cavity.

Preferably there is provided a fluid pump comprising; a piston, rotatable about said eccentric axis and with said eccentric journal, said piston positioned intermediate of ;

(i) an externally toothed gear being of one of said first and second gear of said first means rotatable, and

(ii) an internally toothed gear being of one of said port providing means and cylinder cluster body, at a cavity region created away from the gear contact zone of said external and internally toothed gears, wherein said cavity region is sealed or confined between upper and lower sealing surfaces, save (and preferably provided through at least one of said upper

and lower sealing surfaces) for an opening or openings for fluid communication into and out of said cavity.

In a further aspect the present invention consists in a fixed differential gearing including an externally toothed planet gear journaled for rotation about and carried by an eccentric axle disposed from a crankshaft, said eccentric axle having an eccentric axis parallel to the axis of said crankshaft, said externally toothed planet gear meshing with an internally toothed annular gear which is concentric with said crankshaft axis wherein there is provided mounted for rotation with said eccentric axle and about said eccentric axis, a means for dividing the cavity between said planet and annular gear defined at a region of non contact of said meshing gears, said cavity sealed or confined by upper and lower sealing surfaces to said annular and planet gears, and wherein there is provided an inlet and outlet opening to allow fluid communication to be established between said cavity to the exterior, upon the relative rotation of said annular gear with said planet gear, to allow the establishing of a displacement of fluid onto and out of the divided cavity as said planetary gear rotates relative to said annular gear.

In a further aspect the present invention consists in an axial piston device acting as an internal combustion engine or pump comprising; a crankshaft having a crankshaft axis and carrying an oblique crank journal having an oblique crank axis which is oblique to the crankshaft axis but aligned to intersect therewith at a point, a cylinder cluster body of at least two cylinders, each cylinder containing a complementary piston to each reciprocate along the reciprocating axis of a respective cylinder and each having a cross section matched to the cross section of the cylinder, said cylinder cluster body and pistons mounted to rotate relative to said crankshaft and about said crankshaft axis, each said cylinder in fluid connection with at least one inlet/outlet port therefor,

a reciprocator mounted to rotate relative to the crankshaft about said oblique crank axis, said reciprocator in mechanical engagement with each piston to allow the requisite reciprocating displacement of each piston within its respective cylinder between top dead centre (TDC) and bottom dead centre (BDC) upon the cylinder cluster body rotating relative to and about the said crankshaft axis, rotational constraint means to ensure that said cylinder cluster body and said reciprocator rotate at the same angular rate with respect to said crankshaft and oblique crank journal respectively, port providing means relative to which said cylinder cluster body rotates and ^' each at least one inlet/outlet port of each said cylinder sealably moves save for when said inlet/outlet ports are sequentially presented to cylinder communication means being selected from one or more of a plurality of fluid transfer ports and spark plugs of said port providing means rotational indexing means to index at some rate, the rotation of the cylinder cluster body with the port providing means to present the inlet/outlet ports to said cylinder communication means to time such presentation with the reciprocating motion of each piston in its cylinder so as to cycle through the stages of an engine or pump (such as the 4 stages of a 4 stroke internal combustion engine), wherein said indexing means is epicyclic gearing.

Preferably said indexing means consists of an epicyclic gearing comprising of a sun gear connected to said crankshaft to mesh with one or more planet gears free to rotate on bearings connected to said port providing means, said planet(s) meshing also with an internally toothed annular gear connected to said cylinder cluster body which is concentric to said crankshaft in the same plane as said sun gear.

Preferably said epicyclic gearing is located partially or wholly between the cylinders of said cylinder cluster body.

Preferably said epicyclic gearing is located partially or wholly between said reciprocator and said cylinder cluster body.

Preferably said first and second gears of said first means rotatable are selected from one of a pairing of: (a) and internal and external gear respectively, and

(b) and external and internal gear respectively to each engage as a meshing pair, with an opposite external or internal as the case may be, gear of said port providing means and said cylinder cluster body, wherein for each meshing pair, the number of teeth have prime factorisations which have no common factors.

Preferably said first and second gears of said first means rotatable are selected from one of a pairing of:

(a) and both internal gear respectively, and

(b) and both external gear respectively to each engage as a meshing pair, with an opposite external or internal as the case may be, gear of said port providing means and said cylinder cluster body, wherein for each meshing pair, the number of teeth have prime factorisations which have no common factors.

Preferably said rotational constraint means is positioned so that constraint forces occur at radii outwardly of the PCD of the first gear means.

Preferably said two rotational axes of said intermediate body are positioned at radii remaining outwardly of the PCD of the first gear means.

Preferably a first of said rotational axes of said intermediate body is disposed outwardly of the PCD of said gear teeth of said first means rotatable. Preferably for each intermediate body, an arm is provided extending radially to eccentric axis and outwardly of the PCD of said gear teeth of said first means rotatable, wherein at or towards the distal end of said arm there is provided said first of said rotational axes of said intermediate body.

Preferably said camming surface is located at radii outwardly of the PCD of said first means rotatable or said first gear means of said port providing means as the case may be and from which it is disposed.

Preferably said rotational axis of said contact rollers is located at a radius outwardly of the PCD of said first means rotatable or said first gear means of said port providing means as the case may be and from which it is disposed.

This invention may also be said broadly to consist in the parts, elements and features referred to or indicated in the specification of the application, individually or collectively, and any or all combinations of any two or more of said parts, elements or features, and where specific integers are mentioned herein which have known equivalents in the art to which this invention relates, such known equivalents are deemed to be incorporated herein as if individually set forth.

As used herein the term "and/or" means "and" or "or", or both. As used herein "(s)" following a noun means the plural and/or singular forms of the noun.

The term "comprising" as used in this specification means "consisting at least in part of. When interpreting statements in this specification which include that term, the features, prefaced by that term in each statement, all need to be present but other features can also be present. Related terms such as "comprise" and "comprised" are to be interpreted in the same manner.

It is intended that reference to a range of numbers disclosed herein (for example, 1 to 10) also incorporates reference to all rational numbers within that range (for example, 1, 1.1, 2, 3, 3.9, 4, 5, 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 expressly stated in this application in a similar manner.

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 features of the invention. Unless specifically stated otherwise, reference to such external documents 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 general knowledge in the art.

This invention may also be said broadly to consist in the parts, elements and features referred to or indicated in the specification of the application, individually or collectively, and any or all combination s of any two or more said parts, elements or features, and where specific integers are mentioned herein ^' which have known equivalents in the art to which this invention relates, such known equivalents are deemed to be incorporated herein as if individually set forth. The invention consists in the foregoing and also envisages constructions of which the following gives examples. Brief description of the drawings

Preferred forms of the present invention will now be described with reference to the accompanying drawings in which; Figure 1 is a cross sectional view through a preferred form of the present internal combustion engine being in this case a four-stroke and having a piston and cylinder assembly rotated into the sectional plane and incorporating a planocentric gear set located between the cylinders of the cylinder cluster for driving the cylinder porting; Figure 2 is a cross sectional view through a preferred form of the present internal combustion engine being in this case a four-stroke and having a piston and cylinder assembly rotated into the sectional plane and incorporating a fixed differential gear set located between the reciprocator and cylinder cluster for driving the cylinder porting;

Figure 3 is a cross sectional view through a preferred form of the present internal combustion engine being in this case a four-stroke and having a piston and cylinder assembly rotated into the sectional plane and incorporating a fixed differential gear set located on the opposite side of the reciprocator from the cylinder cluster for driving the cylinder porting;

Figure 4 is a simplified cross section perpendicular to the crankshaft axis of the cylinder cluster, of a compact three cylinder .axial piston engine or pump demonstrating the available space between the cylinders;

Figure 5 is a simplified cross section perpendicular to the crankshaft axis of the cylinder cluster of a compact five cylinder axial piston engine or pump demonstrating the available space between the cylinders;

Figure 6 is a simplified cross section perpendicular to the crankshaft axis of the cylinder cluster of a compact seven cylinder axial piston engine or pump demonstrating the available space between the cylinders; Figure 7 is a simplified cross section of an embodiment of a fixed differential gear set for indexing the rotary valve with the counter-rotating cylinder cluster and crankshaft having gear types and relative sizes selected to yield small overall diameter;

Figure 8 is a simplified cross section of an embodiment of a fixed differential gear set for indexing the rotary valve with the counter-rotating cylinder cluster and crankshaft having gear types and relative sizes selected to yield small overall axial length as is integrated into the engine of Figure 3;

Figure 9 is a simplified cross section of an embodiment of a fixed differential gear set for indexing the rotary valve with the counter-rotating cylinder cluster and crankshaft having both gears in the planet gear assembly as external spur gears;

Figure 10 is a simplified cross section of an embodiment of a fixed differential gear set for indexing the rotary valve with the counter-rotating cylinder cluster and crankshaft having both gears in the planet gear assembly as internal annular gears;

Figure 11 is a simplified cross section of an embodiment of a fixed differential gear set for indexing the rotary valve with the co-rotating cylinder cluster and crankshaft having gear types and relative sizes selected to yield small overall diameter as integrated into the engine of Figure 2; • - Figure 12 is a simplified cross section of an embodiment of a fixed differential gear set for indexing the rotary valve with the co-rotating cylinder cluster and crankshaft having gear types and relative sizes selected to yield small overall axial length;

Figure 13 is a simplified cross section of an embodiment of a fixed differential gear set for indexing the rotary valve with the co-rotating cylinder cluster and crankshaft having both gears in the planet gear assembly as external spur gears;

Figure 14 is a simplified cross section of an embodiment of a fixed differential gear set for indexing the rotary valve with the co-rotating cylinder cluster and crankshaft having both gears in the planet gear assembly as internal annular gears;

Figure 15 is a simplified cross section of an embodiment of a planocentric gear set for indexing the rotary valve with the counter-rotating cylinder cluster and crankshaft having the rotation of the planet gear constrained by intermediate bodies rotating synchronously with but separately to the crankshaft as is integrated into the engine of Figure 1 ;

Figure 15a is a perspective view of the planocentric gear set of Figure 15; Figure 15b is a perspective view of a planocentric gear set similar to that of Figures 15 and 15a but different in that only two intermediate bodies, arrayed non-symmetrically around the axes of the gears are used to constrain the rotation of the planet gear, and the mounting points for the bearings in which the rotational constraint intermediate bodies rotate are displaced outwards radially beyond the peripheral radius of the gear teeth;

• Figure 16 is a simplified cross section of an embodiment of a planocentric gear set for indexing the rotary valve with the co-rotating cylinder cluster and

crankshaft having the rotation of the planet gear constrained by rollers mounted off of the planet gear that roll in an orbital path around the inner surface of cylindrical holes connected to the ported means (rotary valve head);

Figure 16a is a perspective view of the planocentric gear set of Figure 16; Figure 16b is a perspective view of a planocentric gear set similar to that of Figures 16 and 16a but different in that only three rollers displaced outwards radially beyond the peripheral radius of the gear teeth are used to constrain the rotation of the planet gear;

Figure 17 is a simplified cross section of an embodiment of a planocentric gear set for indexing the rotary valve with the counter-rotating cylinder cluster and crankshaft having an internally toothed planet gear with the rotation of the planet gear constrained by rollers mounted off of the ported means (rotary valve head) that roll in an orbital path around the inner surface of cylindrical holes in the planet gear; Figure 18 is a simplified cross section of an embodiment of a planocentric gear set for indexing the rotary valve with the co-rotating cylinder cluster and crankshaft having an internally toothed planet gear with the rotation of the planet gear constrained by intermediate bodies rotating synchronously with but ^' separately to the crankshaft; Figure 19 is a simplified cross section in a plane perpendicular to the crankshaft axis of rotation of a positive displacement gear fluid pump utilising gears that also form part of either a fixed differential or planocentric gear set as shown in Figures 7 to 18 for driving the engine port indexing;

Figure 20 is a simplified cross section of some of the components of an axial piston machine similar to that shown in Figure 1 , illustrating some of the inertial forces imposed upon the engine by the reciprocating components of the engine, and the resultant moment that rotates synchronously with the crankshaft about the crankshaft axis;

Figure 21 shows how the centrifugal forces on two masses separated by a distance along the crankshaft axis and with their centres of mass on opposite

sides of the crankshaft axis, combine to create a moment that rotates synchronously with the crankshaft;

Figure 22 shows a simplified cross sectioned perspective view of a crankshaft for an axial piston machine similar to that shown in Figure 1 showing two possible locations for planet gear eccentric journals or balance masses and their resultant out-of-balance centrifugal forces, along with the moment that they would together create. .

Reference will herein predominantly be made to a four-stroke version of an axial piston engine however it will be appreciated by a person skilled in the art how what is now described can be applied to both two-stroke axial piston engines

(or other stroke) and pumps.

Figure 1 is in essence a simplified cross sectional drawing of a preferred form of a four-stroke version of the engine having a counter-rotating crankshaft

20 and cylinder cluster 6. While the four-stroke engine by necessity has an odd number of cylinders 4, one of the cylinders and its associated piston 2, connecting rod 10 and part of the reciprocator 26 are shown rotated into the plane of the cross section for clarity.

The engine of the present invention and with reference to Figure 1 consists of a crankshaft 20 supported along its length by multiple coaxial bearings 11, 13, 15, 17 that allow the crankshaft to rotate with respect to the cylinder cluster 6 and the ported means 16. In the preferred form the crankshaft 20 operates as an output shaft to the combustion engine. Disposed from, and either forming an integral part of, or securable to the crankshaft 20 is an oblique crank journal 24.

The crankshaft axis 18 and oblique crank axis 22 intersect at a point 28. Disposed from and rotatable on suitable bearings 19, 21 or the like about the oblique crank journal 24 is the reciprocator 26 which controls the reciprocating motion of the pistons 2 within the cylinders 4 of the cylinder cluster 6. The cylinders 4 with the cylinder cluster 6 are rotatable about the crankshaft axis 18 and are suitably mounted with the use of bearings 5, 1, 13 or the like to allow such rotation.

A ported means 16 (also herein referred to as the "rotary valve head") as part of the engine casing 3 is also provided and provides the inlet and outlet porting, and optionally injection and/or other means to allow the engine of the present invention to operate as a pump or in a two or four-stroke internal combustion spark or compression ignition mode (though the ported means 16 will be different depending upon whether the engine is two-stroke, four-stroke or acting as a pump). S.ealing between the ported means 16 and the combustion space of each cylinder 4 is provided by means of floating face seals 14 that rotate with the cylinder cluster 6 and possess a small range of movement normal to the face of the ported means 16, allowing any changes in the relative distance between the cylinder cluster and the ported means to be adapted to by the floating face seals 14. Such a face seal may be similar to those described in US 4,250,843. Together the cylinder cluster 6 with its floating face seals 14 rotatably sliding against the ported means 16 comprises the "rotary valve" for the engine or pump of the present invention.

Two stroke axial piston engines and pumps differ from four-stroke engines in that the transfer of fluids into and out of the cylinder occurs every cycle of piston reciprocation rather than every two cycles of piston reciprocation as for four-stroke engines. This makes it possible to utilise any number of cylinders for a two stroke axial piston engine (rather than just odd numbers as for a four-stroke), but for a rotary valved axial piston engine this requires either more ports in the ported means 16 or a higher rate of rotary valve rotation (relative rotational speed of the cylinder cluster 6 with respect to the ported means 16 as compared to the rotational speed of the crankshaft 20 with respect to the ported means 16) to achieve the required higher rate of porting events for a two-stroke engine or pump. The inlet and exhaust porting events for two-stroke engines and pumps also differ in their timing and duration with respect to piston reciprocation as compared to a four-stroke engine. This will generally result in the dimensions and spacing of ports within the ported means 16 being different for a two stroke engine or pump than it would be for a four-stroke engine.

The cylinder cluster 6 is supported in two bearings 5, 7 that are concentric to the crankshaft axis 18 and which are mounted off of the stationary engine casing 3 and ported means 16 respectively which together encase the rotating engine components. The wide spacing of these two cylinder cluster support bearings 5, 7 provides a stable basis for reacting the large instantaneous moments created by the cyclic fluid pressure forces above each piston 2 within its complementary cylinder 4. The engine illustrated in Figure 3 replaces cylinder cluster support bearing 5 with cylinder cluster support bearing 9 that runs upon the crankshaft 20 rather than being mounted in the engine casing 3 as for cylinder cluster support bearing 5.

A reciprocator rotational constraint means, comprised of a bevel gear 30 mounted on or part of the reciprocator 26 and an identical bevel gear 32 mounted off of the cylinder cluster 6 is provided. These gears engage with each other and restrict the relative rotation of the reciprocator 26 and cylinder cluster 6 such that the reciprocator 26 rotates at the same angular rate about the oblique crank journal 24 as does the cylinder cluster 6 about the crankshaft 20. It also provides the required torque reaction for the output torque of the crankshaft 20 that would otherwise have to be reacted through lateral loads on the pistons 2 within their cylinders 4. This reciprocator rotational constraint means makes it possible for each piston 2 to use a connecting rod 10 having a pivoting joint 8 with multiple degrees of rotational freedom where it links to the piston 2 and another pivoting joint 12 with multiple rotational degrees of freedom where it links to the reciprocator 26. The reciprocator rotational constraint means (e.g. bevel gears 30, 32) ensures that the reciprocator end pivoting joint 12 and the piston end pivoting joint 8 do not significantly rotationally advance or retard to each other at any time. It is in other words a rotation synchronisation gear other forms of this mechanism may also be used such as described in our co-pending NZ patent application filed in August 2008.

. The engine or pump of the present invention is not necessarily restricted to using bevel gears 30, 32 to provide the reciprocator rotational constraint means,

indeed other mechanisms could be used to constrain the relative rotation of the reciprocator 26 and cylinder cluster 6 while allowing the nutating motion of the reciprocator with respect to the cylinder cluster. These include the Cardan Joint (also known as a Universal Joint) such as shown in US Patent 4,491,057, and the Rzeppa Joint (sometimes known as a Ball and Groove Joint) such as shown in US Patent 5,129,752.

The engine or pump of the present invention is not necessarily restricted to. the cylinder axis of each cylinder being parallel. Indeed these cylinders may have an axis at an oblique angle to the crankshaft axis 18. Furthermore the ported means need not have a flat surface upon which the floating face seals 14 run, other surfaces of revolution about the crankshaft axis 18 are also possible. Additionally, and although we have herein described an engine operating in a single sided mode, the engine of the present invention can be utilised so as to have two clusters of pistons 2 and cylinders 4 sharing a common crankshaft 20 and crankshaft axis 18 and arranged to act in substantially opposite directions and in respect of different ported means 16.

The reciprocator rotational restraint means 30, 32 creates a one-to-one indexing between the rotation of the cylinder cluster and the reciprocator. However an indexing is further required between the cylinder cluster 6 and the ported means 16. Without any such additional indexing, a rotation of for example the crankshaft 20 about the crank axis 18 could occur without such rotation inducing a rotation of the reciprocator 26 about the oblique crank journal 24. Such rotation would in such an instance merely drag the cylinder cluster around and would not establish a reciprocating motion for each piston 2 within its complementary cylinder 4. US Patent 5,094,195 describes a reciprocator rotational constraint means comprising of identical bevel gears. However in addition to the use of such bevel gears for ensuring that no relative rotation of the reciprocator 26 and the cylinder cluster 6 occurs, in the engine of the current invention further gearing is required to control the relative rotation of the cylinder cluster 6 and the crankshaft 20. Where reference herein is made to

gearing of the ported means or of the cylinder cluster it is to be appreciated that such gearing is preferably directly engaged to the ported means or cylinder cluster or an extension thereof.

Alternatively however, the gearing may by further intermediate gear stages be coupled with said ported means or cylinder cluster.

A first example of the port indexing gearing is shown in Figure 1 and consists of a planocentric gear set incorporating an externally toothed planet gear 46 and an internally toothed (annular) gear 48 to control the relative rotation and indexing of the cylinders 4, the ported means 16 and the crankshaft 20. The equivalent configuration of planocentric gear set shown in Figure 1 is also shown in isolated cross sectional form in Figure 15 and also in Figure 15a.

In a preferred planocentric gear set as shown in Figure 15, an externally toothed planet gear 46, is disposed from and rotatable (e.g. by use of a suitable bearing or the like) about the planocentric eccentric journal 44 on the crankshaft 20. The planet gear 46 engages with an internally toothed annular gear 48 connected to the ported means 16 and concentric to the crankshaft axis 18. The planet gear 46 is constrained against rotation with respect to the cylinder cluster 6 by two or more intermediate bodies 50 that rotate synchronously with the crankshaft. Each of the intermediate bodies 50 has one journal 52 that engages in a bearing in the planet gear 46 and one journal 54 that engages in a bearing mounted in an extension to the cylinder cluster 6. Both journals 52, 54 have axes parallel to the- crankshaft axis 18. The axes of journals 52, 54 are separated by exactly the same distance as the separation between the axis of the planocentric eccentric journal 44 and the crankshaft axis 18. This is illustrated in Figure 15 and Figure 15a by the dimension labelled as X.

The intermediate bodies 50 are arrayed around the crankshaft axis 18 in a plane perpendicular to the crankshaft axis 18 and at some distance from the crankshaft axis 18 such that the included angle between at least one selection of two of the intermediate bodies 50 is other than 180 degrees and preferably close to 90 degrees so that at any time at least one of the intermediate bodies 50 is

capable of reacting the torque applied to the planet gear 46. Ideally only two instances of intermediate bodies 50 are required to fully constrain the rotation of the planet gear; however more may be employed to retain rotational symmetry. The intermediate bodies 50 are preferably located at as great a distance from the crankshaft axis 18 as is feasible in order to reduce the bearing loads on the intermediate body journals 52 and 54 and the planocentric eccentric journal 44.

For the engine illustrated in Figurel, assuming that it is a five cylinder four-stroke the crankshaft 20 will rotate at five times the speed of the cylinder cluster 6 in the opposite direction. It will have three sets of inlet and outlet ports equally spaced around the ported means 16 with appropriate port geometry to achieve the desired fluid transfer characteristics. To create this desired indexing ratio the planocentric gear set employed must have a planet gear 46 with 5N teeth and an annular gear 48 with 6N teeth where N is a positive integer. To change the five cylinder four-stroke engine of Figure 1 into a two stroke engine or pump with the same indexing drive planocentric gear ratio as for the four-stroke would require twice as many sets of inlet and outlet ports (six sets rather than three sets) equally spaced about the ported means 16 with appropriate port geometry to achieve the desired fluid transfer characteristics.

The location of the planocentric gear set in the engine of Figure 1 is the most preferred location for the port indexing gearing when there is sufficient space between the cylinders 4 of the cylinder cluster 6 (yet having close cylinder spacing) to permit a sufficiently robust planocentric set or other form of port indexing gearing. Locating the port indexing gearing between the cylinders should generally result in a more compact engine with reduced externally apparent gear noise, because the gearing is surrounded by significant mass providing objects, such as the cylinder cluster 6 and bevel gear 32, to attenuate noise from the gearing.

Owing to the similar sizes of gears and the high number of engaged teeth within the planocentric gearing, the planocentric port indexing gear set is likely to be more compact than a suitable epicyclic port indexing gear set. This may

- λλ —

allow for the port indexing gearing to be placed in more beneficial locations within the engine or pump and may also help to reduce the overall mass and size of the engine.

Figure 2 shows another preferred form of port indexing gearing in an engine or pump of the current invention. The engine or pump is the same as that of Figure 1 except in that it has a different form and location of port indexing gearing.

The port indexing gearing of Figure 2 consists of a fixed differential gear set incorporating two gear sets comprised of four gears in total to control the relative rotation and indexing of the cylinder cluster 6, the ported means 16 and the crankshaft 20. The equivalent configuration of fixed differential gear set shown in Figure 2 is also shown in isolated cross-sectional form in Figure 11.

With reference to Figure 2, in the fixed differential gear set a planet gear assembly 33 comprised of two coaxially connected gears; an internally toothed (annular) gear 38, and an externally toothed (spur) gear 36, is disposed from and rotatable (e.g. by use of a suitable bearing or the like) about the eccentric journal 42 on the crankshaft 20. The internally toothed gear 38 of the planet gear assembly engages with an externally toothed gear 40 concentric to the crankshaft axis 18 and connected to the ported means 16. The externally toothed gear 36 of the planet gear assembly engages with an internally toothed gear 34 connected to the cylinder cluster 6 and concentric to and rotatable about the crankshaft axis 18. The two sets of meshing gears 34, 36 and 38, 40 in this fixed differential gear set both feature high numbers of teeth engaged at any time, lowering the individual tooth loadings and allowing relatively smaller gears to be used. The relative sizes of the planet gear assembly gears 36 and 38 in this embodiment are selected to create a radially compact implementation of the fixed differential gear set. This has the advantage of allowing placement of the gears at locations to take advantage of providing gear noise attenuation as is herein described.

The configuration of fixed differential gear set in Figure 2 results in the less preferred co-rotating engine configuration in which the cylinder cluster 6

rotates in the same direction as the crankshaft 20. Co-rotating axial piston machines are less preferred for several reasons; for a co-rotating machine the total angular momentum of the engine is the sum of the crankshaft 20 angular momentum and the components that rotate synchronously with the cylinder cluster 6, whereas for counter-rotating it is the difference between these two values, resulting in much lower total angular momentum for the counter-rotating

■ machine. A lower angular momentum is desirable as it reduces the torque reaction imposed on whatever body the axial piston machine is mounted within when the axial piston machine is accelerated or rotated. Co-rotation also increases the speed of the crankshaft 20 with respect to the ported means 16, with a probable resultant increase in bearing friction losses, and it can increase rotary valve sliding speeds as less ports are required (two sets of ports for a five cylinder four-stroke co-rotating engine versus three sets for a five cylinder fours- stroke contra-rotating engine), with potentially detrimental effects upon seal life and friction. This also applies not just to fixed differential but also to the planocentrically geared version of the engine.

The indexing drives for the engines and pumps of the current invention require precise indexing gear ratios between the crankshaft 20 and the cylinder cluster 6 rotations in order to control the indexing of the rotary valve. For a four- stroke engine of the current invention the ratio between the rotation speed of the cylinder cluster 6 and the crankshaft for both co-rotating and counter-rotating engines will generally need to be 1 :M where M is the number of cylinders 4 in the cylinder cluster.

Planocentric gear sets such as described with reference to Figure 1, the indexing drive gear ratio is created by what is in effect a single gear stage, and as such the number of gear teeth on each of the gears in the gear set must be some multiple of the ratio dictated by the required port indexing drive ratio. In the example of the five-cylinder counter-rotating four-stroke engine of Figure 1, the planet gear 46 must have 5N teeth and the annular gear 48 must have 6N teeth where N is a positive integer. Given that most practical designs of externally

toothed gears will have greater than 20 teeth it becomes necessary to choose a value of N greater than 1 for this planocentric gear set, and if N is greater than 1 each tooth will only come into cyclic contact with a subset of the teeth on the meshing gear, possibly increasing the harmonic noise and wear of the gear set. This will in general be the case for similar engines having less than 20 cylinders, which accounts for almost all likely designs.

Fixed differential gear sets such as described with reference to Figures 2, 3, 7, 8, 9, 10, 11, 12, 13 and 14 employ two sets of meshing gears to achieve the necessary port indexing gear ratio and this makes it possible to choose hunting tooth ratios for each of the two sets of meshing gears in the gear set that in combination create the desired indexing drive gear ratio.

For the four-stroke co-rotating engine of Figure 2, assuming that it has five cylinders, the crankshaft 20 will rotate at five times the speed of the cylinder cluster 6 in the same direction. It will have two sets of inlet and outlet ports equally spaced around the ported means 16 with appropriate port geometry to achieve the desired fluid transfer characteristics. To create this desired indexing ratio the fixed differential gear set employed must have the product of the number of gear teeth of planet spur gear 36 times the number of gear teeth of ported means spur gear 40 divided by the product of annular planet gear 38 times the number of teeth of annular cylinder cluster gear 34 equal to 0.8, (gear 36*gear 40)/(gear 38*gear 34)=0.8, as can be easily confirmed by those skilled in the art. An example of an appropriate choice of gear tooth numbers that satisfy this relationship while also giving desirable relative gear sizes for a compact design and meshing pairs of gears and in which each tooth will come successively into contact with every tooth on the meshing gear so as to produce even wear and reduced harmonic noise is: gear 34, 71 teeth, gear 36, 64 teeth, gear 38, 80 teeth and gear 40, 71 teeth. In this example the module of gears 34, 36 is greater than that of gears 38, 40 by a factor of 1.28 = (80-71)/(71-64), as is required in order for the two meshing sets of gears 34, 36 and 38, 40 to have planet gears 36, 38 as part of the same planet gear assembly 33 rotating on the

fixed differential eccentric 42. The gears in each meshing pair of gears 34, 36 and 38, 40 have tooth numbers that are not divisible by any common factor other than one - in other words; each of the meshing pairs of gears 34, 36 and 38, 40 have tooth numbers with prime factorisations that share no common factors. The use of a spur gear 38 and an internally toothed gear 36 on the planet gear assembly 33 is advantageous as it means that both of the gear sets contribute to the overall desired ratio. In the example of the previous paragraph the ratios for the two gear sets are 64/71 and 71/80, which multiply together to yield an overall ratio of 0.8. If two spur gears or two annular gears were used in place of the combination of a spur gear 36 and an annular gear 38 to create the planet gear assembly 33 of the fixed differential then one of the gear ratios would be greater than 1 and the other would have to be much less than 1 in order to yield the desired ratio of 0.8, this would necessitate a fixed differential gear set in which the two planet gears have much different diameters as shown in Figures 13, 14, which can be disadvantageous for overall engine packaging owing to the likely greater overall length of the planet gear assembly.

The similar sizes of the gears and the high number of engaged teeth within the fixed differential gearing of Figure 2 means that the fixed differential port indexing gear set is likely to be more compact than a suitable epicyclic port indexing gear set. This may allow for the fixed differential port indexing gearing to be placed in more beneficial locations within the engine or pump and may also help to reduce the overall mass and size of the engine.

To change the five cylinder four-stroke engine of Figure 2 into a two stroke engine or pump with the same indexing drive fixed differential gear ratio as for the four-stroke would require twice as many sets of inlet and outlet ports (four sets rather than two sets) equally spaced about the ported means 16 with appropriate port geometry to achieve the desired fluid transfer characteristics.

The location of the fixed differential gear set between the cylinder cluster 6 and the reciprocator 26 in the engine of Figure 2 is a less preferred location for the port indexing gearing that may be necessitated by there being insufficient

space to implement sufficiently robust port indexing gearing in the most preferred location between the cylinders 4. However; if connecting rods 10 of greater than the minimum length required to fit the reciprocator, cylinders and pistons 2 together are used, then piston side forces and overall engine friction may be reduced due to the reduced average angle between the longitudinal axes of the connecting rods and the axis of piston reciprocation. This reduced angle can in turn result in . reduced side forces on the piston 2 due to the better alignment of gas pressure and piston inertial forces with the axis of the connecting rod 10. In making the connecting rods longer to reduce piston side forces, a space between the cylinder cluster 6 and the reciprocator 26 may be created in which the port indexing gearing can be located without otherwise increasing the overall dimensions of the engine. In other words, longer connecting rods can reduce piston side forces and hence friction (as is also the case for conventional engines) due to the reduction in the angle between the conrod and the axis of the cylinders. Longer connecting rods may also usefully open up a space suitable for the gearing. This location of the gearing can allow for the gear noise attenuation as described to be utilised.

Figure 3 shows another preferred form of port indexing gearing in an engine of the current invention. The engine is the same as that of Figure 1 and Figure 2 except in that it has a different form and location of fixed differential port indexing gearing.

The port indexing gearing of Figure 3 consists of a fixed differential gear set incorporating four gears in total to control the relative rotation and indexing of the cylinders 4, the ported means 16 and the crankshaft 20. The equivalent configuration of fixed differential gear set shown in Figure 3 is also shown in isolated cross-sectional form in Figure 8.

In the fixed differential gear set of Figure 3 a planet gear assembly 133 comprised of two coaxially connected gears, one externally toothed (spur) gear 138 and one internally toothed (annular) gear 136, is disposed from and rotatable on a suitable bearing or the like about the eccentric journal 142 on the crankshaft

20. The externally toothed gear 138 of the planet gear assembly engages with an internally toothed gear 140 connected to the ported means 16 by means of the engine casing 3 and is concentric to the crankshaft axis 18, while the internally toothed gear 136 of the planet gear assembly engages with an externally toothed gear 134 connected to the cylinder cluster 6 and concentric to and rotatable about the crankshaft axis 18. This configuration of fixed differential gear set results in the preferred counter-rotating engine configuration in which the cylinder cluster 6 rotates in the opposite direction to the crankshaft 20. The two sets of meshing gears 134, 136 and 138, 140 in this fixed differential gear set both feature high numbers of teeth engaged at any time, lowering the individual tooth loadings and allowing relatively smaller gears to be used. The relative sizes of the planet gear assembly gears 136 and 138 are selected to allow an axially compact implementation of the fixed differential gear set that has a small impact on engine length (i.e. in the axial direction). By way of example of an appropriate selection of tooth numbers for this fixed differential gear set of Figure 3; assuming that it has five cylinders, the crankshaft 20 will rotate at five times the speed of the cylinder cluster 6 in the opposite direction. It will have three sets of inlet and outlet ports equally spaced around the ported means 16 with appropriate port geometry to achieve the desired fluid transfer characteristics. To create this desired indexing ratio the fixed differential gear set employed must have the product of the number of gear teeth of planet annular gear 136 and the number of gear teeth of ported means annular gear 140 divided by the product of the number of teeth of planet spur gear ^" 138 and the number of teeth of cylinder cluster spur gear 134 equal to 1.2, (gear 136*gear 140)/(gear 138*gear 134)=1.2, as can be confirmed using well known gear design principles. An example of an appropriate choice of gear tooth numbers that satisfy this relationship while also giving desirable relative gear sizes for a compact design and meshing pairs of gears in which each tooth will come successively into contact with every tooth on the meshing gear so as to produce even wear and reduced harmonic noise is: gear 134, 105 teeth, gear 136,

116 teeth, gear 138, 58 teeth and gear 140, 63 teeth. In this example the module of gears 134, 136 is less than that of gears 138, 140 by a factor of 0.45 = (63- 58)/( 116-105), as is required in order for the two meshing sets of gears 134, 136 and 138, 140 to have planet gears 136, 138 as part of the same planet gear assembly 133 rotating on the fixed differential eccentric 142. The gears in each meshing pair of gears 134, 136 and 138, 140 have tooth numbers that are not divisible by any common factor other than one - in other words; each of the meshing pairs of gears 34, 36 and 38, 40 have tooth numbers with prime factorisations that share no common factors. To change the five cylinder four-stroke engine of Figure 3 into a two stroke engine or pump with the same indexing drive fixed differential gear ratio as for the four-stroke would require twice as many sets of inlet and outlet ports (four sets rather than two sets) equally spaced about the ported means 16 with appropriate port geometry to achieve the desired fluid transfer characteristics. The location of the fixed differential gear set on the far side of the reciprocator 26 from the cylinder cluster 6 in the engine of Figure 3 is a less preferred position for the port indexing gearing that increases the overall length of the engine, but it may be necessitated by there being insufficient space to implement sufficiently robust port indexing gearing in the most preferred location between the cylinders 4. However; this location on the far side of the reciprocator from the cylinder cluster has an advantage in that it simplifies engine design and assembly and does not require the transmission of torque from the ported means 16 through the middle of the cylinder cluster 6 as for the indexing drive locations of Figure 1 and Figure 2. This permits more rigid and lightweight transmission of torque directly from the engine casing 3 (as at gear 140) attached to the ported means 16.

Figure 4 shows a simplified cross section taken perpendicular to the crankshaft axis 18 of the cylinder cluster 6 of a compactly designed three cylinder engine or pump illustrating the very small space 56 available between the cylinders 4 for integrating port indexing gearing components or torque

transmitting means such as the crankshaft 20 or an extension from ported means 16. This is likely to necessitate the port indexing gearing being located on the far side of the reciprocator 26 from the cylinder cluster as depicted in Figure 3.

Figure 5 shows a simplified cross section taken perpendicular to the crankshaft axis 18 of the cylinder cluster 6 of a compactly designed five cylinder engine or pump illustrating the relatively small space 56 available between the cylinders 4 for integrating , port indexing gearing components or torque transmitting means such as the crankshaft 20 or an extension from ported means 16 into the preferred location between the cylinders. This is likely to lead to, or allow for, the port indexing gearing being located between the cylinder cluster and reciprocator 26 as depicted in Figure 2 or on the far side of the reciprocator from the cylinder cluster as depicted in Figure 3 or on the side of the cylinder cluster furthest from the reciprocator.

Figure 6 shows a simplified cross section taken perpendicular to the crankshaft axis 18 of the cylinder cluster 6 of a compactly designed seven cylinder engine or pump illustrating the relatively large space 56 available between the cylinders 4 that should make it possible to integrate port indexing gearing components or torque transmitting means such as the crankshaft 20 or an extension from ported means 16 into the preferred location between the cylinders as depicted in Figure 1.

Several variations to the planocentric and fixed differential gearing will now be described with reference to Figures 7 to 18.

Figure 7 shows, in isolated cross-sectional form, a fixed differential gear set similar to that in Figure 3 with the preferred combination of an internally toothed gear 136 and an externally toothed gear 138 on the planet gear assembly 133 that helps to produce a more compact gear set. The internally toothed gear 140 attached to the ported means 16 and the externally toothed gear 134 attached to the cylinder cluster 6 means that the engine will have the preferred counter- rotating cylinder cluster 6 and crankshaft 20. The relative gear sizes of the planet gear assembly gears 136 and 138 are selected to yield a radially compact design.

Figure 8 shows, in isolated cross-sectional form, a fixed differential gear set similar to that in Figure 3 and Figure 7 with the preferred combination of an internally toothed gear 136 and an externally toothed gear 138 on the planet gear assembly 133 that helps to produce a more compact gear set. The internally toothed gear 140 attached to the ported means 16 and the externally toothed gear 134 attached to the cylinder cluster 6 means that the engine will have the preferred counter-rotating cylinder cluster 6 and crankshaft 20. The relative gear sizes of the planet gear assembly gears 136 and 138 are selected to yield an axially compact design. Figure 9 shows, in isolated cross-sectional form, a fixed differential gear set similar to that in Figure 3 but with the less preferred combination of two externally toothed gears 236 and 238 on the planet gear assembly 233 that increases the size of the gear set for a given torque, but may have some manufacturing advantages. The relative gear pitch diameters of gears 236 and 238 are selected to give the engine the preferred counter-rotating cylinder cluster 6 and crankshaft 20. If planet gear 238 is instead made larger than planet gear 236 with corresponding changes in the gears with which they mesh then the engine would have the less preferred co-rotating cylinder cluster 6 and crankshaft 20 as is depicted in Figure 13. Figure 10 shows, in isolated cross-sectional form, a fixed differential gear set similar to that in Figure 3 but with the less preferred combination of two internally toothed gears 336 and 338 on the planet gear assembly 333 that increases the size of the gear set for a given torque. The relative gear pitch diameters of gears 336 and 338 are selected to give the engine the preferred counter-rotating cylinder cluster 6 and crankshaft 20.

Figure 11 shows, in isolated cross-sectional form, a fixed differential gear set similar to that in Figure 2 with the preferred combination of an internally toothed gear 38 and an externally toothed gear 36 on the planet gear assembly 33 that helps to produce a more compact gear set. The externally toothed gear 40 attached to the ported means 16 and the internally toothed gear 34 attached to the

cylinder cluster 6 means that the engine will have the less preferred co-rotating cylinder cluster 6 and crankshaft 20. The relative gear sizes of the planet gear assembly gears 36 and 38 are selected to yield a radially compact design.

Figure 12 shows, in isolated cross-sectional form, a fixed differential gear set similar to that in Figure 2 and Figure 11 with the preferred combination of an internally toothed gear 36 and an externally toothed gear 38 on the planet gear assembly 33 that helps to produce a more compact gear set. The externally toothed gear 40 attached to the ported means 16 and the internally toothed gear 34 attached to the cylinder cluster 6 means that the engine will have the less preferred co-rotating cylinder cluster 6 and crankshaft 20. The relative gear sizes of the planet gear assembly gears 36 and 38 are selected to yield an axially compact design.

Figure 13 shows, in isolated cross-sectional form, a fixed differential gear set similar to that in Figure 2 but with the less preferred combination of two externally toothed gears 236 and 238 on the planet gear assembly 233 that increases the size of the gear set for a given torque, but may have some manufacturing advantages. The relative gear pitch diameters of gears 236 and 238 are selected to give the engine the less preferred co-rotating cylinder cluster 6 and crankshaft 20. Figure 14 shows, in isolated cross-sectional form, a fixed differential gear set similar to that in Figure 2 but with the less preferred combination of two internally toothed gears 336 and 338 on the planet gear assembly 333 that increases the size of the gear set for a given torque. The relative gear pitch diameters of gears 336 and 338 are selected to give the engine the less preferred co-rotating cylinder cluster 6 and crankshaft 20.

As a summary Table I below illustrates the various configurations of gearing utilising the fixed differential gear set.

TABLE I

Figure 15 shows, in isolated cross-sectional form, a planocentric gear set similar to that in the engine of Figure 1, with the preferred combination of an externally toothed planet gear 46, disposed from and rotatable (e.g. on a suitable bearing or the like) about the planocentric eccentric journal 44 on the crankshaft 20, and engaging with internally toothed gear 48. The internally toothed gear 48 is attached to the ported means 16 and is concentric with the crankshaft axis 18. Two or more intermediate bodies 50, each having two separate parallel journals, separated by a distance X equal to the eccentricity of the planet carrying journal 44 from the crankshaft axis 18. One journal 52 of each intermediate body is free

to rotate within a complementary bearing mounted to the externally toothed planet gear 46, the second intermediate body journal 54 of each intermediate body is free to rotate within a complementary bearing mounted on to the cylinder cluster 6. The intermediate bodies 50 rotate synchronously with, but separately to, the crankshaft 20 holding the planet gear irrotational with respect to the cylinder cluster 6, giving the engine the preferred configuration of counter- rotating cylinder cluster 6 and crankshaft 20.

The intermediate body 50 crank arrangement is advantageous in that owing to their ability to react loads in both tension and compression a minimum of only two such bodies are required to rotationally constrain the planet gear 46, while the rotational constraint roller 74 system of Figure 16 being reliant on contact pressure between the rollers 74 and their complementary cylindrical camming surfaces or holes 70 requires three or more rollers. The space required for the intermediate body rotational constraint system may also be smaller than for the rotational constraint roller system of Figure 16, owing to the lack of large diameter cylindrical camming surfaces 70.

Figure 15a is a simplified perspective view of the planocentric gear set configuration of Figure 15, and shows the large numbers of teeth engaged at any time in the planocentric gear set, which serves to reduce the individual tooth loading, thereby reducing the size of the gears required to transmit a given torque. For the same gear ratio and overall diameter the planocentric eccentric journal 44 of Figure 15 has greater eccentricity from the crankshaft axis 18 than the fixed differential eccentric journal 42 of Figure 7 leading to lower bearing forces for the same torque capacity. Figure 15b is a a simplified perspective view of a planocentric gear set similar to that of Figure 15 and 15a, but having the bearing mounting points for the rotational constraint intermediate bodies 50 displaced radially outwards to outside of the periphery of the gears 46, 48 of the planocentric gear set in order to reduce the forces in the intermediate bodies for a given level of torque transmission. This is explained further, below.

Figure 16 shows, in isolated cross-sectional form, a planocentric gear set similar to that in the engine of Figure 1, but different in terms of the method used to rotationally constrain the planet with respect to the ported means 16. The gear set has the preferred combination of an externally toothed planet gear 46, disposed from and rotatable (e.g. on a suitable bearing or the like) about the planocentric eccentric journal 44 on the crankshaft 20, and engaging with internally toothed gear 48. The internally toothed gear 48 is attached to the cylinder cluster 6 and rotates relative to and is concentric with the crankshaft axis 18. Three or more rollers 74 mounted off of the planet gear 46 orbit in rolling contact within complementary cylindrical camming surface 70 connected to the ported means 16.

The cylindrical camming surface 70 have a diameter equal to the outside diameter of the rollers 74 plus two times the eccentricity 'X' of the planet journal 44 from the crankshaft axis 18. The axis 68 of each cylindrical camming surface 70 is the same distance from the crankshaft axis 18 as is its complementary roller's axis 72 from the axis of the planet journal 44. The rollers 74 are located at the largest convenient distance from the axis of the planet gear 46 in order to reduce bearing forces for a given torque transmitted through the planocentric gear set. The rollers 74 are preferably disposed at equal angular spacings about the planet gear axis to aid in reducing any torsional irrigidity in the planocentric gear set and to reduce bearing and roller forces by keeping the forces on the rollers as much as possible in a tangential direction to the crankshaft axis 18. The rotational constraint rollers 74 orbiting in rolling contact within their complementary cylindrical camming surfaces 70 in the ported means 16 give the engine the less preferred configuration of co-rotating cylinder cluster and crankshaft.

Figure 16a shows a simplified perspective view of the planocentric gear configuration of Figure 16.

Figure 16b is a a simplified perspective view of a planocentric gear set very similar to that of Figure 16 and 16a, but having the mounting points for the

rotational constraint rollers 74 displaced radially outwards to outside of the periphery of the gears 46, 48 of the planocentric gear set in order to reduce the forces in the rollers for a given level of torque transmission. This is explained further, below. Figure 17 shows, in isolated cross-sectional form, a planocentric gear set similar to that in the engine of Figure 1 and the gear set of Figure 16, but utilising a roller-in-hole based planet gear rotational constraint system and having the less preferred combination of an internally toothed planet gear 146, disposed from and rotatable (e.g. on a suitable bearing or the like) about the planocentric eccentric journal 44 on the crankshaft 20, and engaging with externally toothed gear 148. The externally toothed gear is attached to the cylinder cluster 6 and rotates relative to and is concentric with the crankshaft axis 18.

Three or more rollers 174 mounted off of an extension from the ported means 16 orbit in rolling contact within complementary cylindrical camming surfaces 170 in the planet gear 146, said cylindrical camming surfaces 170 having a diameter equal to the outside diameter of the rollers 174 plus two times the eccentricity 'X' of the planet journal 44 from the crankshaft axis 18. This is a less preferred arrangement owing to the higher roller 174 forces created by the smaller distance of the roller axes 172 from the crankshaft axis 18 that is necessitated by the need to keep the diameter of the cylindrical camming surfaces 170 within the body of the planet gear 146 compared with having the rollers mounted off of the planet gear 146.

The axis 168 of each cylindrical camming surfaces 170 is the same distance from the planet journal's 44 axis as is its complementary roller's axis 172 from the crankshaft axis 18. The rollers 174 are located at the largest convenient distance from the crankshaft axis 18 in order to reduce bearing forces for a given torque transmitted through the planocentric gear set. The rollers 174 are preferably equi-spaced about the crankshaft axis to help reduce any torsional irrigidity in the planocentric gear set and reduce bearing and roller forces by keeping the forces on the rollers as much as possible in a tangential direction to

the crankshaft axis 18. The rotational constraint rollers 174 orbiting in rolling contact within their complementary cylindrical camming surfaces 170 in the planet gear 146 give the engine the preferred configuration of counter-rotating cylinder cluster 6 and crankshaft 20. Figure 18 shows, in isolated cross-sectional form, a planocentric gear set similar to that in the engine of Figure 1 and Figure 15, with the less preferred combination of an internally toothed planet gear 246, disposed from and rotatable (e.g. on a suitable bearing or the like) about the planocentric eccentric journal 44 on the crankshaft 20, and engaging with externally toothed gear 248 concentric with the crankshaft axis 18 and is attached to the ported means 16. The internally toothed planet gear 246 is less preferred because it is likely to be harder to manufacture, with greater axial length and mass which can in turn lead to higher loads and frictional losses for the planocentric planet eccentric journal 44. Two or more intermediate bodies 50, each having two separate parallel journals, one journal 52 of which is free to rotate within a complementary bearing mounted to the internally toothed planet gear 246, the second intermediate body journal 54 of which is free to rotate within a complementary bearing mounted on to the cylinder cluster 6. The intermediate bodies rotate synchronously with, but separately to the crankshaft holding the planet gear irrotational with respect to the cylinder cluster, giving the engine the less preferred configuration of co-rotating cylinder cluster and crankshaft.

The intermediate bodies providing the rotational constraint means for the planet gear of the planocentric gear sets of the present invention are not restricted to having eccentric journals axially separated as shown in Figures 1, 15, 15a, 15b, 18 and being in general form cranks. The separate parallel axes of rotation of the intermediate bodies could alternatively be nested one within the other with both bearings in the same axial plane but eccentric to each other as for the crank journal axes shown in Figures 1, 15, 18. This has the advantage of reducing cantilever forces on the intermediate body bearings, but requires one bearing to be much larger diameter than the other, even though they share similar loads.

For the planocentric rotational constraint intermediate bodies, it is preferable that a minimum of two instances of intermediate bodies are at 90 degrees about the crankshaft axis 18 and otherwise an odd number of three or more intermediate bodies equi-spaced about the crankshaft axis are provided to most advantageously react to the rotational constraint torque.

As a summary; Table II below illustrates the various configurations of gearing utilising the planocentric gear set in Figures 15 to 18. TABLE II

Cylinder

AttacheCTo , .Cluster/ ,

Attached To Ported

Figure Crankshaft ^"*

Meansl, ¹ ^ hψ

*.„ Gluster,T * ^s * Rotation ^! J

Orientation rotational internal annular to

15 constraint crank Counter-rotating external planet to planet rotational internal annular to

15a constraint crank Counter-rotating external planet to planet rotational internal annular to

15b constraint crank C ounter-rotating external planet to planet

rotational constraint internal annular

16 Co-rotating holes to planet to external

In general, as far as size reduction of gearing is concerned, in order to allow compact design and greater liberty in placement of the index gearing, a spur gear (planet gear) meshing inside an annulus gear where a greater number of teeth are engaged at any time than for a spur gear meshing against a spur gear will in general tend to result in a smaller overall gear package for the same

transmitted torque. This can offer potential space savings over both parallel layshafts and epicyclic gear sets for planocentric and fixed differential gearing. As a result significant freedom exists in the placing of the index gearing.

With the crankshaft extending through fixed differential gearing on the cylinder cluster 6 side of the reciprocator 26, a fixed differential gear set in which the fixed differential gear set is located on the cylinder cluster side of the reciprocator can be provided. The crankshaft 20 or an extension of the crankshaft extends axially to both sides of the planet carrying eccentric journal 42 of the fixed differential gear set. With the crankshaft extending through the planocentric gearing on the cylinder cluster 6 side of the reciprocator 26, a planocentric gear set can be • located on the cylinder cluster side of the reciprocator in which the crankshaft 20 or an extension of the crankshaft extends axially to both sides of the planet carrying eccentric journal 44 of the planocentric gear set. With the crankshaft extending through fixed differential gearing on the far side of the reciprocator 26 from the cylinder cluster 6 (in other words; the tail end of the engine or pump), a fixed differential gear set can be located on the side of the reciprocator furthest from the cylinder cluster in which the crankshaft 20 or an extension of the crankshaft extends axially to both sides of the planet assembly carrying eccentric journal 42 of the fixed differential gear set.

With the crankshaft extending through planocentric gearing at the tail end of the engine or pump, a planocentric gear set can be located on the side of the reciprocator furthest from the cylinder cluster 6 in which the crankshaft 20 or an extension of the crankshaft extends axially to both sides of the planet carrying eccentric journal 44 of the planocentric gear set.

For the fixed differential gearing located between the cylinders 4 of the cylinder cluster 6, the fixed differential gear set having four gears can be located either partially or wholly between the cylinders of said engine or pump, and is used to control the relative rotation of the crankshaft 20 and the rotary valve that

is formed by the sliding interface between the face seals 14 of the cylinder cluster 6 and the ported means 16.

For the planocentric gearing located between the cylinders 4 of the cylinder cluster 6, the planocentric gear set can be located either partially or wholly between the cylinders of said engine or pump, and is used to control the relative rotation of the crankshaft 20 and the rotary valve.

For the fixed differential gearing located between the cylinder cluster 6 and the reciprocator 26, the fixed differential gear set can be located either partially or wholly between the cylinder cluster and the reciprocator of said engine or pump, and is used to control the relative rotation of the crankshaft 20 and the rotary valve.

For the planocentric gearing located between the cylinder cluster 6 and the reciprocator 26, the planocentric gear set can be located either partially or wholly between the cylinder cluster and the reciprocator of said engine or pump, and is used to control the relative rotation of the crankshaft 20 and the rotary valve.

For the fixed differential gear set located on the far side of the cylinder cluster 6 from the reciprocator 26 (in other words; above the cylinder cluster), the fixed differential gear set can be located either partially or wholly above the cylinder cluster of said engine or pump, and is used to control the relative rotation of the crankshaft 20 and the rotary valve.

A planocentric gear set that provides the port indexing gearing may also be located above the cylinder cluster 6, either partially or wholly above the cylinder cluster 6 at the end of the cylinder cluster furthest from the reciprocator of said engine or pump, and is used to control the relative rotation of the crankshaft 20, cylinder cluster 6 and rotary valve.

A fixed differential gear set that provides the port indexing gearing for the engine or pump by controlling the relative rotation of the crankshaft 20 and the rotary valve may be located on the far side of the reciprocator 26 from the cylinder cluster 6.

Likewise a planocentric gear set that provides the port indexing gearing for the engine or pump by controlling the relative rotation of the crankshaft 20 and the rotary valve may be located on the far side of the reciprocator 26 from the cylinder cluster 6. As shown in Figures 15, 15a, 15b, 18 the bearing mounts for the rotational constraint intermediate body journals 52, 54 in either the planet gear 46, 246 or the assembly to which the rotation of the planet gear is restrained by the intermediate bodies 50 (in the case of Figures 15, 15a, 15b, 18 the cylinder cluster 6) are preferably made so that they allow a small amount of motion in a radial direction, depicted by the arrow labelled 'Y', (in other words; radial compliance) with respect to the rotational axis of the object or assembly that they mounted within (crankshaft axis 18 in the case of intermediate body journal 54 and the axis of the planet gear journal 44 in the case of intermediate body journal 52), in response to any applied radial force. The forces transmitted by the intermediate bodies 50 should predominantly be tangential with respect to the crankshaft axis 18 and the axis of the eccentric journal 44, though there will be a smaller radial force component. The radially compliant bearing mounts help to reduce any unnecessary large loads on the planocentric gear set's bearings created by small dimensional mismatches in the planocentric gear set's components arising from manufacturing inaccuracies or thermally or stress induced dimensional changes. As shown in Figures 15, 15a, 15b, 18 the radially compliant mounts for the intermediate bodies 50 are preferably not in the planet gear 46, 246, as the planet gear orbits at synchronous speed with the crankshaft 20, inducing large oscillating lateral accelerations within the planet gear that would tend to induce unwanted movements in a radially compliant bearing mount in the planet gear 46, 246.

Figure 15a shows a simplified perspective view of the planocentric gear set configuration of Figure 15 showing one possible method for implementing a radially compliant bearing mount through use of flexures 51 machined into the intermediate bodies' 50 bearing mount structure attached to the cylinder cluster

6. The bearing mount points in the cylinder cluster 6 are able to move in a radial direction depicted by the arrows labelled 'Y' with respect to the crankshaft axis 18 while remaining relatively rigid in the tangential direction to transmit the required constraining torque. Referring to Figures 16, 16a, and 17, and in a manner similar to that described above for the intermediate body rotational constraints of Figures 15, 15a, 18, preferably either the mounting of the rollers 74, 174 in the planet gear

46, 146 or their complementary holes or cylindrical camming surface 70, 170 within the body connected to the ported means 16 incorporate a small amount of radial compliance to prevent the rollers 74, 174 from being able to generate unnecessarily large loads in a radial direction with respect to the planet gear's axis (concentric with the planet gear eccentric journal 44). Such unnecessary forces could arise as a result of manufacturing inaccuracies or thermal or stress induced dimensional changes encountered during engine or pump operation. Alternatively, with reference to Figure 16a, the complementary holes or cylindrical camming surface 70 may be made very slightly elliptical with their major axis 'V passing through the crankshaft axis 18 while their minor axis 'U' is tangential to the crankshaft axis in order to help prevent unnecessary radial loading. Figure 15b shows a simplified perspective drawing of a planocentric gear set similar to that of Figures 15 and 15a, having the preferred counter-rotating cylinder cluster 6 and crankshaft 20 and in which the annular gear 48 is attached to the ported means 16. The mounting points for the bearings in which the journals 52, 54 of the rotational constraint intermediate bodies 50 rotate are displaced radially outwards from the planocentric eccentric journal 44 and the crankshaft axis 18 respectively. This is enabled by utilising planet arms 55, attached to the planet gear 46 and extending radially outwards beyond the periphery of the planet gear's teeth, and in this case arranged asymmetrically about the axis of the planet gear owing to there being only the minimum number of intermediate bodies (two) used to constrain the planet gear's rotation to the

cylinder cluster. Displacing the intermediate bodies' bearing mounts outwards radially beyond the periphery of the planet gear's teeth has the beneficial effect of reducing the forces that have to be reacted by the intermediate bodies in order for the planocentric gear set to transmit a given torque, this in turn reduces the forces on the planocentric eccentric journal 44 and should reduce the frictional losses for the planocentric gear set while also possibly allowing the size and mass of the intermediate bodies and their bearings to be reduced. Radially compliant bearing mounts for the rotational constraint intermediate bodies can be employed as for the planocentric gear set of Figures 15, 15a. The outwardly radially displaced rotational constraint intermediate bodies could also be similarly applied to the planocentric gear set of Figure 18.

Figure 16b shows a simplified perspective drawing of a planocentric gear set similar to that of Figures 16 and 16a, having the less preferred co-rotating cylinder cluster 6 and crankshaft 20 and in which the rotation of the planet gear 46 is constrained to the ported means 16 through the use of rotational constraint rollers 74 that roll in continuous orbital contact around the inner surface of complementary cylindrical holes 70 attached to the ported means. The annular gear 48 is attached to the cylinder cluster 6. The mounting points for the three (the' minimum number that will work satisfactorily) rotational constraint rollers are arrayed symmetrically about the planocentric eccentric journal 44 and are displaced radially outwards along planet arms 57 that are attached to the planet gear but extend radially outwards beyond the periphery of the planet gear's teeth. This has the beneficial effect of reducing the forces that have to be reacted by the rotational constraint rollers in order for the planocentric gear set to transmit a given torque, this reduces the forces on the planocentric eccentric journal 44 and should reduce the frictional losses for the planocentric gear set while also possibly allowing the size and weight of the rotational constraint rollers to be reduced. Radially compliant bearing mounts for the rotational constraint rollers as previously described can still be employed as for the planocentric gear set of

Figures 16, 16a. The outwardly radially displaced rotational constraint rollers could also be similarly applied to the planocentric gear set of Figure 17.

Figure 19 shows a cross-section perpendicular to the engine crankshaft axis 18, of a gear pump formed from two gears, being part of either a planocentric gear set or a fixed differential gear set that acts as the port indexing gearing for an axial piston engine or pump. The externally toothed planet gear 64 is disposed from and rotatable on a suitable bearing or the like about a fixed differential eccentric journal 42 or a planocentric eccentric journal 44 on the crankshaft 20 of the engine. The externally toothed planet gear 64 engages with, and rotates relative to, the internally toothed gear 62. The two gears have the same axial length and are both sandwiched between two axially separated sealing surfaces such that a sealed cavity exists between the gears. A separating block or piston 66, attached to the crankshaft 20 and having a small clearance between itself and the teeth of gear 62 and gear 64 divides the sealed cavity into two parts, separating the inlet port 60 from the outlet port 58 of the gear pump which are also part of the crankshaft 20. The teeth of the gears pump fluid from the inlet port to the outlet port as they move past the surfaces of the separating block.

Axial piston machines incorporating reciprocators can generally be dynamically balanced to a high standard as the out of balance forces and couples created by the rotating and reciprocating components are predominantly of the same frequency as the crankshaft rotation and the higher frequency out of balance forces and moments that are present in many conventional in-line cylinder engines are either absent or of generally reduced magnitude. Figure 20 is a simplified cross section of some of the key components of an axial piston machine illustrating the inertial forces FAl, FA2 imposed upon the engine by the reciprocating motion of the pistons 2, connecting rods 10 and reciprocator 26 that combine to yield a net moment Ml acting upon the machine. While this moment Ml is shown as being on the centreline of the engine, moments are by their nature non-localised. Due to the symmetry of the cylinder layout for engines having three or more cylinders 4, the net moment Ml will rotate synchronously

with the crankshaft 20. To reduce vibration it is useful to balance this rotating moment Ml with another rotating moment of opposite orientation.

Balancing of an axial piston machine is generally accomplished by the addition of appropriate masses to the crankshaft 20 at suitable radii and locations along the length of the crankshaft. Figure 21 shows how the centrifugal forces FRl, FR2 on two masses 80, 82 on opposite sides of the crankshaft axis 18 and separated by a distance along the crankshaft axis 18 can combine to create a moment M2 that rotates synchronously with the crankshaft 20. By appropriate sizing of these masses 80 and 82, their radii from the crankshaft axis 18, and the distance separating them along the crankshaft axis 18, the dynamic unbalance of the axial piston machine as illustrated in Figure 20 can potentially be greatly reduced.

It is normally advantageous to minimise the size of the balance masses required in order to reduce engine weight as the balance masses often serve no other useful function. This can be facilitated by situating balance masses as far from the axis of rotation as possible within other design constraints as the centrifugal force created by a balance mass is proportional to the distance of its centre of mass from the crankshaft axis. In the case of creating a moment using two balance masses on opposite sides of the crankshaft axis 18 as shown in Figure 21, the greater the axial separation, the greater will be the moment created. Towards this end it is useful to have one out of balance mass on either side of the oblique crank section 24 of the crankshaft 20 in order to increase their axial separation and hopefully reduce the mass required to balance the moment Ml created by the reciprocating components of the engine or pump. Figure 22 illustrates a simplified cross sectioned perspective view of a crankshaft 20 for an axial piston machine showing two possible locations 84, 86 for eccentric journals or balance masses on either side of the oblique crank journal 24 of the crankshaft 20, either of which could be suitable for mounting eccentric planet gears 46, 146, 246 or planet gear assemblies 33, 133, 233, 333 of a planocentric or fixed differential gear set for port indexing gearing. The offset

from the crankshaft axis 18 of the eccentric journals at these two locations 84, 86 is depicted such that the out of balance forces FR3 and FR4 respectively, produced by the out of balance mass or eccentric journal and its associated planet gear or planet gear assembly, would produce a force couple M3 of opposite orientation to that created by the inertia of the reciprocating components of the axial piston machine as shown in Figure 20, and thus should enable the dynamic balancing of the axial piston machine to be achieved with the use of less dedicated balancing mass. Selection of other orientations of planet carrying eccentric journals with respect to the crankshaft axis 18 and the oblique crank axis 22 would in general be less advantageous for engine balancing.

Hence, preferably the angular positioning of said planet gear 46, 146 or 246 or planet gear assembly 33, 133, 233 or 333 carrying eccentric journal 42, 44 on said crankshaft 20 about said crankshaft axis 18 with relation to said oblique crank axis 22 results in the mass of said planet gear or planet gear assembly contributing to the dynamic balancing of the inertial loads created by the motion of said reciprocator and/or said pistons.

The engine of Figures 1 shows a planocentric gear set situated between the cylinders 4 of the cylinder cluster 6. The planocentric eccentric journal 44 on the crankshaft 20 is oriented in a plane perpendicular to the crankshaft axis 18 on the crankshaft 20 with respect to the oblique crank axis 22 such that the centrifugal force created by the planocentric eccentric journal 44 and the planet gear 46 and the intermediate bodies 50 as the crankshaft rotates about its axis 18 in conjunction with a balancing mass on the opposite side of the crankshaft axis 18 on the other side of the oblique crank journal 24 should contribute to the dynamic balancing of the inertial forces produced by the reciprocating components of the engine including the reciprocator 26, the connecting rods 10 and the pistons 2 as shown in Figure 20.

The engine of Figure 2 shows a fixed differential port indexing gear set located between the cylinder cluster 6 and the reciprocator 26. The fixed differential eccentric journal 42 on the crankshaft 20 is oriented in a plane

perpendicular to the crankshaft axis 18 on the crankshaft 20 with respect to the oblique crank axis 22 such that the centrifugal force created by both the fixed differential eccentric journal 42 and the fixed differential planet gear assembly 33 as the crankshaft 20 rotates about its axis 18 in conjunction with a balancing mass on the opposite side of the crankshaft axis on the other side of the oblique crank journal 24 should contribute to the dynamic balancing of the inertial forces produced by the oblique crank journal 24 and the reciprocating components of the engine including the reciprocator 26, the connecting rods 10 and the pistons 2. For the fixed differential port gearing of Figure 3, the fixed differential eccentric journal 42 on the crankshaft 20 is oriented in a plane perpendicular to the crankshaft axis 18 on the crankshaft 20 with respect to the oblique crank axis 22 such that the centrifugal force created by both the fixed differential eccentric journal 42 and the fixed differential planet gear assembly 133 as the crankshaft rotates about its axis 18 in conjunction with the centrifugal force created by the out of balance centre of mass of the oblique crank journal 24, reciprocator 26 and reciprocator bearings 19, 21 contributes to the dynamic balancing of the inertial forces produced by the oblique crank journal 24 and the reciprocating components of the engine including the reciprocator 26, the connecting rods 10 and the pistons 2 as shown in Figure 20.

In general, the planet gear carrying eccentric journal 42 or planet gear assembly carrying eccentric journal 44 will be most advantageously situated for balancing the inertial couple created by the reciprocating components and oblique crank journal 24 of the axial piston machine if its centre of mass is situated as far as possible from the oblique crank journal's axis 22 in the plane in which both the oblique crank axis and the crankshaft axis 18 lie.

A surprising advantage of the planocentric as shown in Figures 15, 16, 17 or 18 over a fixed differential as shown in Figure 7-14 of the same overall diameter is that the planet gear eccentric journal 44 has greater eccentricity than for the fixed differential eccentric journal 42 meaning that it can make a greater

contribution to the balancing of the engine or pump for a given mass of planet gear.

Advantages

Reducing the cylinder pitch spacing as much as is practicable in an axial piston engine or pump with a rotary valve can produce a number of benefits;

• it can save size and mass,

• it can reduce the mass of the reciprocating components with potential consequent reductions in balance masses and bearing friction,

• it may allow the reduction of the sliding speeds of the rotary valve seals at a given engine speed by reducing the radial distance of the rotary valve openings from the axis of rotation, high sliding speeds can be extremely detrimental to rotary valve durability, and can also have greater frictional losses and wear,

• it may reduce the piston friction through reduction of the lateral forces on the pistons and connecting rods created by centrifugal forces as the cylinder cluster spins about its axis.

A more compact form of port indexing gearing such as a planocentric or fixed differential gear set as compared to epicyclic or parallel layshaft may potentially facilitate a reduction in the cylinder pitch spacing. The invention may also provide the advantage that a port indexing drive gear set is located where it is of reduced externally apparent gear noise.

The invention may also have the advantage of offering a port indexing drive fixed differential gear set which aims to create even wear on the teeth and reduced harmonic gear noise. Another advantage that may be provided is a method for creating an engine that imposes a minimum of inertial torque reaction loads on whatever vehicle or structure it is built in to in response to changes in engine speed or orientation. External engine torque reaction during engine acceleration (blipping throttle) and also when changing engine orientation are largely eliminated by the angular momentum of the crankshaft being cancelled out by the counter-rotating cylinder

cluster and reciprocator, this is a potential advantage for engine mounting and vehicle handling in vehicular applications.

The planocentric gear set may have advantages over the fixed differential set in that it is likely to be more compact and efficient with only a single gear stage comprised of two internally meshing gears with a relatively large number of teeth in contact. In the preferred forms the mechanism to constrain the rotation of the planet gear utilises bearings with frictional losses that are expected to be low, and the mounting for the bearings in the rotational constraint mechanisms of the preferred forms should also limit the creation of undesirable bearing forces that could otherwise be created by small dimensional mismatches between the various components of the planocentric gear set.

The planocentric gear set having planet gear rotational constraint means located at radii outside of the periphery of the planocentric gear teeth should potentially further improve the efficiency of the port indexing drive.