WO1990005863A1

WO1990005863A1 - Motion transfer apparatus

Info

Publication number: WO1990005863A1
Application number: PCT/AU1989/000502
Authority: WO
Inventors: Michael James Durack
Original assignee: Durack M J
Priority date: 1988-11-18
Filing date: 1989-11-20
Publication date: 1990-05-31

Abstract

A motion conversion or motion transfer apparatus comprises a first member (5) rotatable about a first axis (6) and inclined thereto. A second member (7) is inclined at the same angle as the first member and is constrained to rotate about an axis (8) spaced from but parallel to the first axis. When a working surface of the first member is brought into contact with a working surface of the second member, rotation of the first member causes rotation and reciprocation of the second member on its axis. Similarly reciprocating motion induced in the second member causes rotation of both the first and second members. Practical applications of the invention include fluid pumps and motors as well as impact tools and other power transfer mechanisms.

Description

MOTION TRANSFER APPARATUS

This invention is related to a motion transfer apparatus of the type wherein a rotational motion may be converted to reciprocating motion or vice versa.

Several aspects embodying the fundamental principles of the invention have particular although not exclusive application to fluid pumping and fluid drive apparatus as well as a variety of power transfer mechanisms which may or may not include a reciprocating motion or vibration in the rotatable final drive.

Fluid pumps and motors in common use includes apparatus of the axial-piston type, in which an angled end face on a rotating "swash plate" operates to reciprocate axial pistons in contact with the swash plate, and the pistons either force fluid in and out of a valving system in the case of a pump or are forced in reciprocating motion in the case of a motor.

In order to permit the ends of the pistons in current designs of swash-plate apparatus to co-operate with the inclined rotating surface of the swash plate, complex articulate joints including ball joints and slipper pads are typically interposed between pistons and swash plate, leading to increased cost and complexity of the apparatus. As there is no control over rotation of the pistons, valving for controlling fluid flow must be achieved by motion of other moving components, again increasing the complexity of the apparatus.

Motion conversion apparatus which convert rotational motion to reciprocating motion (and vice versa) are well known and take many forms such as motors, pumps and the like.

Typical examples of "swash plate" apparatus are described in United States Patent Number 4365940 and Australian Patent Numbers 512 192, 436390, 419076, 522406 and 439184.

Other types of motion transfer or motion conversion apparatus are exemplified in United States Patents Number 4437016, 2386,675, 2578559 and 1521364.

United States Patent Number 2,622,567 entitled "Rotatable Piston Machine" describes a two stroke internal combustion engine comprising a rotatable swash plate assembly associated with one or more reciprocating pistons. The pistons are positively compelled to rotate during reciprocal movement by means of one or more dogs on the pistons which engage in a complex pattern of arcuate grooves in the swashplate. Means are provided to retain the pistons captive to the swash-plate to retain engagement between the piston dogs and the swash-plate grooves.

Each of the above prior art mechanisms is an adaptation of well known mechanical principles and generally speaking each mechanism is highly complex, expensive to construct and otherwise inherently limited in their respective fields of application. It would appear that the only commercially successful application of these prior art motion conversion apparatus is in the field of pumping of liquids.

It is an aim of the present invention to overcome or ameliorate the disadvantages of motion transfer or motion conversion apparatus adapted for use as fluid pumps, motors and various power transfer devices and to provide a simple relatively inexpensive apparatus which may be adaptable to a wide range of end uses in a number of alternative configurations.

From investigations of prior art motion transfer or conversion devices, it is believed that the present invention may embody a hitherto undiscovered principle of operation or at least a principle of operation which has not been applied for useful purposes.

According to one aspect of the invention there is provided a motion transfer apparatus comprising:- a first member rotatable about a first axis, said first member including at least one first planar working surface lying in a first plane inclined to said first axis; and,

a second member reciprocable along and rotatable about a second axis spaced from and substantially parallel to said first axis said second member including at least one second planar working surface lying in a second plane inclined to said first plane, said second plane being substantially parallel to said first plane to permit complementary engagement between said first and second working surfaces, said motion transfer apparatus characterized in that rotation of said first member about said first axis causes reciprocation and rotation of said second member about said second axis by interaction of respective resultants of reactive forces between said first and second working surfaces.

Alternatively or additionally, displacement of said second member towards said first member causes rotation of said second and first members about respective rotational axes by interaction of respective resultants of reactive forces between said first and second working surfaces.

Suitably said first member may comprise more than one parallel first working surfaces. Similarly said second member may comprise more than one parallel second working surfaces.

Preferably rotation of said second member is induced by displacement, from the second rotational axis, of opposing reactive forces between the respective planes of respective working surfaces of said first and second members.

Preferably displacement of reactive forces from said second rotational axis induces a torque in said second member about said second rotational axis.

If desired, either or both working surface may be formed as a complex shape such as a cone whereby selected forms of relative motion between the first and second members may be achieved.

The first and second axes may be inclined to one another at any desired angle, and may be placed in different planes such that any desired form of relative motion between the members may be achieved. However, it is preferred that the first and second axes be parallel such that the first and second working surfaces may be formed at substantially equal inclinations to the first and second axes respectively and form a co-operative pair of surfaces.

The first and second surfaces may be maintained in operable conjunction by biasing means, and the biasing means may include any suitable biasing means such as springs, limit stops arid/or fluid pressure. Alternatively, one of the members may be formed with a pair of opposed internal working surfaces having a cavity therebetween and the other member may be formed with a projection having a pair of opposed external working surfaces thereon, the opposed internal working surfaces "being co-operable with the opposed external working surfaces for maintaining respective pairs of external and internal surfaces in operable contact.

The working surfaces may be fixed at pre-determined angles to the respective axes, or, if desired, may be attached through adjustment means whereby the inclination may be varied such that the reciprocating stroke of the reciprocable member may be varied.

Suitably, the members may be formed with cooperating cylindrical surfaces such that a fluid chamber is formed therebetween, and porting means may be disposed along or around the co-operating surfaces and/or along the cooperating surface of a third member such as a housing whereby fluid transfer to and from the chamber may be controlled.

The motion transfer apparatus may be formed as a hydraulic pump or motor of the swash-plate type, the first member being in the form of a swash plate rotatable in a housing about a swash plate axis, and the second member being in the form of a piston slidable along and rotatable about a piston axis, and an end face of the piston being inclined to piston axis at an inclination such that it may co-operate with the swash-plate, whereby the construction of a swashplate motor may be simplified. If desired, the piston may constitute one of a plurality of pistons disposed about respective axes parallel to one another, the pistons cooperating with a common swash-plate.

In a preferred embodiment of the invention, the first member may be in the form of a drive shaft rotatable about an axis within a housing, the drive shaft having a drive portion including an opposed pair of inclined drive faces between which is disposed an annular piston having driver faces co-operable with the drive faces, the annular piston being reciprocable and rotatable within the housing in co-operation with the drive shaft. The housing, piston or drive shaft may include fluid ports co-operable with portions of other members for controlling fluid flow about the piston.

In order that the invention may be more easily understood and put into practical effect, reference will now be made to the accompanying drawings which illustrate a number of preferred embodiments and wherein:-

FIG 1 shows schematically one aspect of the invention.

FIG 2 shows schematically an alternative aspect of the invention.

FIGS 3-6 show a fluid pump mechanism according to the invention.

FIGS 7 and 8 illustrate a fluid motor according to the invention.

FIGS 9 and 10 show schematically the principle underlying the operation of the apparatus of FIGS 7 and 8.

FIGS 11 shows schematically portion of a mathematical model of the invention. FIGS 12-16 illustrate theoretical performance criteria of the mathematical model of FIG 11.

FIGS 17-20 illustrate a wheel hub motor according to the invention.

FIG 1 shows one aspect of the invention which, for the purposes of simplicity may be considered as a pair of spaced discs 1, la inclined to an constrained to rotate about axis 2. The inclined pair of spaced discs 1, la thus sweep a generally cylindrical volume through a 360° rotation.

Located between the inwardly opposed surfaces of discs 1, la is a further disc 3 constrained to rotate about axis 4 and is able to move freely therealong. The externally opposed surfaces of disc 3 are in sliding contact with the inwardly opposed surfaces of discs 1, la and the surfaces are considered to be frictionless.

As discs 1, la are caused to rotate in the direction shown, a "camming" or "wedging" force is applied to disc 3 to urge it to move upwardly along axis 4. As portion of disc 3 is sandwiched between opposing discs 1, la there are a plurality of forces being applied between the overlapped engaging portions of discs 1, la and disc 3.

These forces are resolved as two main resultants-- one force causing disc 3 to move upwardly along axis 4 and another causing disc 3 to rotate about axis 4 in the same rotational direction as disc 1, la. As discs 1, la rotate about axis 2 through 180° as shown in phantom, disc 3 is elevated to a maximum position. After rotation through a further 180° to complete a single revolution discs 1, la return to the original position shown. It will be noted that disc 3 rotates at the same rotational speed as discs 1, la regardless of whether the diameters of discs 1, la and 3 are the same or different.

FIG 2 shows an alternative embodiment of the invention.

A disc 5 is inclined to and constrained to rotate about axis 6 to sweep a cylindrical volume. A cylindrical member 7 is constrained to rotate about an axis 8 spaced from and parallel to axis 6 but is otherwise free to slide along axis 8. Cylindrical member 7 has an inclined face 7a lying in the same plane as the upper surface of disc 5. Once again, substantially frictionless engagement is assumed between the abutting surfaces of cylindrical member 7 and disc 5.

In the rest position shown the static forces applied between the member 7 and disc 5 are a downward force due to the mass of member 7 and a reactive force exerted by disc 5. The resultant of these forces is a lateral force represented by arrow 9 which is countered by an opposite reactive force 9a exerted by the constraint of member 7 to move along or to rotate about axis 8.

At equilibrium, all resultants of active and reactive forces pass through axis 8 at the interface of the abutting respective surfaces of disc 5 and member 7.

As disc 5 is rotated through an infinitesimally small angle, the leading surface of disc 5 imposes a

"wedging" or "camming" action on the inclined face of member

7 whereby the point of intersection of the resultant active and reactive forces moves outwardly in a radial direction away from both rotational axes 6 and 8.

The equilibrium of resultant active and reactive forces is thus disturbed and, in effect, the shift of the centre of reaction of resultant forces away from the rotational axis 8 of member 7 creates a moment arm through which reactive forces in the system apply a torque to member 7 to restore the equilibrium of resultant active forces. As disc 5 is rotated stepwise by infinitesimal angular steps through 180° to the position shown in phantom, member 7 also rotates in the same angular direction through 180°.

Continued rotation of disc 5 thus causes member 7 not only to reciprocate along axis 8 but also to rotate at the same angular speed as disc 5.

The reverse of the above procedure is also possible.

By applying a downwardly directed axial force on member 7 whilst in any position elevated above the lowermost position shown in bold outline, the interaction of reactive forces disturbs the equilibrium otherwise existing in a rest position and induces a shift of the centre of reactive forces away from rotational axis 8 of member 7 thus creating a moment arm through which restorative reactive forces act to apply a torque to member 7 to rotate member 7 about its axis 8 while simultaneously inducing a rotation (in the same angular direction) of disc 5.

By applying forces alternating in opposing directions along axis 8 by opposing members 7 and 7b (shown in phantom), disc 5 can be caused to continuously rotate about axis 6.

It can be seen therefore that the apparatus according to the invention may be utilized to translate rotational motion to reciprocatory motion or vice versa. A particular advantage arising from the invention however is that whether rotational motion is translated to reciprocatory motion or vice versa, both of the co-acting members rotate at the same angular speed. This phenomenon readily lends itself to exploitation by providing a very simple and inexpensive means for valving pumps or motors of the swash-plate type as well as other advantages in power transfer mechanisms described later with reference to other aspects of the invention.

FIGS 3-5 show a fluid pump generally in accordance with an aspect of the invention shown in FIG 1.

The fluid pump comprises a rotor assembly 10 as shown in FIG 3, the rotor assembly 10 being eccentrically located within a housing 11 as shown in FIG 4 which is a top plan view. Rotor assembly 10 comprises a generally cylindrical body 12 with a lower bearing shaft 13 and an upper drive shaft 14. Body 12 has an annular slotted recess 15 extending about the periphery of body 12, slotted recess 15 being inclined to a central rotational axis of the rotor assembly 10.

A generally annular piston member 16 is slidingly located within recess 15 and is capable of eccentric movement within the recess about a central shaft 17 by virtue of an enlarged aperture 16a in piston member 16.

As rotor assembly 10 is constrained to move about its rotational axis and piston member 16 is constrained to rotate about an axis located centrally of housing 11, the respective rotational axes being spaced and parallel, the portion of piston member 16 which protrudes from slot 15 reciprocates between a maximum and minimum position shown in

FIG 3 as the rotor assembly rotates through 180°. FIG 5 is a perspective view illustrating the positions of maximum and minimum displacement of piston member 16.

Inlet and outlet ports 18, 19 formed in housing 11 are shown in FIG 4. The shaded regions 18, 19 shown in FIGS

3 and 5 represent the alignment of piston member 16 with the outlet ports at the maximum and minimum positions of reciprocation.

The operation of the pump will now be described with reference to FIGS 3 and 4.

The rotor as sembly 10 commences rotation in an anti-clockwise direction with the piston member 16 at its lowermost position shown in phantom in FIG 3. In this position both the inlet port 18 and outlet port 19 are closed by piston member 16. As the rotor assembly 10 continues to rotate, both the inlet port 18 and outlet port 19 are progressively opened, the inlet port 18 and outlet- port 19 are progressively opened, the inlet port 18 being opened below piston member 16 and the outlet port 19 being opened above piston member 16. The piston member 16 forces fluid out of the outlet port 19 as it elevates to its maximum position as shown in FIG 3 whereupon inlet port 18 and outlet port 19 are again closed.

As the rotor assembly rotates through a further

180°, the inlet port 18 opens into the region above piston member 16 and a low pressure region is created above piston member 16 to draw fluid into the pump.

With suitable adaptation the pump described above is particularly suitable for "down hole" bore pumps as it is compact, capable of producing a very large head and otherwise requires relatively lower power. In a typical "down hole" bore pump the housing 11 is lined with a resilient rubber or plastics material to reduce the effects of wear caused by particles of sand or grit commonly formed in bore water.

FIG 6 illustrates an alternative embodiment of a pump described with reference to FIGS 3-5.

In the rotor assembly 20, body member 21 includes a plurality of piston members 22 with respective planes of inclination phased progressively through 180°. Each piston member 22 is associated with a respective pair of inlet and outlet ports in a housing (not shown) to achieve a high capacity pump with minimal pressure or volumetric surging.

To test the effectiveness of the invention in the form of a motor, prototype in the form of a simple hydraulic motor was constructed in accordance with the exploded view shown in FIG 7. The assembly of FIG 7 is conveniently shown in schematic form in FIG 8.

In FIG 7 the apparatus comprises a body member 30, opposed head members 31, 32 and side plates 33, 34 to facilitate assembly of the apparatus by bolts 35 passing through apertures 36 in side plates 33, 34 to corresponding threaded apertures 37.

Located within recesses 38, 39 in head members 31, 32 are thrust bearings 40 to receive rotatable drive members or pins 42 each of which has a planar working face 41a, 42a inclined with respect to a longitudinal rotational axis. The angle of inclination of respective faces 41a, 42a is identical and located between those working faces is a piston member 43 having parallel working faces 43a, 43b complementary to the angle of inclination of working faces 41a, 42a.

Rotatable drive members or pins 41, 42 were fitted with gear members (not shown) which meshed with a gear set on a common shaft (not shown) for the purposes of preliminary tests.

OPERATION The device of FIGS 7 and 8 is simple but not simple to explain. It produces torque from two rotating inclined surfaces forced to maintain alignment with each other while circulating about displaced axes. The particular design tested was a single cylinder, double acting hydraulic motor with a theoretical charge of 96 cc per revolution. The piston has a stroke (which can be varied from zero) of up to 17mm, and a diameter of 60mm. The swept volume is simply cylinder cross sectional area x stroke x 2 (double acting). As the piston is pushed by the high pressure fluid it applies axial and rotary forces on the opposite pin. The axial forces cannot cause any movement and therefore do not work but the rotary forces cause the pins to rotate and thus rotary power can be extracted. OPERATIONAL TESTS

Procedure

The motor was coupled directly to an output shaft which carried a flywheel and a disc brake (not shown). It was supported by trunnion bearings so that its torque reaction could be measured. The torque reaction (T) of the motor frame (not shown) is equal and opposite to the torque imposed upon the drive shaft (not shown) by the motor. Hence it was possible to compute the mechanical efficiency (n) as the ratio of the torque x running speed (w) and the pressure drop across the motor (P) x the displacement (q). Torque reactions were measured by adding weights to the end of an arm extended horizontally out from the motor. Various weights were applied and the pressure drop across the motor was recorded when the applied moment was just balanced by the motor torque reaction. This was achieved by gradually applying the brake until the weights just started to lift.

Observations

The motor ran remarkably well considering that it was of single cylinder design. The large flywheel kept the output speed very steady and this threw the reactions of the pulsating fluid flows onto the .torque reaction and caused the torque measuring arm to vibrate somewhat. This is normal for machines of this class, and was expected. As in current practice, the vibration can be significantly reduced by designing the motor to have more cylinders. Also, the motor would always be firmly mounted to some base which would usually have more inertia than the driven load and as a consequence, the vibration displacement amplitudes would be very much smaller than in our tests.

One advantage of the invention is the ease with which it lends itself to double action and a design with six cylinders (for example) would have twelve pulsations per revolution.

For certain preferred initial angular orientations of the drive shaft, the motor was self starting. A multi cylinder motor would be self starting from any initial position. The motor exhibited a stalled torque of the same magnitude as the running torque, as expected, which means that it has the characteristic (typical of hydraulic motors in general) of being able to be started under full load.

The motor was run at high speed (1000 RPM) without load for a short time. It operated very well and could have gone much faster but the test rig was not adequately protected against possible accidents. It became apparent through this test with a large displacement, single cylinder design, that machines with smaller individual cylinders and multiple cylinders will be able to be run at very high speeds. By comparison with conventional machines this aspect of the invention does not have any of the same limitations on allowable operating speeds and this means that for given operating pressures and displacements, higher powers will be possible.

When attempting to produce torques which demanded pressures greater than about twelve Mpa, the pressure which was needed increased, but the reaction torque did not. This indicated a phenomenon which could be attributed to either the particular embodiment or to the aspect of the invention itself.

The possible reasons for the phenomenon are:

Jamming of the rolling element bearings which support the pins;

Loss of lubrication film at either (or both);

- The pin - piston interface;

- The piston - cylinder wall interface.

To check the first of these, tapered roller bearings which provided the support for the axial forces were moved and thrust bearings were inserted. This gave only marginal change in the operation, which may have been due to slight change in the stiffness of the pin supports. Thus the resistance observed at higher pressures seems to be due to one or both of the other reasons above.

It was estimated that the forces exerted by the piston sides upon the cylinder walls are of the same order as those exerted upon the pins by the piston faces. If the friction coefficients in both of those sliding contacts are similar, the amount of energy which is expended by the piston in wall friction is (2πRx0.707/ρ) 16 times as high as that

Table 1: Torque and mechanical efficiency test results

which it expends in pin friction. It can be shown that by a simple re-design of the piston the wall forces can be eliminated, whilst at the same time the relative stiffness of the pin to that of the piston can be significantly increased.

Hence it is concluded that the operating pressures will be able to be increased to values currently in use in compatible devices, and that the phenomenon can be adequately overcome using suitable designs.

The measured torque and mechanical efficiency results are provided in Table 1. KINEMATICS

As the pins rotate, so the piston must also rotate, as well as move axially, in order to conform with the slanting end faces of the pins. The angular velocities of the piston and the pins are exactly equal to each other at all times and the axial motion of the piston is purely sinusoidal. The relative motion of the piston and the pins is constant at all points of contact as viewed along the axis of the motor. This is simple to prove. FIG 9 shows an axial schematic of the motor. The distance between the central axis of the piston (O₁) and the pins (O₂) is ρ = R - r.

Consider any point (A) of contact between the piston and a pin. In a plane perpendicular to the axes, the motion of the pin at "A" is at right angles to the line joining A and O₂ and has magnitude ω x AO₂. Similarly, the motion of the piston at "A" is at right angles to AO₁ and has magnitude ω x AO₁. The relative velocity is the difference given by ρ ω directed always at right angles to the straight line through the pin and piston centers. This proves that the component of the relative velocity (normal to the motor axis) is constant for a given operating speed.

If the angle of slant of the piston is ø and the rotational displacement is θ the actual sliding velocity is ρω (1 + tan² ∅ sin² θ)^0.5. For ∅ = 45°, as in this motor the result is a sinusoidal variation between ρω and √2 ρω , twice per revolution. The sliding velocity is important to the operation of the motor because of its effect on the lubrication regime between the piston and the pins.

FORCES AND MOMENTS

If it is assumed for the initial analyses that there is no friction, upper bound values can be obtained for the forces and moments acting throughout the important parts of the mechanism. The true forces can then be ascertained by including friction effects, and limiting coefficients of friction can be determined.

It is not valuable to assume that the mechanism is infinitely stiff and it is necessary to realize that in the tested configuration, the resistance to radial relative movement of the pin is significantly lower than radial relative movement of the piston in the cylinder.

With the above considerations in mind, assuming firstly that there is no friction, and considering the cylinder walls to be infinitely stiff in comparison with the pins, the configurations which give rise to the greatest forces are when the piston is at either top or bottom dead centre.

The forces acting on the piston are shown in Figure

8 for the case when the piston is at bottom dead centre with high pressure oil below. The reaction of the pin on the piston is at the "heel" of the pin because of its greater flexibility. The force (F) due to oil pressure acts on the centre of the lower face of the piston and the pin reaction force is equal and opposite (F) on the top face, but offset due to the flexibility of the pin. These two forces impart a moment to the piston which tends to rotate it out of alignment in the cylinder and cause the equal and opposite wall reactions (f). The wall forces on the sides of the piston (f) are 0.707F but they can be reduced drastically by redesigning the arrangement as shown in Figure 7. The forces (f) can also be reduced in this arrangement by improving the relative stiffness of the pins with respect to that of the piston against the cylinder wall. The slanting part of the piston can be made suitably flexible that the reaction of the pin (force F) can be forced into alignment with the central force due to the pressure of the oil. Thus there will be no moment tending to twist the pistons into the walls, and the wall reactions (f) will be zero. f= Fx = L tan θ + R/Cos θ

y L + 2 R tan θ

THEORETICAL TORQUE

Figure 10 is a view down the axis of the apparatus. The line of action of the pressure force (F) has a component normal to the axis (F_n) which is always directed radially towards the tip of the piston. The pin reaction has the normal (to axis) component (also F_n) in line with it but the bearings supporting the pin can only provide reactions radial to the pin and thus there is an offset between the two forces which causes the pin to rotate. For any rotation (θ) as shown the theoretical torque (T_t) is ρ sinθ x F_n. The average theoretical torque (T_m) is 2/πx the peak torque

(at θ = π/2) and this is 15.3 x the pressure differential across the piston. Table 2, provides a comparison of the theoretical torques and consequent efficiencies (neglecting volumetric losses) with those from the results shown in Table 1.

Table 2: Theoretical torque compared with test results

It can be seen in Table 2 that the measured efficiency values based on measured pressure differentials and measured torque reactions, agree very closely to the ratios of measured and calculated torques.

PORTING Hydraulic fluid enters and leaves the cylinder as the piston uncovers the ports in the cylinder wall. There are lead in and lead out slots in the piston to assist the timing of fluid transfer, see FIG 7. There are two advantages which are inherent in the mechanism as compared to conventional piston motors:

The porting can be made larger;

The fluid can be made to flow in one side and out the other, thus avoiding much almost entirely any flow reversals.

These allow either higher operating speeds or operation with much more viscous fluid.

After testing the initial prototype as shown in FIGS 7 and 8, a detailed theoretical analysis of the invention was carried out and a mathematical model was developed to consider mechanical efficiency when acting as a pump or motor. The study was primarily aimed at exploring the application possibilities of the mechanism to hydraulic transmissions where a minimum acceptable level of efficiency is required.

EFFICIENCY STUDIES OF THE PISTON MECHANISM

OF THE PRESENT INVENTION AS APPLIED

TO HYDRAULIC MOTORS AND PUMPS 1. INTRODUCTION AND SUMMARY

The study showed that mechanical efficiencies of above 92% with the mechanism acting as a motor, or slightly higher as a pump, can be achieved at an assumed dynamic friction coefficient of 0.05. This level of efficiency is considered to be competitive with vane and gear pumps and motors and with gerotor type motors. The apparatus according to the invention has advantages over these other types of equipment including the potential of improved volumetric efficiency as the main sealing requirement is a diametrical piston fit in a cylindrical bore. A motor according to the invention has the potential to provide similar size to displacement ratios as the very compact gerotor type motors, but with a higher speed capability, leading to an enlarged market potential.

Similar efficiency levels can be achieved as a water pump, with the possibility of replacing much less efficient centrifugal pumps in critical applications. The compact size of a pump according to the invention would permit convenient application as a down-the-bore water pump, replacing existing multi-stage centrifugal pump designs with a single stage high efficiency piston pump according to the invention.

in summary, these studies confirm that hydraulic pumps and motors using the principle of the present invention can be made to have satisfactory efficiency levels. The same efficiency levels also provide useful performance as a water pump.

2. BASIC DESIGN

The basic design analysed is shown on Figure 11. This is one of many possible configurations according to the invention, chosen as being simple to manufacture and being capable of operating at medium hydraulic pressures (nominally 20 MPa, 3,000PSI). Other designs may be devised that are more efficient, but are expected to be more complex. and less compact. Similarly, other designs may be devised that are more compact, but are expected to be less efficient and to have a lower pressure capability.

Referring to FIG 11, only part of a single element or piston assembly is shown, consisting of a piston 50 that rotates with the body 51 and shaft 52 and reciprocates up and down in the cylinder bore as the shaft rotates. The components shown provide for a power stroke as a motor over 180° of each rotation.

A complete element would have a second body on top of the piston providing a power stroke over the second 180° of rotation so that the element provides for two power strokes for each shaft revolution.

The high pressure area is shown at 53, pushing the piston 50 against the angled face 54 of body 51. The piston 50 is shown with a top and bottom sealing lands 55 on its circumference. A pressure feed drilling 57 connects the pressure in the shaft space 56, which is always equal to the high pressure, to the annulus between the sealing lands 55, 56. This arrangement is designed to bring the point of pressure application on the piston 50 as close as possible to the angled body face 5.5, to minimize any. tendency of cockling. The low pressure area is shown at 60.

The angled body face has a hydrostatic balance area 58, also connected to the shaft space, to minimize the mechanical loading between the piston 50 and the body 51.

The bottom face of the body 51 is also shown with a hydrostatic balance area 59. It is possible either to make this area large enough to fully support the thrust load from the piston 50, or to make tandem elements with two opposed pistons that balance out all thrust loads. In either case there is no theoretical efficiency loss at the angled body thrust face, and this condition is assumed for this study.

In operation, the shaft 52 body 51, and piston 50 all rotate, with the piston 50 being forced to a planetary action with respect of the body by the eccentricity of the cylinder bore 60. It is this planetary action that provides the vertical reciprocation of the piston 50 which in turn provides the displacement of the assembly.

There is a relatively low sliding velocity at the interface between the piston 50 and body 51 with considerable mechanical load. There is high sliding velocity between the piston 50 and bore 60 but with only very light loading, derivative entirely on the friction at the inclined face of body 51.

All main lateral loads at the piston 50 are fully supported by the body and transmitted to the shaft 52, where it is possible to mount high efficiency rolling bearings, so that the losses from lateral loads are relatively low.

The main mechanical loss is generated at the interface between piston 50 and the inclined face of body 51.

MATHEMATICAL MODEL

A mathematical model was compiled taking into account all the main forces and their derivative friction forces. The effects of component inertia, viscous drag and component deformation were ignored. Experience with similar modelling of hydraulic pumps and motors have shown that any variances introduced by these factors are well covered by the assumption of a coefficient of friction.

The possible effects of porting were also discounted for the following reasons:-

The simplest porting produces a side-load onto the piston against the cylinder bore as the piston crosses the port. The effect of such side-load is relatively unimportant at speed as it is only effective for a small part of the rotation, depending on detail porting design. However, the effect is much more severe as a motor which must successfully start from rest under load from the least favorable position. It is possible to design porting that is more complex but avoid the momentary side loads of the simplest design. For this reason the porting losses are not taken as being losses inherent in the design of the mechanism.

The model was formulated into an equation solving computer program, TK Solver, as listed in Appendix A. This program accepts all the equations of force and moment equality, with the defined input variables, and reiteratively calculates the solutions to the output variables. The analysis is fully three dimensional and calculates, at any specified angle of rotation, the friction force at the cam face and the resulting reaction and friction forces at the cylinder wall. From this the instantaneous power losses are calculated for a set pressure and shaft speed.

Appendix A shows first a list of input and output variables followed by a list of equations. The program also calculates lists and tables, as shown in Appendix B, and can plot the results.

All the analysis discussed below is calculated at a shaft speed of 1 rad/sec and at 1 MPa pressure. This is done merely for convenience and the calculations of efficiency are independent of speed and pressure because the equations used take no account of variations in friction due to speed and load.

Figure 12 shows the calculated instantaneous input power as a motor (output power as a pump) over 180° of rotation. The complete element would provide two such power pulses for each revolution.

Figure 13 shows the variation of instantaneous power loss over 180° of rotation. It can be seen that the power loss peaks at 90°, due primarily to the higher sliding velocity between the piston 50 and body 51 at this point.

The shape of the curve is very close to sinusoidal, and a the mean loss is calculated on this basis. It can be seen that the mean power loss almost coincides with the instantaneous loss at 45° rotation, and the curves shown on Figures 5 and 6 are derived from analysis at 45° rotation.

Figure 14 shows the instantaneous mechanical efficiency as a motor over the 180°. The efficiency remains at good levels over a wide range of positions, which will be a beneficial characteristic for starting from rest as a motor.

Figure 15 shows the variation of mean mechanical efficiency, calculated at the 45° position, with changes in eccentricity ratio, expressed as a proportion of piston diameter. A peak efficiency in excess of 92% is shown at an eccentricity ratio of 0.06, with levels above 90% over a range from 0.03 to 0.15.

In practice the displacement for size characteristic of the element is enhanced by larger eccentricities, but values above 0.10 are not readily achievable because the shaft diameter becomes too small.

A design with small eccentricity will tend to be shorter and larger in diameter than a design with a larger eccentricity ratio.

Figure 16 shows the results of a similar study with variations of piston angle. Peak efficiency occurs at about

52°, with 90% being available over a range from about 30° to 70°. This variation does not change greatly with changes in eccentricity.

The displacement for size characteristics are enhanced with larger piston angles. A larger piston diameter is required to give the same displacement at smaller piston angles.

In summary, the greatest efficiency is achieved with eccentricity ratios between 0.05 and 0.10 and with piston angle of between 40° and 60°. Improvements in efficiency above 92% (at 0.05 friction coefficient) can only be achieved with a design that allow a higher degree of hydrostatic balance at the inclined body face 51

* BDIA = BOFACT*POIA

* CDIA = BDIA-2*ΕOCEN-CLTH

* COCKLE = sqrt (X3-2+Z3-2)*2/(ODIA+CLTH)

"coordinate equations

* X2 = ECCEN* cos(THETA )

* Z2 = -ECCEN*s in (THETA)

* Y2 = X2*tan(ALPHA)+PLTH /2

* Y3 = X3*t an (ALPHA)

* Y4 = X4*tan(ALPHA)+PATH/2

* (X4-X2)-2+(Z4-Z2)-2 = (PDIA/2)-2

"pressure forces * PY1 = PRESS*ODIA-2*pi()/4

* PY2 = -PRESS*PDIA-2*pi()/4

* PX1 = -PY1*tan(ALPHA)

* PX2 = -PY2*tan(ALPHA) "force relationships

* RY3 = R3*cos(ALPHA)

* RX3 = -R3*sin(ALPHA)

* RX4 = -R4*cos(atan2((Z4-Z2), (X4-X2)))

* RZ4 = -R4*sin(atan2((Z4-Z2), (X4-X2)))

"frictional resultants

* F3 = MU3*R3

* FZ3 = F3*cos(atan2(sin(THETA)*cos(ALPHA), cos(THETA)))*PM

* FX3 = F3*cos(ALPHA)*sin(atan2(sin(THETA)* cos(ALPHA), cos(THETA)))*PM

* FY3 = F3*sin(ALPHA)*sin(atan2(sin(THETA)*cos(ALPHA), cos(THETA)))*PM * F4 = MU4*RA

* F4V = sqrt((EOCE N*tan(ALPHA)*sin(THE TA))-2+PDIA-2/4)

* FY4 = F4*EOCEN*tan(ALPHA)*sin(THETA)/F4V*PM * FX4 = F4*PDIA/2/F4V*sin(atan2((Z4-Z2), (X4-X2)))*FM

* FZ4 = -F4*PDIA/2/F4V*cos(atan2((Z4-Z2), (X4-X2)))*PM

"resolving in directions * PX1+PX2 +RX3-RX4+FX3 +FX4 = O

* RZ4+FZ3 +FZ4 = O "taking moments about O, O * PX2*Y2-PY2*X2+X3*((RX3+FX3)*tan(ALPHA) -RY3-FY3)+(RX4+FX4)*Y4-FY4*X4 = O * PX2*Z2+RX3* Z3+RX4*Z4-RZ4* X4+FX3 *Z3 -FZ3*X 3+FX4*Z4-FZ4*X4 = O

"instantaneous power by F3

* VEL3 = OWEGA*EOCEN*sqrt(1+ tan(ALPHA)-2*sin(THETA)-2)

* POWER3 = VEL3*F3

"instanteous power by F4 * VEL4 = OMEGA*sqrt((EOCEN*tan(ALPHA)*s i n(THETA)) -2+(PDIA/2) -2)

* POWER4 = F4*VEL4

"power by piston

* VEL2 = OMEGA*EOCEN*tan(ALPHA)*sin(THETA)

* POWER2 = VEL2*PRESS*PDIA-2*pi()/4

"calculation of cam power losses

* POWER5 = (PDIA-2-CTDlA-2)* PI()/4*PRESS*CTDIA/2*OMEGA*MU 5

* POWER6 = PDIA-2*Pi() /4*PRESS* tan(ALPHA)*BDIA/2*OMEGA*MU6

"efficiency etc * T I LOSS = PCWER3+POWER4-+POWER 5 +POWER6

* IEFF = (POWER2-TILOSS*step(FM, O))/(POWER2+TI LOSS*step(O, PM)) * M POWER = PD I A-2*pi()/4*PRESS*EOCEN * 2* tan (ALPHA)/pi()

* MEFF = (MPOWER-TILOSS*step(PMO))/(MPOWER+TILOSS*step (O, PM)) "checking equations

FIGS 17-20 illustrate an hydraulic wheel hub motor according to the invention. This embodiment is particularly suitable as a drive means for earth working machinery and other hydrostatically powered vehicles.

The drive motor shown comprises six single piston double acting hydrostatic drive elements with planetary gears driving a ring gear. The left hand side of Fig 17 shows a section through a piston/cylinder assembly while for the sake of simplicity the right hand side of Fig 17 shows hydraulic oil galleries associated with a piston/cylinder assembly.

Similarly, for the sake of simplicity, Figs 19 and 20 show partial cross sectional views of various quadrants through Fig 17.

In Fig 17, the motor comprises a body 70 comprising a first generally cylindrical section 71 and a second generally cylindrical section 72. An end plate 70a is affixed to one end of body 70 while a rotatable wheel hub portion 73 is rotatably mounted in body section 72 by means of a stub axle 74 and bearing assemblies 75, 76.

Located in the central portion of body 70 is a spool valve 77 slidable in an oil distribution enclosure 78 which forms a series of oil galleries 79 between inner body wall 80 and the spool valve 77.

A piston 81 is slidably and rotatably located in a cylinder defined by a cylinder liner 82 having a plurality of oil galleries 83 communicating via apertures 84 in body wall

80 with the oil galleries 79 of distribution enclosure 78.

Oil galleries 83 in turn communicate with opposed ports 85

(shown in phantom) in cylinder wall 82.

Piston 81 comprises a pair of oppositely inclined faces 86, 87 formed with annular recesses 88 in which are located seals 89. A region intermediate faces 86, 87 is formed with opposing recesses 90, 90a separated by opposing piston walls 91 which sealingly engage with an internal wall 92 of cylinder liner 82. A hydrostatic balance channel 93 is provided between recess 90 in piston 81 and lower recess 88 in the lower inclined face 87 of piston 81.

In the upper portion of cylinder liner 82 is located a retaining member 94 which rotatably locates a pin

95 having an inclined face 96 complementary to upper inclined face 86 of piston 81. A spacer or shim pack 97 is positioned above pin 95.

At the lower end of piston 81 is located a drive shaft 98 having an inclined face 99 engaging a complementary inclined face 87 of piston 81. Drive shaft 98 is journalled in a main bearing 100 and an oil seal 101 is provided between the drive shaft 98 and body portion 72.

A pinion gear 102 is mounted on drive shaft 98 and engages in a planetary manner with ring gear 103 mounted by grub screws 73a or the like on wheel hub portion 73. The end of shaft 98 is journalled in a bearing/oil seal assembly 104 associated with the body section 72 of body 70. A hydrostatic balance gallery 105 is provided between recess 88 at the interface of piston 81 and drive shaft 98. Gallery

105 is in fluid communication with a hydrostatic balance chamber 106 at the end of drive shaft 98.

A plurality of wheel studs 107 are provided in wheel hub portion 75 to facilitate attachment of a road wheel, drive sprocket or the like.

Fig 18 shows an external view of the wheel hub motor of Fig 17. End plate 70a and body sections 71, 72 are suitable connected by bolts 108.

Inlet and outlet ports 109, 110 are associated with spool valve 77 which in turn is operable by hydraulic pressure via speed control gallery 111. Port 112 is a case drain leakage port. A mounting flange 113 is provided on body 70 for mounting the motor to a vehicle or the like.

FIGS 19 and 20 show partial cross sections through the apparatus of Fig 17.

FIG 19A shows a cross section of a quadrant through A-A in Fig 17.

FIG 19B shows a cross section of a quadrant through B-B in Fig 17 .

FIG 19C shows a cross section of a quadrant through C-C in Fig 17.

FIG 19D shows a cross section of a quadrant through D-D in Fig 17.

FIG 20E shows a cross section of a quadrant through E-E in Fig 17.

FIG 20F shows a cross section of a quadrant through F-F in Fig 17.

FIG 20G shows a cross section of a quadrant through

G-G in Fig 17.

FIG 20H shows a cross section of a quadrant through H-H in Fig 17.

The six motor units shown schematically in FIG 19 are arranged in alternating groups of three whereby upon operation of spool valve 77 a group of three motors may be selectively isolated to permit high speed low torque applications. The wheel hub motor can thus be considered as a two speed motor although speeds with six or three motors operative can be infinitely varied for each range.

Each group of three drive elements is arranged such that the relative positions of the piston in each cylinder is phased to provide smooth operation.

The operation of the wheel hub motor will now be described with reference to Fig 17 although the general principle of operation has already been described with reference to Fig 2.

In Fig 17, depending on the desired direction of rotation of the wheel hub, pressurized hydraulic fluid is supplied to the inlet port. The pressurized oil finds its way via spool valve 77, distribution enclosure 78, apertures 84 and galleries 83 to cylinder port 85 whereupon it enters recess 90. Pressurized oil proceeds via gallery 83a in piston 81 (shown in phantom) to enter the cylinder between piston face 86 and retaining member 94. The high pressure oil drives piston 81 downwardly by rotation against the inclined face of drive. shaft 98 whereby drive shaft 98 is constrained to rotate. The upper inclined face 86 of piston 81 also bears against "idler" pin 95 which is also constrained to rotate.

During the downward movement piston 81 rotates through 180° as it reaches the bottom of its stroke as represented on the right hand side of FIG 17 and this in turn caused drive shaft 98 to rotate through 180° in the same direction.

At the bottom end of the piston stroke, oil gallery

83a is ported to low pressure and high pressure oil is ported via gallery 83b in piston 81 to the interior of chamber 114 whereupon it acts upon the effective cross-sectional area of the piston to drive it back along its rotational axis. Once again, as the piston is constrained by bearing faces 86, 87 against the bearing faces of pin 95 and drive shaft 98 respectively, pin 95 and drive shaft 98 both rotate in the same direction as piston 81 as it moves back into the cylinder.

It can be seen therefore that the arrangement of

FIGS 17 and 20 comprises a simple, self valving arrangement of six drive cells phased to produce a smooth torque output and oil flow. The group of six cells produce twelve sinusoidal torque pulses per revolution of the cells which when geared by a planetary gear arrangement results in a smooth and powerful torque in the wheel hub.

A particular advantage of the wheel hub motor described above is that an extremely high volumetric capacity per revolution (i.e. torque to weight ratio) is achievable partly through the inherently compact design and partly through the very large gear ratios permissable with this design. Volumetric/weight ratios of 20 cc/kg are readily achievable with this type of motor compared with the present best of prior art motors of around 12 cc/kg. In addition, high rotational speeds are obtainable as a result of relatively small rubbing speeds within the mechanis

- according to the invention together with the excellent porting afforded by the rotating piston. Inertial effects are small as the piston is fully restrained, is of low mass and reciprocates through relatively small distances.

Yet another advantage of the invention, in its various aspects, is that the degree of reciprocatory movement

(displacement) of the second member may be selectively varied in a particular apparatus by varying the distances between the respective rotational axes of the first and second members. In practical application this would permit pump and/or motor displacement to be selectively varied while in mechanical energy or motion transfer apparatus such as impact hammers, hammer drills and the like, the amount of impact energy and/or the stroke of a reciprocatory member may be selectively varied

A major advantage of motors of the type described above is their durability. Because of the internal hydrostatic balance between the components, the motors may be run with very high back pressures without causing a correspondingly high load in the various bearings.

Although the invention, in its various aspects, has been described with reference to pumps and motors, various structural modifications to the broad principles illustrated in FIGS 1 and 2 may be employed as rotational/vibrational drive mechanisms for electric, pneumatic and hydraulic tools such as impact drills, descaling devices, impact hammers, mining tools and the like.

It will be readily apparent to a skilled addressee that many modifications and variations will be possible according to the present invention and that many practical adaptions to the underlying principles will be possible without departing from the spirit and scope of the invention.

Claims

1 . A motion transfer apparatus comprising: - a first member rotatable about a first axis , said first member including at least one first planar working surface lying in a first plane inclined to said first axis; and,

a second member reciprocable along and rotatable about a second axis spaced from and substantially parallel to said first axis said second member including at least one second planar working surface lying in a second plane inclined to said first plane , said second plane being substantially parallel to said first" plane to permit complementary engagement between said first and second working s ur f ace s , s aid motion tran s f er apparatu s characterized in that rotation of said first member about said first axis causes reciprocation and rotation of said second member about said second axis by interaction of respective resultants of reactive forces between said first and second working surfaces .

2. An apparatus as claimed in claim 1 wherein displacement of said second member towards said first member causes rotation of said second and first members about respective rotational axes by interaction of respective resultants of reactive forces between said first and second working surfaces .

3. An apparatus as claimed in claim 1 or claim 2 wherein said first member comprises more than one parallel first working surface .

4. An apparatus as claimed in any preceding claim wherein said second member comprises more than one parallel second working surface .

5. An apparatus as claimed in any preceding claim wherein rotation of said second member is induced by misalignment , of opposing reactive forces between the respective planes of respective working surfaces of said first and second members .

6. An apparatus as claimed in claim 5 wherein misalignment of opposing reactive forces creates a moment which induces a torque in said second member about said second rotational axis.

7. An apparatus as claimed in any preceding claim wherein respective planar working surfaces of said first and second members are inclined relative to respective rotational axes in the range 20º-70º.

8. An apparatus as claimed in claim 7 wherein respective planar working surfaces, of said first and second members are inclined relative to respective rotational axes in the range 35°-60°.

9. An apparatus as claimed in any preceding claim wherein said first and second working surfaces are maintained in operable conjunction by biassing means.

10. An apparatus as claimed in any one of claims 1-8 wherein one of said first and second members comprises a pair of internally opposed working surfaces and the other of said first and second-members comprises externally opposed working surfaces, said internally opposed working surfaces being cooperable with said externally opposed working surfaces to permit interaction of respective resultants of reactive forces between respective pairs of internally and externally opposed working surfaces.

11. An apparatus as claimed in any one of claims 1-9 comprising a fluid pump or motor wherein said first member comprises a swash-plate member rotatable in a housing about said first axis, said second member comprising a piston member reciprocable along and rotatable about said second axis.

12. An apparatus as claimed in claim 12 wherein said second member comprises a pair of piston members located on opposite sides of said swash-plate member, said pair of piston members being reciprocable along and rotatable about said second axis.

13. An apparatus according to any one of claims 1-10 comprising a fluid pump or motor wherein said first member comprises a drive shaft assembly rotatable about an axis within a housing, the drive shaft assembly comprising a drive portion including an opposed pair of inclined drive faces between which is disposed a piston member having inclined driver faces complementary to and co-operable with the inclined drive faces, the annular piston being reciprocable and rotatable within the housing in co-operation with the drive shaft.

14. An apparatus according to claim 13 wherein said inclined driver faces be in parallel planes.

15. An apparatus according to claim 13 wherein said inclined driver faces lie in oppositely inclined planes.

16. An apparatus according to any one of claims 1-10 or claims 13-15 including a plurality of first and second members operatively connected to a common drive means.

17. A fluid pump substantially as hereinbefore described with reference to FIGS 1-6 of the accompanying drawings.

18. A fluid motor substantially as hereinbefore described with reference to FIGS 1, 2, 7-11 and 17-20.

19. A wheel hub motor substantially as hereinbefore described with reference to FIGS 17-20 of the accompany drawings.

20. An apparatus as claimed in any preceding claim wherein linear displacement of said second member may be selectively varied by varying the spatial separation of said first and second axes.