An axial piston machine is a machine in which a plurality of axially extending cylinders, together comprising the cylinder cluster, are arranged in a generally rotationally symmetrical layout around a central axis coincident with the rotational axis of a crankshaft. Each cylinder has an axis parallel or slightly inclined to that of the other cylinders. Each cylinder contains a reciprocating piston that may reciprocate along the cylinder axis. Axial piston machines may offer a number of potential advantages over other multi-cylinder piston machine configurations including: reductions in size and weight, simplified fluid porting, and the ability to achieve close to perfect balancing of the dynamic inertial forces.
There are a number of different mechanisms that can be used to drive the reciprocating motion of the pistons in their cylinders, two of the most common types being swashplate drives and wobbleplate or Z-Crank drives. While terminology can vary, a swashplate is in effect a cam surface attached to and rotating with the crankshaft that drives or is driven by the reciprocating linear motion of the pistons. Each piston has a bearing or bearings attached to it that slides or rolls over the surface of the swashplate cam surface. Each piston also has some form of linear bearing such as the side of the piston within its cylinder that re-acts the lateral forces created by the action of the piston-driving bearings when on the inclined surface of the swashplate. Piston-swashplate bearings will generally have a high sliding or rolling speed over the swashplate, which is proportional to engine rotation speed and radius from rotational axis. While this arrangement is adequate for axial piston machines having relatively low piston speeds such as compressors and hydraulic pumps or motors, modern internal combustion engines commonly have much higher piston reciprocity speeds, and the high inertial loads and bearing sliding or rolling speeds in a swashplate drive operating with high piston speeds can lead to high frictional losses that make swash plate drives less attractive for internal combustion engines. Z-Crank drives employ an intermediate body known variously as a wobbleplate, reciprocator or spider that rotates on bearings mounted on a crank section inclined to and intersecting with the crankshaft's rotational axis at an acute angle hereinafter referred to as the “crank angle” at a point hereinafter referred to as “point X”. The reciprocator is restrained against rotation with respect to the cylinder cluster by a torque restraint mechanism that may be implemented using a variety of mechanisms so that the rotation of the inclined crank causes the reciprocator to nutate or vice versa. The torque restraint mechanism thereby synchronising the reciprocator rotation with the cylinder clusters rotation. U.S. Pat. No. 4,235,116 describes such a mechanism. So does WO9859160.
The radial distance from point X to the point where features of the reciprocator are provided for rotation constraint, remain fixed.
The connection between the reciprocator and pistons can take many forms but generally connection rods having joints with two or more rotational degrees of freedom are utilised at both ends to connect to the piston and the reciprocator respectively. WO9859160 shows an example.
In a typical Z-crank drive configuration it is common that the connection rods (con-rods) are connected by joints with multiple degrees of freedom to both the piston and the reciprocator, the big end connection being where the connection rods connect to the reciprocator. The connection rods preferably take peak loads when the connection rods are substantially parallel to the cylinder axis. This reduces significant side loading of the piston onto the cylinder wall.
The radial distance from point X to each big end connection point remains fixed. Prior art Z-crank engine arrangements are known where the cylinder cluster rotates relative to the stationary engine housing. An example is described in WO9859160. This type of engine requires an indexing mechanism to index the rotation of the cylinder cluster relative to the ports. In WO9859160 the indexing is achieved by bevelled gears. Synchronisation is also achieved by bevelled gear.
As described in WO9859160, the synchronisation bevelled gear carried by the reciprocator operates at a distance from point X that is fixed and that is substantially smaller than the distance from point X to the big end connection point of each connection rod. One of the disadvantages of this position of the synchronisation bevelled gears is that the pitch line forces of the teeth in contact are high and the torsional stiffness is low due to the torque being reacted at a relatively small radius.
As shown in FIG. 1, a further disadvantage is that due to the gears 1000, meshing on a median MA line, the reciprocator 9A may need to be configured around the gear positioning in order to place the big end connection on the equatorial plane EA. This can result in the load transfers of the input forces (by the piston 11A) via the con-rods 12A being transferred via the reciprocator 9A in less direct paths and increased stress on or masses of the reciprocator structure.
A solution to these problems is to make the diameter of the bevelled gear bigger. This is described in FIG. 5b of U.S. Pat. No. 7,117,828 where its diameter is such as to place the bevelled gear around the outside of the connection rods. But a disadvantage of this is that the overall size of the engine increases. This adds mass at relatively large radius and therefore substantial inertia.
Described in FIG. 4 of U.S. Pat. No. 5,109,754 is the use of a bevelled gear placed close to the location of the big end connection.
In such a configuration, the bevelled gears (needing to meet at a median line) will dictate where the big end connection can be made in order to prevent the bevelled gears clashing with the big end connection and/or connection rods.
This may result in the big end connection being placed substantially away from or to the equatorial plane of the reciprocator. This may generate linear vibrations of the 2nd or higher order due to the pistons then being controlled for movement in significantly non-sinusoidal manner. Vibrations of this type would be very detrimental to the operation of an engine at high speeds or in vibration sensitive applications.
Bevel gears do not beneficially lend themselves for use near the big end connection and/or connection rods. At gear PCR's allowing operation clear inside of the con rods, a bevel gear is subjected to high pitch line forces, due to re-acting the required torque at a small radius. Also the positioning of the bevel gear compromises the reciprocator structure by operating through space between the con-rod connection point and the centre of nutation. The result is a low maximum cylinder pressure, and hence low performance, that can be tolerated using practicable materials and construction methods.
At gear PCR's allowing operation clear outside of con rods and con rod to reciprocator connections, a bevel gear is subjected to high pitch line speeds at high operating speeds typical in automotive IC engines. High pitch line speeds require high precision and mechanically stiff components to maintain precise gear tooth engagement and lubrication.
Such precision is not practicable in a high performance lightweight IC engine. Positioning of a bevel gear outside of the con rods also requires a large gear component and related mounting structure with significant inertia and mass, leading to high inertia forces being applied to bearings and structures, which in turn must be made larger to carry these increased loads.
The practicable result is a low maximum operating speed, and hence low performance with large relative engine size and mass.
In addition to the above packaging limitations, bevel gears also have a number of intrinsic mechanical limitations related to torque transfer, speed, tooth strength etc. As a consequence of these limitations a number of alternative rotational restraint mechanisms have been provided. Such as for example, roller and slot arrangements, as disclosed in U.S. Pat. Nos. 1,948,827 and 2,917,931. Such mechanisms are typically located in a radial location on the equatorial plane of the reciprocator, corresponding to the same plane as the connecting rod to reciprocator connector. While these substitutions have addressed a number of the issues intrinsic to bevel gears they have also introduced further limitations such as restraint tip speed, path and rotational velocity which all affect the overall performance envelope of the engine.
A restraint mechanism located off the equatorial plane is disclosed in U.S. Pat. No. 2,182,213. However, further disadvantages of this restraint mechanism is that only this particular engine embodiment is allowed, the restraint mechanism is always in a fixed relationship, leading to lubrication, wearing and/or reliability issues.
Restraint mechanisms presented in the prior art make no mention or attempt to achieve the required torsional stiffness for realistically operating in a high speed automotive sized IC engine. Low torsional stiffness is a characteristic apparent in much of the prior art, and especially those systems where the big end is located close to point X.
It is accordingly an object of this invention to provide improvements in or relating to an axial piston machine that offer freedom of design of rotation constraint mechanisms that do not significantly compromise engine operation, particularly but not solely at high or low speeds, provide improved packaging of the engine components, improved performance and/or that at least offer the public a useful choice.
USPTO Backgrounds
USPTO Backgrounds