US3695146A - Hydrostatic motor or pump and hydrostatic transmissions - Google Patents

Abstract

This radial-piston, stationary-block, variable displacement motor or pump has special features that enable it to operate far beyond the operating pressure and speed ranges of current hydrostatic motors. Hydrostatic bearings or shoes with multiple recesses at the base of each piston bear on a variable eccentric attached to the output shaft of the motor. The bearing recesses are supplied through two sets of orifices, one set open at all times and the other set open a greater proportion of the time at high-speed, low-stroke operation than at low-speed, high-stroke operation. Valving is done by spool valves driven by a shaftmounted eccentric of fixed eccentricity, adjustable as to phase. The cylinders have spherical bearing surfaces and are so mounted that the entire cylinder can undergo the oscillating motion required to maintain the piston''s centerline alignment with the center of the eccentric.

Description

United States Patent Orshansky, Jr.

[ 1 Oct. 3, 1972 [54] HYDROSTATIC MOTOR OR PUMP AND HYDROSTATIC TRANSMISSIONS [72] Inventor: Elias Orshansky, Jr., San Francisco,

Calif.

[73] Assignee: URS Corporation, San Mateo, Calif.

[22] Filed: April 22, 1970 [21] Appl. No.: 28,289

Related U.S. Application Data [63] Continuation of Ser. No. 657,731, July 31,

1967, abandoned.

[52] U.S. Cl. ..91/478, 91/481, 91/488 [51] Int. Cl ..F0lb 3/10 [58] Field of Search ..91/476, 478, 481, 488; 417/465; 74/55; 92/1 19 [56] References Cited UNITED STATES PATENTS 1,303,795 5/1919 Hicks, Jr. ..74/55 2,347,663 5/1944 Carnahan ..92/1 19 2,456,077 12/ 1948 Orshansky, Jr. ..91/481 2,573,472 10/1951 Martin ..74/790 3,320,902 5/1967 Paschke ..417/213 3,245,354 4/1966 Gregor 103/162 X 1,817,735 8/1931 Clark ..74/571 X 2,288,963 7/1942 Von Tavel ..103/38 X 2,348,958 5/1944 Celio .......74/571 X 2,456,077 12/1948 Orshansky, Jr 103/38 2,573,472 10/1951 Martin ..103/38 2,900,839 8/ 1959 Mackintosh 74/5 71 3,074,294 1] 1963 Wooley ..74/ 675 3,320,902 5/ 1967 'Paschke ..103/178 FOREIGN PATENTS OR APPLICATIONS 403,290 4/1873 ltaly ..417/465 71,772 l/1928 Sweden ..417/465 884,556 12/1961 Great Britain ..103/174 Primary Examiner-William L. Freeh Attorney-Owen, Wickersham & Erickson 1 1 ABSI RACT This radial-piston, stationary-block, variable displacement motor or pump has special features that enable it to operate far beyond the operating pressure and speed ranges of current hydrostatic motors. Hydrostatic bearings or shoes with multiple recesses at the base of each piston bear on a variable eccentric attached to the output shaft of the motor. The bearing recesses are supplied through two sets of orifices, one set open at all times and the other set open a greater proportion of the time at highspeed, low-stroke operation than at low-speed, high-stroke operation. Valving is done by spool valves driven by a shaftmounted eccentric of fixed eccentricity, adjustable as to phase. The cylinders have spherical bearing surfaces and are so mounted that the entire cylinder can undergo the oscillating motion required to maintain the pistons centerline alignment with the center of the eccentric.

17 Claims, 5 Drawing Figures PATENIEnuma :srz

SHEEI 1 BF 5 v INVENTOR ELI ORSHANSKY BY O fl/M ZAMZHM ATTORNEYS PATENTED 3 I97? 3. 695. 146

SHEET 2 OF 5 INVENTOR ELI ORSHANSKY 0M, WM Q ATTORNEYS PATENTEDucI 3 I972 swim a or 5 INVENTOR. ELI ORSHANSKY ATTORNEYS PATENTEUBNB I912 3.695.146

sum 5 [IF 5 INVENTOR. ELI ORSHANSKY ATTORNEYS" HYDROSTATIC MOTOR R PUMP AND HYDROSTATIC TRANSMISSIONS This application is :a continuation'of application Ser. No. 657,731, filed July 31, 1967, now abandoned.

This invention relates to improvements in hydrostatic motors and pumps, especially for use inhydrostatic transmissions.

The hydrostatic pumps and motors of this invention are capable of transmitting far greater power per unit weight andvolume than do conventional units. My new units-have relatively few parts, many of which are interchangeable, have increased efficiency, have a lower temperature rise, and are more reliable andare easier and less expensive to maintain than are conventional units. A truck, tractor or locomotive transmission system employing these units is relatively small, consumes less fuel than comparable conventional units, and has low starting power requirements.

Two features of particular importance in the hydrostatic system are its capacity to endure sustained-overload and its capability of developing maximum drawbar pull with low engine horsepower required.

A hydrostatic drive of this invention includes an engine-driven hydrostatic pump, a hydrostatic motor, pump and motor stroke controllers, a control system, scavenging and supercharge pumps, relief and check valves, braking valves, and an oil reservoir filtration and cooling system. Both the pump and the motor are variable and incorporate step-up or reduction gearing which also serves as the stroke-varying mechanisms for these units. Some scavenging pumps are driven by the output and some by the input. The scavenging pumps driven by the output lubricate the units under the condition of towing to prevent damage. The units may be filled with oil, and provision is available for putting hydraulic units on zero stroke for the towing operation even without the availability of auxiliary power.

The hydrostatic pumps and motors are the heart of the system. They are radial-piston stationary-block units with special features enabling operation far beyond the operating pressure and speed ranges of current hydraulic motors.

Hydrostatic bearings or shoes, with multiple recesses at the base of each piston, bear on a variable eccentric, of spherical outer surface, attached to the output shaft of the motor. The bearing recesses are supplied through two sets of orifices, one set being open at all times and the other being open a greater proportion. of time at high-speed, low-stroke operation than at lowspeed, high-stroke operation; this feature significantly reduces operating temperatures.

Valving is accomplished by means of spool valves also driven by a shaft-mounted eccentric, which has fixed eccentricity but is adjustable in phase. The cylinders have spherical bearing surfaces and are mounted so that the entire cylinder can undergo the oscillating motion required to maintain the pistons centerline alignment with the eccentric. By this means piston sideload forces are essentially eliminated.

Motor stroke control incorporates a planetary speed reduction of approximately 3:1 within the motor envelope in the space that would be required for control elements were there no reduction. It includes two sets of small balancing gears which eliminate output torque forces on the control when the motor transmits maximum torque. Control preload provided by supercharge pressure overcomes torque forces when the motor transmits less than maximum torque. Under all conditions, average torque forces transmittedto the control are zero or small.

The hydrostatic pump is similar to the motor in all details except that the pump control permits fully reversible operation.

The, invention incorporates a novel cooling system to dissipate heat produced by friction of the slippers on the eccentric. Incorporation in the units of variable quantity. oil feed to the hydrostatic bearing pads results in'a very significant decrease in temperature rise with only aslight increase in power loss. Calculations indicate, asan example, that the hydrostatic bearing of a 2,000 hp motor of this invention operating at a maximum speed (without variable feed) would experience a total power loss of 30.1 hp and a temperature rise (from an assumed 200 F supply temperature) of 73.l F. With variable feed, the power loss rises to 35.3 hp, but the temperature rise is limited to 43. 1 F. This is a lower loss and a far smaller temperature rise than would be experienced in a hydraulic motor of more conventional design.

The hydrostatic drive of this invention is capable of exerting torque dependent only onpressure and size of the hydraulic motor and entirely independent of input horsepower into the pump. Thus, maximum drawbar pullimay be developed statically with the prime mover providing only enough horsepower to supply oil under pressure and to makeup for system leakage. Therefore, only a small fraction of maximum prime mover hor sepower is needed to develop maximum starting tractive eflort,and to keepalocomotive or truck rolling at slow speed. In contrast, torque converters or electric drives would require essentially maximum engine power.

The hydrostatic drive of this invention has other advantages over hydrodynamic and electric drive systems. After the hydrostatic drive of this invention has started. a train or truck and it is in motion, the train or truck will continue to move while the engine comes up to full horsepower. Initially it will move slowly, due to the small input horsepower, but, as the engine comes up to full rating, proper speed will be reached. It is unnecessary to overstress the engine and develop unused horsepower (which must be dissipated) in order to start the train or truck.

On the other hand, a hydrodynamic drive can only develop maximum torque by absorbing substantially the full rated horsepower from the prime mover. If prime mover horsepower is reduced, maximum starting torque cannot be developed. Similarly, an electric drive requires large amounts of current to develop full starting torque in the motor, and the prime mover is required to run at high horsepower levels to supply this current.

The hydrostatic drive of this invention is capable of developing output torque with no input torque, as it would do if fed from an accumulator. In contrast, a hydrodynamic drive is properly called a torque converter, in that it converts input torque to higher values of stall torque, as also does an electric drive. For these, maximum input torque is required to generate maximum output torque.

The hydrostatic drive of this invention has a tremendous overload capacity. If maximum operating pressure is exceeded in the system, because of extreme starting conditions, and relief valves pop open, the actual amount of horsepower lost in the transmission need not be great, and the engine would be operating far below its maximum rating; therefore, only a relatively small amount of energy is dissipated through the coolers.

With either a hydrodynamic or electric drive, under conditions of start, tremendous amounts of energy must be dissipated and the engine is asked to operate near its maximum rating. Both factors place a great burden on the cooling system, which must be considerably larger or at least far more critically loaded than the cooling system required for a hydrostatic drive of this invention.

Finally, with a hydrostatic drive of this invention, engine capacity is determined by maximum speed, or maximum speed on a grade, and not by starting conditions. This may enable a more favorable choice in the final drive reduction, and it may be possible to operate a locomotive or other vehicles at a higher speed with the same horsepower, or under conditions which may be more favorable on grades than could be done with other drive systems.

An important point to note in this comparison is that advantages and disadvantages of the various systems stem from their inherent characteristics. It is not possible to improve hydrodynamic or electric drives so that they accomplish the same things as the hydrostatic drive of this invention because, inherently, they develop torque as a function of both input speed and horsepower, while the hydrostatic drive does not.

Very low values of anticipated system weight and volume per unit horsepower are achievable because of the special characteristics of the motors and pumps of this invention. The transmission of this invention is calculated to have output per given displacement of two to four times that of any unit now known, a torque multiplication range of greater than 16:1, nearly twice that achieved by previous drives, an overall efficiency of 80 percent at low speed and 85 percent over the greater part of the range, a reasonable tolerance for dirt, requiring no less than IO-micron filtration, and low production costs, incorporating some 40 percent fewer parts than in previous units, with the parts not posing any unusual manufacturing problems.

Other objects and advantages of the invention will appear from the following description of a preferred form thereof.

In the drawings:

FIG. 1 is a view in longitudinal section of a motorpump embodying the principles of the invention, taken along the line l-1 in FIG. 3.

FIG. 2 is a view in section taken along the line 2-2 in FIG. 1.

FIG. 3 is a view in section taken along the line 3--3 in FIG. 1.

FIG. 4 is a view in section taken along the line 4-4 in FIG. 1.

FIG. 5 is a view in section taken along the line 5-5 in FIG. 1.

The motor-pump of this invention will be described first as a motor, although it is equally capable of acting as a pump. A cylinder block 21 is provided with an annular circumferential inlet manifold 22 to which high pressure oil is supplied by a pump like the motor, but not shown here, and an annular outlet manifold 23 returns the fluid to the pump at low pressure. The manifolds 22 and 23 are connected to valve assemblies 24, by means of respective branch passages 25 and 26. In the specific form illustrated, there are seven such valve assemblies 24, each valve assembly 24 connecting with one piston assembly 30 in a cylinder 31. While seven cylinders 31 are shown, other numbers may be used, such as five, nine, or any other number, an odd number of cylinders 31 producing a smoother pumping cycle than an even number of cylinders 31.

The valves 24 operate in timed relationship with the pistons 30 so that, describing the unit as a motor, when a piston 30 is on top dead center, its associated valve 24 is moving downwardly in the direction shown by an arrow A in FIG. 1 and is about to connect a passage 27 with the passage 26. High pressure oil is then to be admitted to the cylinder 31. As soon as the piston assembly 30 is connected by its valve 24 to the high pressure manifold 22, a downward force is exerted on the piston 30 in the direction of arrow B in FIG. 1, causing an average force F to be produced at the center C of an eccentric 32. This force F rotates the eccentric 32 about the center 33 of its shaft 34, in the direction of the arrow D. The eccentric assembly serves as the output member of the motor and comprises the eccentric 32 and the shaft 34, the two parts enabling one to change the eccentricity or stroke and thereby to vary the speed and torque of the motor at a given inlet pressure.

The valve assemblies 24 move in timed relationship with the pistons 30, and the valve 24 is driven by an eccentric 35, the position of which is adjusted as the center of the eccentric 32 is adjusted, in order to maintain a constant relationship of between the two. The means for adjusting the main eccentric 32 about the shaft 34 and the valve eccentric 35 about this shaft will be described further in this description. Before proceeding with the description of the stroke-adjusting mechanism and the valve-timing mechanism, the valve assemblies and then the piston assemblies will be described in detail.

Each valve assembly 24 includes a sleeve 40, which fits in a bore 41 in the cylinder block casting 21 and has a port 42 connecting with the passage 25, a port 43 connecting with the passage 26, and a port 44 connecting with the passage 27 which leads to the piston assembly 30. The valve sleeve 40 is located within the casting 21 by means of a pin 45.

A valve 46 is reciprocated in the sleeve 40. The valve 46 has an extension 47 which acts as a pin, by means of which the valve 46 is driven from the eccentric 35. Each valve pin 47 has a shoe 48 fitted on it, which is free to oscillate, and all these shoes 48 fit into a valve drive track 49, which surrounds the eccentric 35. The valve drive track 49 is pinned to one of the valves 46 but not to the others, to give freedom of motion and at the same time the proper registry of lubricating holes.

The valves 46 and valve sleeve assembly 24 may be constructed in any of several different ways. In a first such way, the valve sleeve 40 may be made flexible so that it will fit within the bore 41 but not with a press fit. The parts may be so proportioned that the pressure acting on the outer periphery of the sleeve .40 tends to compress the sleeve 40 as pressure rises, thereby reducing the clearance between the sleeve 40 and the valve 46. This may give very good control of leakages at low speeds and the necessary flow for cooling and lubrication at high speeds. A second way of constructing the valve sleeve 40 is to have seals between it and the bore 41 to prevent leakages from one passage to the other. A third way is to press-fit the sleeve 40 into the bore 41 while affording considerable clearance between the sleeve and the valve. In this case it is necessary to put piston rings on the valve 46 and so to position the piston rings that the splits of the rings always ride along the struts in the valve sleeve. This may be the best construction from a practical standpoint.

Each piston assembly generally indicated at 30, comprises a member 50, which has spherical sealing surfaces, 51 and 52 and fits into a seat 53. Within the member 50 is a bore 54 in which a piston 55 reciprocates up and down. The member 50 is prevented from separating from the member 53 by a ring 56 which is urged toward the spherical surface 52 by a spring 57. The piston 55 terminates in a pad 58 on its lower end, as shown in FIG. I. The bottom surface 5 9 of this pad 58 is spherical and matches the spherical outer surface 60 of the eccentric 32. The piston pad 58 has two pairs of recesses 61 and 62 each one of which is fed by a flow restrictor 63 or 64, each preferably comprising an orifice of a capillary tube 65 or 66. One restrictor 63 is shown feeding a recess 61, and the other restrictor 64 is shown feeding a recess 62. This construction enables the piston 30 to operate as a hydrostatic bearing on the surface of the eccentric 32.

Referring to FIG. 3, the pistons 30 assume a different angle with respect to the casting 21 as rotation of the eccentric 32 occurs. However, the force exerted by the piston 30 on the eccentric 32 is always on the center line E of the piston 30, and there are no side forces produced on the piston 30 within the bore 54 of the spherical member 50 as a function of the driving torque. The reaction force, as shown by arrow R, which is a result of the oil pressure pushing a piston 30 radially inwardly, tends to push its member 50 radially outwardly, as shown in FIG. 1. This force is counteracted by the pressure acting around the circumference of the member 50, which tends to push the member 50 radially inwardly. By properly selecting the areas, it is possible to balance this member 50 to any desired degree, and the driving force of the piston 30 results in a balanced hydraulic reaction and does not fall on the piston 30 itself as a side load. This is extremely important, because at high torque and high speeds in units of large horsepower, piston sideload is always a limiting point of performance.

The piston 30 also bears on the spherical eccentric 32, thereby eliminating any deflection or misalignment problems in this preferred construction. However, it is not always necessary to limit the construction to a spherical eccentric 32, for a cylindrical member with spherical bearing surfaces could be used and for some uses a cylindrical outer surface may be used, although this is less desirable.

The piston pads 58 are restrained from separation away from the spherical eccentric surface 60 by rings 67 which are retained in position by plates 68. The pad 58 on the lower end of the piston 30 acts as a hydrostatic bearing. The recesses 61 and 62 are each fed by one orifice or restrictor 63 or 64. The orifice 63 communicates with the pressure within the cylinder space at all times. The orifice 64 communicates with a groove 69 and is plugged on top of the piston 30 by a plug 70. As the piston 30 reciprocates up and down, the groove 69 comes in contact with and out of contact with a groove 71 located in the spherical member 50 and connected to the pressure in the piston space 72 by drilled passage 73.

As the piston 30 moves up and down, the groove 69 passes the groove 71, and therefore the flow through the orifice 64 is interrupted. However, as the stroke of the piston 30 decreases, the groove 69 is in contact with the groove 71 a greater proportion of the time. Therefore the flow through this orifice becomes greater. Eventually, in minimum stroke, the groove 69 is at all times connected with the groove 71, and in that case the flow through the restrictor 64 is constant. The effect of this is as follows: At full stroke the motor operates at low speed and does not need as great a quantity of fluid flow as is needed at high speed. Therefore the restrictor 63 is sufficient to feed it, and a greater flow of oil would produce greater leakage. However, as the stroke decreases and the speed increases, more and more fluid is necessary to lubricate and cool the bearings, and since the pressure drops with increase in speed, the effect of leakage decreases. Therefore, introducing a variable orifice configuration enables the motor to go to far higher speeds at an acceptable temperature rise than would be possible with a single orifice. This is a very valuable feature of the device.

In order to vary the stroke of the pistons 30, and therefore the displacement of the motor, the outer eccentric 32 is rotated about the shaft 34. As seen in FIG. 3 the center of rotation is at 33. Any place in between the stroke will be reduced from its maximum. The force shown by arrow E in FIG. 3 produces a torque on the eccentric about the point G, and the reaction at this point produces a torque on the shaft 34 about the point 33. The direction of force as shown by arrow C is not always constant, as three or four cylinders are under pressure at any time, and the effect is that this force swings through an arc of approximately 24 to 25, producing a torque fluctuation in the eccentric 32 and the shaft 34 about the points G and 33.

It is essential to be able to adjust the eccentric position without adverse effect of this fluctuation of force. For that reason, a novel construction is resorted to. The eccentric 32 has on its end a gear which meshes with an internal gear 81 and therefore transmits its torque to it. The shaft 34 has splined to it a gear member 82 by means of spline 83 and transmits its torque to that member. Thus the torque from the motor proceeds along two paths. :The gear 81 is made integral with the sleeve member 40, on the outer face of which are gear teeth 84. A bearing 85 supports the shaft load, and a bearing 86 supports the load between the inner shaft and the outer shaft. Since normally there is no rotation at that point except for control adjustment, this bearing 86 can be much lighter.

Meshing with the gear 84 are three pinions 87 (FIG. 4). These pinions are supported on pins 88 which are carried in a carrier 89. The carrier 89 is the final output member of the motor and drives the final drive by means of spline 90. The planet pinions 87 also mesh with an internal gear 91. Normally this gear 91 is stationary, but it can be adjusted through an angle for control of eccentricity. The gear 82 also meshes with three planets 92, which are supported on the pin 88 and the carrier 89. The planets 92 and 87 are independent of each other, and the planets 92 mesh with an internal gear 93 which is also normally stationary but is adjustable through a small arc for controlling the eccentricity. Thus, by adjusting the position of the gears 91 and 93, the position of the eccentric 32 is adjusted with respect to the shaft 34, independently of the rotation of the eccentric and shaft members.

The control gearing also acts as a reduction gear, since, in the configuration shown, the carrier 89 runs at approximately 3:1 speed reduction with respect to the shaft 34.

The torque transmitted by the gears 82 and 84 react on the ring gears 93 and 91, acting on both gears in the same direction. In order to balance out this force, the gears 94 and 95 (FIG. 4) are provided, which mesh with externally cut gears on part of the outer circumference of the ring gears and also mesh with each other. Thus, if the torque force on the. gear 93 acts in the direction of arrow II, this force tends to rotate the gear 94 in the direction of arrow J, and the torque force acting on the gear 91 in the direction of K tends to rotate the gear 93 in the direction of arrow L. These torque forces cancel each other out; so there is no tendency for the gears 93 and 91 to move under the influence of equal torque.

In order to obtain adjustment of these gears, each gear 93 and 91 is equipped with pistons 96 and 97. These pistons 96 and 97 are fed by control pressure and can rotate the gears 93 and 91 through the necessary are to achieve control. Since, in certain settings of the eccentric, the torque is not evenly distributed between the gear 92 and the gear 91 but fluctuates from one to the other to a lesser or greater extent, thepistons 96 and 97 produce a preload in conjunction with similar pistons attached to gear 91. In this manner the gear teeth of the gears 94 and 95 and the points where they mesh with the gears 93 and 91, are preloaded, and there is no takeup of backlash during the transfer of portion of the torque from the gear 93 to the gear 91 and vice versa. Hence, the control function of this device'is not dependent upon transmitted torque.

The balance of the eccentrics will now be described (see FIG. 3). The center of mass of the eccentric 32 and of the pistons lies at the point C, and since the point C is rotated about the point G, this center of mass swings around that point. The eccentric 32 is equipped with counterweights 100 and 101, the moments of which are such as to equal the moment of the unbalance of the eccentric 32 and the piston assembly 30. Therefore, irrespective of the rotation of the eccentric 32 about the point G and due to the presence of the counterweights 100 and 101, the total mass of the counterweight, eccentric, and piston assembly will always remain at the point G. However, since-the point G is itself offset from the point 33, which is the center of rotation, it is necessary to balance the aforementioned mass by a still additional set of counterweights 102 and I 103. Thus the counterweights 102 and 103 always balance the total mass of the adjustable eccentric and counterweights which is centered at the point G, irrespective of rotation of the eccentric 32 for adjustment purposes. This makes the control insensitive to centrifugal forces as well, so that there is no effect of reaction on the control as a function of speed.

It was stated previously that the valve eccentric always has to be out of phase with the line connecting the point C with the point 33 in FIG. 2, this line representing the crank arm of the eccentric 32 in any position of adjustment so that as eccentric 32 is rotated about shaft 34, the valve eccentric 35 likewise has to be rotated about the concentric portion of the shaft to maintain this relationship. The mechanism by which this is accomplished is shown in FIGS. 5 and l. A cam ring is fastened to the eccentric 32, a section through the rim of that ring 110 being shown in FIG. 5. The center of that ring 110 is at the point C, which is the same point as shown in FIG. 3. Consequently, the center of this ring rotates about the point G. This cam ring 110 fits into a slot 111, which is cut in counterweight. 112, the counterweight 112 being attached to the eccentric 35 by means of bolts 1 13. On the opposite side of this eccentric 35 is the counterweight 114, and the total mass of these two counterweights l 12 and 1 14 balances the mass of the valve and eccentric assembly.

The counterweights and valve eccentric assembly rotate about the point 33 since they are mounted on the concentric portion of the shaft 34. Therefore as rotation of the eccentric 32 occurs, the cam 110 rotates with it and, by interacting with the slot 111 through bearing blocks 115, causes an adjustment of the eccentric which drives the valves. The geometry of this construction is such that the valve eccentric will effectively remain 90 away from the crank radius of the main eccentric in all positions of adjustment.

To those skilled in the art to which this invention relates, many changes in construction and widely differing embodiments and applications of the invention will suggest themselves without departing from the spirit and scope of the invention. The disclosures and the description herein are purely illustrative and are not intended to be in any sense limiting.

Iclaim:

l. A hydraulic mechanism comprising a support, a cylinder mounted for universal movement in the support, a piston reciprocable with sliding contact in the cylinder means for admitting and exhausting fluid from the chamber formed between the piston and cylinder, means for maintaining the hydraulic forces acting on said cylinder in a balanced state at all times, a crank, means on the crank in engagement with the piston for mounting the piston in sliding contact with the crank, recesses in the piston portion in sliding contact with the crank, means for supplying lubricant continuously to each of the recesses, and additional means for selectively supplying lubricant continuously or intermittently or to cut off the supply of lubricant to each of the said recesses.

2. A device as defined in claim 1 wherein the mounting of the cylinder in the support comprises a spherical surface on the cylinder and socket means in the support engaging said spherical surface at zones on opposite sides of a diametral plane through said spherical surface, and further including means for supplying fluid to a portion of the zone on the side of said diametral plane opposite the crank and at a pressure equal to that in the chamber whereby to substantially counteract the fluid pressure acting between the piston and cylinder urging the cylinder against the socket means. 7

3. A device as defined in claim 1 wherein the means for supplying lubricant continuously comprises a conduit in the piston connecting each of said recesses to the end of the piston in the cylinder.

4. A hydraulic mechanism comprising a support, a cylinder mounted for universal movement'in the support, a piston reciprocable with sliding contact in the cylinder, means for admitting and exhausting fluid from the chamber formed between the piston and cylinder, means for maintaining the hydraulic forces acting on said cylinder in a balanced state at all times, a crank, means on the crank in engagementwith the piston for mounting the piston in sliding contact with the crank, recesses in the piston portion in sliding contact with the crank, means for supplying lubricant continuously to each of the recesses comprising a conduit in the piston connecting each of said recesses to the end of the piston in the cylinder, and additional means for selectively supplying lubricant continuously or intermittently or cutting off the supply of lubricant to each of said recesses, comprising a supply conduit in the wall of the cylinder, a conduit in the piston connecting each of said recesses with a wall of the piston, and means for varying the throw of the crank, whereby the conduit in the piston is continuously in register with the supply conduit, or continuously out of register with said supply conduit, or is intermittently in register with the said supply conduit.

5. A hydraulic mechanism comprising a support, a cylinder mounted for universal movement in the support, a piston reciprocable with sliding contact in the cylinder, means for admitting and exhausting fluid from the chamber formed between the piston and cylinder, means for maintaining the hydraulic forces acting on said cylinder in a balanced state at all times, a crank, and means on the crank in engagement with the piston for mounting the piston in sliding contact with the crank, said crank being an eccentric and further including means for angularly adjusting the eccentric whereby to vary the travel of the piston in the cylinder.

6. A hydraulic mechanism comprising a support, a plurality of cylinders angularly spaced in the support and mounted for universal movement therein, an eccentric rotatably mounted in the support, a piston reciprocably mounted in each cylinder in sliding contact therewith and in sliding engagement with the eccentric, means for admitting and exhausting fluid in timed relation to and from the chambers formed between each of the pistons and cylinders, means for maintaining the hydraulic forces acting on said cylinder in a balanced state at all times, and means for angularly adjusting the eccentric to adjust the travel of the pistons in the cylinders.

7. The hydraulic mechanism of claim 6 having means for maintaining the hydraulic forces acting on said piston in a balanced state at all times.

8. A device as defined in claim 6 wherein the means for angularly adjusting the eccentric comprises a gear train and means for drivin said, train.

9. A device as defined i n claim 6 wherein the means for admitting and exhausting fluid comprises a valve for each piston and cylinder, a valve eccentric rotatably adjustably mounted on the shaft, and means connecting the valves to the said valve eccentric to be driven thereby.

10. A device as defined in claim 9 wherein the means connecting the valves to the valve eccentric comprises a track slidably mounted on the valve eccentric, means slidably connecting each valve to the track, and means pivotally securing only one of said valves to the track.

11. A device as defined in claim 9 wherein the means for adjusting the valve eccentric comprises means defining a slot secured to the first mentioned eccentric and means defining a tongue secured to the valve eccentric and extending into the slot.

12. A device as defined in claim 6 wherein the eccentric comprises a shaft with an eccentric portion and an annulus having a spherical surface eccentrically mounted on the eccentric portion in sliding engagement with each piston.

13. A device as defined in claim 9 further including balance weights secured to opposite axial sides of the first mentioned eccentric to dynamically balance the eccentric and pistons for rotation about the axis of the eccentric portion of the shaft and other balance weights secured to the shaft on opposite sides of said first mentioned eccentric and the balance weights secured thereto to dynamically balance the mass of the said first mentioned eccentric and balanceweights for rotation about the axis of the shaft.

14. A device as defined in claim 12 wherein the means for angularly adjusting the eccentric comprises a gear train connected to the annulus, a gear train connected to the shaft, means connecting said trains to a common shaft, and separate means for driving each train.

15. A device as defined in claim 14 wherein the separate means comprises an internal gear connected to each of the trains and power means connected to each of said internal gears for adjustably rotating each of said internal gears.

16. A device as defined in claim 15 further including means for locking the internal gears against rotation by reaction of the gear trains.

17. A device as defined in claim 16 wherein the locking means comprises separate gears in meshing engagement with each internal gear and each other.

Flo-1050' UNITED STATES PATENT OFFICE (5/69) m a q a 1 CERTEFICATE OF CECTION Patent No. 3,695,146 Dated October 3, 1972 Inventor(s) Elias O'rshansky, Jr. I

It is certified that error appears in the above-identified patent and that said Letters Patent are hereby corrected as shown below:

Left-hand column of first (unnumbered) page of the patent, in TY the heading, after "Assignee:" "URS Corporation" should read --URS Systems Corporation--. Right-hand column of first (urn-- numbered) page in the first line under foreign patents or applications, '4/1873" should read -=-4/l9-+3--; second line under foreign patents or applications, "1/1928" should read --ll/l928-. Column 6, line 41, "arrow E in FIG. 3" should read --arrow F in FIG. 3--.

Signed and sealed this 17th day of April 1973 (SEAL) Attest:

EDWARD M.FLETCHER,JR. ROBERT GOTTSCHALK Attesting Officer Commissioner of Patents

Claims (17)

1. A hydraulic mechanism comprising a support, a cylInder mounted for universal movement in the support, a piston reciprocable with sliding contact in the cylinder means for admitting and exhausting fluid from the chamber formed between the piston and cylinder, means for maintaining the hydraulic forces acting on said cylinder in a balanced state at all times, a crank, means on the crank in engagement with the piston for mounting the piston in sliding contact with the crank, recesses in the piston portion in sliding contact with the crank, means for supplying lubricant continuously to each of the recesses, and additional means for selectively supplying lubricant continuously or intermittently or to cut off the supply of lubricant to each of the said recesses.

2. A device as defined in claim 1 wherein the mounting of the cylinder in the support comprises a spherical surface on the cylinder and socket means in the support engaging said spherical surface at zones on opposite sides of a diametral plane through said spherical surface, and further including means for supplying fluid to a portion of the zone on the side of said diametral plane opposite the crank and at a pressure equal to that in the chamber whereby to substantially counteract the fluid pressure acting between the piston and cylinder urging the cylinder against the socket means.

3. A device as defined in claim 1 wherein the means for supplying lubricant continuously comprises a conduit in the piston connecting each of said recesses to the end of the piston in the cylinder.

4. A hydraulic mechanism comprising a support, a cylinder mounted for universal movement in the support, a piston reciprocable with sliding contact in the cylinder, means for admitting and exhausting fluid from the chamber formed between the piston and cylinder, means for maintaining the hydraulic forces acting on said cylinder in a balanced state at all times, a crank, means on the crank in engagement with the piston for mounting the piston in sliding contact with the crank, recesses in the piston portion in sliding contact with the crank, means for supplying lubricant continuously to each of the recesses comprising a conduit in the piston connecting each of said recesses to the end of the piston in the cylinder, and additional means for selectively supplying lubricant continuously or intermittently or cutting off the supply of lubricant to each of said recesses, comprising a supply conduit in the wall of the cylinder, a conduit in the piston connecting each of said recesses with a wall of the piston, and means for varying the throw of the crank, whereby the conduit in the piston is continuously in register with the supply conduit, or continuously out of register with said supply conduit, or is intermittently in register with the said supply conduit.

5. A hydraulic mechanism comprising a support, a cylinder mounted for universal movement in the support, a piston reciprocable with sliding contact in the cylinder, means for admitting and exhausting fluid from the chamber formed between the piston and cylinder, means for maintaining the hydraulic forces acting on said cylinder in a balanced state at all times, a crank, and means on the crank in engagement with the piston for mounting the piston in sliding contact with the crank, said crank being an eccentric and further including means for angularly adjusting the eccentric whereby to vary the travel of the piston in the cylinder.

6. A hydraulic mechanism comprising a support, a plurality of cylinders angularly spaced in the support and mounted for universal movement therein, an eccentric rotatably mounted in the support, a piston reciprocably mounted in each cylinder in sliding contact therewith and in sliding engagement with the eccentric, means for admitting and exhausting fluid in timed relation to and from the chambers formed between each of the pistons and cylinders, means for maintaining the hydraulic forces acting on said cylinder in a balanced state at all times, and means for angularly adjusting the eccentric to adjust the travel of the pistons iN the cylinders.

7. The hydraulic mechanism of claim 6 having means for maintaining the hydraulic forces acting on said piston in a balanced state at all times.

8. A device as defined in claim 6 wherein the means for angularly adjusting the eccentric comprises a gear train and means for driving said train.

9. A device as defined in claim 6 wherein the means for admitting and exhausting fluid comprises a valve for each piston and cylinder, a valve eccentric rotatably adjustably mounted on the shaft, and means connecting the valves to the said valve eccentric to be driven thereby.

10. A device as defined in claim 9 wherein the means connecting the valves to the valve eccentric comprises a track slidably mounted on the valve eccentric, means slidably connecting each valve to the track, and means pivotally securing only one of said valves to the track.

11. A device as defined in claim 9 wherein the means for adjusting the valve eccentric comprises means defining a slot secured to the first mentioned eccentric and means defining a tongue secured to the valve eccentric and extending into the slot.

12. A device as defined in claim 6 wherein the eccentric comprises a shaft with an eccentric portion and an annulus having a spherical surface eccentrically mounted on the eccentric portion in sliding engagement with each piston.

13. A device as defined in claim 9 further including balance weights secured to opposite axial sides of the first mentioned eccentric to dynamically balance the eccentric and pistons for rotation about the axis of the eccentric portion of the shaft and other balance weights secured to the shaft on opposite sides of said first mentioned eccentric and the balance weights secured thereto to dynamically balance the mass of the said first mentioned eccentric and balance weights for rotation about the axis of the shaft.

14. A device as defined in claim 12 wherein the means for angularly adjusting the eccentric comprises a gear train connected to the annulus, a gear train connected to the shaft, means connecting said trains to a common shaft, and separate means for driving each train.

15. A device as defined in claim 14 wherein the separate means comprises an internal gear connected to each of the trains and power means connected to each of said internal gears for adjustably rotating each of said internal gears.

16. A device as defined in claim 15 further including means for locking the internal gears against rotation by reaction of the gear trains.

17. A device as defined in claim 16 wherein the locking means comprises separate gears in meshing engagement with each internal gear and each other.

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