US3188973A - Hydraulic pumps and motors - Google Patents

Description

June 15, 1965 FlRTH ET L 3,188,973

HYDRAULIC PpMPs AND MOTORS Original Filed April 14, 1960 s She ets-Sheet 1 INVE N TORS June 15, 1965 D. FIRTH ETAL 3,188,973

HYDRAULIC PUMPS AND MOTORS Original Filed April 14, 1960 5 Sheets-Sheet 2 0 INVENTORS ud/w! BY (I ATTORNEY June 15, 1965 D. FIRTH ETAL 7 HYDRAULIC PUMPS AND MOTORS Original Filed April 14, 1960 3 Sheets-Sheet 3 II 125 52/ 0 I18 126 IOI Q INVENTORS Mail 754% M R M United States Patent Ofi ice 3 ,188,913 Patented June 15, 1955 3,188,973 HYDRAULIC PUMPS AND MOTORS Donald Firth and James 1). Hamilton, Glasgow, Scotland,

assignors to Council for Scientific and Industrial Research, London, England, a body corporate of the United Kingdom Original application Apr. 14, 1960, Ser. No. 22,335, new Patent No. 3,120,816, dated Feb. 11, 1964. Divided and this application Oct. 24, 1963, Ser. No. 318,679 Claims priority, application Great Britain, May 1, 1959, 14,890/59 1 Claim. (Cl. 103-474) This application is a division of application Serial No. 22,335 of Donald Firth and Roger H. Y. Hancock filed April 14, 1960, now US. Patent No. 3,120,816 issued February 11, 1964, entitled Hydraulic Pumps and Motors.

This invention relates to hydraulic machines of the positive displacement type, whether pumps or motors, in which the reaction to the thrust exerted on a piston by the fluid pressure in a cylinder is taken through a slipper or shoe on the piston rod working on a tilted or eccentric reaction surface. The machine may be a swash plate machine, in which the reaction surface is constituted by the inclined surface of the swash plate, or a radial cylinder machine in which the reaction surface is a relatively fixed cam or eccentric on the axis of the machine shaft.

One source of loss of efficiency in positive displacement pumps and motors in which the piston rod bears through a slipper against an inclined or eccentric surface is that, especially at low speeds, the slipper tends to tilt, and this materially increases the risk of metal-to-metal contact and high friction.

A positive displacement hydraulic motor or pump according to the present invention has each piston and slipper assembly designed to operate with a high degree of hydrostatic balance of pressures between the piston and the slipper bearing surface, i.e. the pressure of a lubricant film at the slipper surface is substantially a direct function of the bearing load. In this way, metal-tometal friction and the tendency of the slipper to tip or tilt are materially reduced or even eliminated.

To this end, the piston and slipper assembly is conveniently designed so that oil at the working pressure in the cylinder at any given instant is available at the slipper bearing surface which engages the reaction membereccentric, swash plate, or the likeand which controls the reciprocatory motion of the piston.

Preferably the piston and the slipper, together With any intervening connecting rod, are drilled through in register to provide a continuous oil duct which includes a metering restriction at or near the end adjacent the Working face of the slipper, and the latter had oil-retaining recesses adjacent its normally leading and trailing edges from which oil can leak at a controlled rate dependent on the resistance of the metering restriction.

Practical embodiments of the invention will now be particularly described, by Way of example only, with reference to the accompanying drawings in which:

FIG. 1 is an axial cross-section of a swash plate pump or motor, all but one piston and cylinder being omitted for convenience;

FIG. 2 is an enlarged sectional view of a piston and cylinder;

FIG. 3 is a plan view of the slipper in FIG. 2;

FIG. 4 is a fragmentary sectional view similar to FIG. 2 showing a detail modification;

FIG. 5 is a view similar to FIG. 3 of a modified form of slipper;

FIG. 6 is a sectional view on the line VIVI of FIG. 5 and FIG. 7 is a radial fragmentary section through a radial cylinder eccentric type pump or motor.

The pump illustrated in FIG. 1 consists of a main frame having front and back end plates 1, 2 clamped by four pillars (not shown). Each plate 1, 2 carries a journal bearing 4, 5 respectively for a short rigid drive shaft 6 Adjacent the bearing 5 in the back end plate 2, shaft 6 is formed with a locking taper section 7 on which is locked a cylinder block 8. This block is drawn up on the taper by a back-nut 9 on the shaft. The cylinder block 8 contains a number of cylinders 10 whose axes are mutually inclined inwards towards the back end plate 2. A piston 11 in each cylinder is reciprocable under the control of a normally fixed swash plate 12 carried on trunnions (not shown) by which it can be angularly adjusted on an axis normal to the shaft 6. The swash plate 12 has a central conical aperture 14 through which the shaft 6 passes, the dimensions of this aperture being sufficient to allow for adjustment of the angle of the swash plate to the shaft 6.

The working face of the swash plate is recessed at 15a to accommodate an annular bearing pad 15 and an annular slipper plate 16. The latter is free to rotate under the frictional drag of slippers 17 each of which is engaged with a respective piston 11. For clarity of illustration in FIG. 1, only one cylinder 10, piston 11 and slipper 17 is shown. The bearing pad 15 is locked against rotation by means of a dowel. Thus, the friction between the piston slippers 17 and the slipper plate 16 causes the latter to tend to follow the slippers 17 around the shaft 6, while the eccentricity of the plate, due to the tilt of the swash plate 12, with respect to the axis of the shaft 6, causes the slippers 17 to trace a path over the working surface of the slipper plate which is not of constant configuration. Thus, wear of the plate is distributed over an area greater than the annulus which would be traced by a single slipper 17 if the plate 16 were stationary.

Each piston 11 consists of a hollow sleeve 54 (FIG. 2) closed at its outer end to form a crown 55 and which is a snug fit in the cylidner 10 and has four external oil control grooves 60, 61, 62, 63. Into the outer or working end 56 of the sleeve 54 is screwed a headed stem 64 having an axial oil feed bore 65 passing therethrough. The outer end of this stem is shouldered at 66 and has a spigot which passes through an oil seal 68 in the piston crown 55. An oil seal 57 prevents leakage of oil between the head 65 of the stem and the end 56 of the sleeve. Coaxially with the bore 65, the crown 55 is drilled through to form an oil duct 67 which opens into a hemispherical socket 58 which forms a seating for a hemispherical pivot 70 integral with the slipper 17. The pivot is retained in the socket 58 by means of a spring ring 59.

The hemispherical pivot 71) is flattened at 71 to leave a small pocket 72 around the end of the oil duct 67 within the seating 58. A radial oilway 73 is drilled through the pivot and leads into a constriction 74 which communicates with a central recess 75 in the face of the slipper 17 (see also FIG. 3). This recess can thus receive oil from the cylinder 10 through the bore 65 in the stem 64, the duct 67 in the piston crown 55, the pocket 72, the oilway '73 in the hemispherical pivot 70, and the constriction 74. Thus it will be seen that the pocket 72 must always be proportioned so that the pivot 70 never closes the duct 67 at a limit position of its angular deflection.

The sliding surface of the slipper 17 also has an annular groove 76 adjacent its periphery, this groove having no communication with the central recess 75, but communicating with the atmosphere through four radial notches 77, 78, 79, 80. The sliding surface of the slipper 17 thus effectively consists of five lands-an inner annulus 81 and four peripheral part- annuli 82, 83, 84 and 85. For convenience of identification on the drawing, these lands are horizontally shaded in FIG. 3.

The purpose of the above-described arrangement of oil recess and grooves is to combat the tendency of a slipper 17 to tip on the swash plate due to friction between the spherical pivot 70 and its seating 58. This tendency increases with increase of thrust between the slipper and the swash plate, which occurs over approximately one half .of each revolution ofthe particular cylinder concerned, and is coextensive with the oil delivery stroke of the piston. Since all the coacting thrust surfaces are fed with oil at delivery pressure (neglecting any pressure drop due to flow resistance in the circuit 65, 57, 72, 73, 74) the load will be transmitted at each bearing point through a film of oil at a pressure which is a direct function of the axial thrust on the piston. Thus, there will be a film of oil between the lands 81 84 and the slipper plate 16, and also between the pivot '70 and its seating 58. The whole piston and slipper assembly 11, 17 is thus substantially hydrostatically balanced.

In practice, there will be a pressure,

81 84 and the slipper plate 16, but this can be kept to a very low value by careful design and' accurate drop in the circuit 65 {74 due to leakage between the slipper lands s at-3,973

In both forms of slipper, the surface areas of the lands, together with their shapes and positions on the slipper surface, are chosen so that the forces between the slipper 17 and the slipper plate 16 on the swash plate 12 during a delivery stroke of the associated piston 11, which forces tend to promote metal-to-metal contact between the parts, are counterbalanced by forces in the oil films which separate them, these latter forces being directly derived from the instantaneous pressure of oil in the cylinder 10.

A similar hydrostatic balance is not quite achieved by the hemisphericaljpivot 70 and coacting seating 58 as shown in FIG. 2, since the projected area of the latter on a plane normal to the piston axis is less than the area piston.

machining of the bearing surfaces. Furthermore, if the clearance between the bearing surfaces at any point increases so that the pressure in the central recess 75 falls,

the hydrostatic balance of the piston and slipper assembly is upset. The pressure on the piston 11 thereupon tends to reduce the clearance until equilibrium is restored. If the slipper 17 tends to tilt about a point on its. leading edge, so that the adjacent arc of the slipper tends to make metal-to-metal contact with the slipper plate 16, the moment of the piston thrust about this point opposes this tilting, so that the system is self-compensating to a considerable degree.

The degree of self-compensation is partly dependent on the cross-sectional area of each notch 77-80 which is presented to the escaping oil, and by careful design of these notches and their interconnecting annular. groove 76 a dashpot action can be induced in which the oil pressure changes are slower than the changes in mechan-' ical thrust, so that oscillation of the slipper 17 is minimized.

FIGS. 5 and. 6 illustrate a modified oil film control arrangement. ranged recesses 75a, 75b, 75c fed with oil through metering ducts 74a, 74b, 740. The recesses are "of sector shape,

. and of equal radius and angle andare surrounded by a separate concentric groove 76 which communicates with atmosphere through three, eq'uiangularly spaced notches 77a, 78a, 79a. This arrangement of recesses and groove forms four lands, the central land 81a resembling in shape a three-spoked wheel and the outer lands.82a,'83a, 84a being part-annular in shape.

The operation of the modified slipper, however, is somewhat different from that of the slipper described with reference to FIG. 3, in. that tilting is now opposed by differential hydraulic pressures across the face of the slipper.

For example, let'it be. assumed that the slipper 17 in FIGS. 5 and 6 is moving over the slipper plate;16 along the line of the section plane marked VIVI inFIG. 5., the notch 79a leading. There will be a tendency for the ends of the lands 83,41 and 84a on either side of the notch 79a to dig into the surface of the slipper. plate 16, while points on the lands 81a, 82a diametrically opposite the notch 79a will tend to lift. As soon as any such displacement commences, the clearance between the land 81a and the slipper plate 16 will increase and. oil will escape into the annular groove 76, and out through, the notches 77a, 78a. Its pressure at thiszonewill accordingly be reduced, while at the diametrically opposite sideof the annular land 81a the clearance will be reduced and, the oil film pressure will be'increased. This change in pressure produces a restoring couple so thatthe modified system is hydrostaticallyselflcompensating. r

Here there are three symmetrically arof the working'face 56 of. the piston. Consequently, if hydrostatic balance is to be achieved at this point also of the piston-slipper assembly, the arrangement of FIG. 4 is adopted. a In this figure, the hemispherical male pivot 70a is formed on the piston 11 and the coacting socket seating 58a is formed on the slipper 17, the projected area of the seating on a plane normal to the piston axis being made equal to the area ofthe working face 56 of the By'this arrangement the surfaces can be proportioned so that a similar balance between mechanical and hydrostatic forces is achieved as at the working face of the slipper 17. I

It is to be understood that, with appropriate design modifications, the pump described above may be used which works in a complementary little end bearing assembly in thepiston'106. The latter reciprocates in a cylinder 107 in a cylinder block 103, and each cylinder has a detachable. head 109 which defines, Withthe piston 106 a compression space 10.

Each slipper 10 2 has the surface which bears on the maineccentric 101 lubricated by oil under pressure from the working space 110 above the piston 106. For this purpose, each piston rod 103 is drilled axially at 111 from the crown of the part-spherical little end 104. Oppositethe point 112 at which the drilling 111 opens through the crown of the little end,'the crown of the pistonltl is pierced by a coaxial bore 113. The. month 114 of this bore atthe part-spherical bearing surface 105 is enlarged to. ensure continuity of the oil flow path between the bore 13 and the drilling 111 at all working angles'between them during each stroke of the piston 106. Alternatively, if preferred, a flat may be machined on the crown of the little end 104 concentric with the drilling 111 for the same purpose.

The lower'end of the drilling 111 is bifurcated, each branch drilling 115., 116 leading to a respective shouldered cavity 117, 118; This cavity has a metering plug 119, 120 screwed th-ereinto, the plug being provided with an axial I duct which includes a short metering section 121, 122 of stantially symmetrically about both the longitudinal and transverse centre lines of this surface.

When the motor is r-unning,'each recess 1 23, 124 be- 7 comes. filled with oil pumped into it from the working space 110 of the cylinder by way of the drilling or 116 andmeteringduct 121 or 122. Since the hydrostatic pressure of the oil in the recess at any instant is of the same order as the pressure in the working space 110 of the cylinder,-the shoe l02 tends to fioat on'an oil film trapped. between it and the bearing. surface of the main eccentric 101. Asthe oil leaks away through the work- 'its metering duct 118 falls.

ing clearance between the shoe and the eccentric, a pressure drop is established across the metering duct 121 or 122 tending to reduce the pressure of oil forming the lubricating film. The thickness of this film therefore tends to be reduced, thus reducing the leakage flow and re-establishing the pressure equilibrium across the metering duct 121 or 122.

This automatic balance mechanism serves not only to maintain a constant thickness of oil film at the working surfaces but also to oppose the tendency for the shoe 102 to tilt. Thus, assuming that at a given point in the travel of a shoe on the main eccentric 101 in the direction of the arrow R, leading edge 125 of the shoe tends to rise, the clearance between the shoe and the eccentric is increased locally and the rate of leakage of oil from the adjacent recess 123 is increased. As explained above, this brings about a reduction in the hydrostatic pressure of oil in that recess and a consequent reduction in pressure of the oil film. At the same time, the trailing edge 126 of the shoe will tend to bite into the eccentric, the clearance between the bearing surfaces on the shoe and the eccentric being thereby reduced. The resistance to the leakage of oil between these surfaces from the recess 120 therefore increases, and the pressure of oil in the recess 120 consequently rises as the pressure drop across Hence, a couple is exerted on the shoe 102 which tends to oppose the couple causing the initial tilting, and the system tends to be self-correcting as well as floating. Hence, the friction between the working parts is kept to a minimum.

Provided that the actual area of the bearing surface on the shoe 102 is kept relatively small, the total friction between the shoe and the eccentric 101 can be kept to a low value. This is partly due to the fact that the area exposed to friction is less, and partly to the fact that the resistance of the oil leakage path between the shoe and the eccentric is reduced, so that a greater volume of oil leaks away from the recesses 123, 124subject only to the control exercised by the metering duct 121 or 122. The higher rate of leakage leads to the establishment of a thicker film of oil which requires less energy to shear than a thin film, and the loss of thrust at the piston due to this leakage is offset by the reduced absorption of power at the friction surface, especially at creep speeds.

The hydrostatically balance floating feature of the shoe thus leads to increased motor efficiency.

The machine illustrated in FIG. 7 is equally capable of acting as a pump when the eccentric 101 is externally driven.

It will be obvious to those skilled in the art that various changes may be made without departing from the scope of the invention and therefore the invention is not limited to what is shown in the drawings and described in the specification but only as indicated in the appended claim.

What is claimed is:

In a positive displacement hydraulic machine, in combination; a radial cylinder; a piston reciprocable in said cylinder; a slipper articulated to said piston; a piston thrust reaction member on which said slipper is slidable; said piston thrust reaction member comprising an eccentric; at least two separate oil retaining recesses symmetrically disposed with respect to the midpoint of the sliding face of said slipper; each of said recesses extending from a point adjacent the center of said sliding face towards the periphery thereof and bounded on all sides by a continuous land; separate ducts communicating between the pressure side of said piston and said recesses; and a metering constriction in each duct for controlling the supply of working fluid to said recess whereby said slipper is restored to its normal operating position by at least one of said recesses when said slipper is moved from said reaction member.

References Cited by the Examiner UNITED STATES PATENTS 2,604,856 7/52 Henrichsen 103-162 FOREIGN PATENTS 827,756 2/60 Great Britain.

LAURENCE V. EFNER, Primary Examiner.

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