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Publication numberUS3895565 A
Publication typeGrant
Publication dateJul 22, 1975
Filing dateFeb 12, 1973
Priority dateFeb 12, 1973
Also published asDE2406659A1
Publication numberUS 3895565 A, US 3895565A, US-A-3895565, US3895565 A, US3895565A
InventorsHenry Schottler
Original AssigneeHenry Schottler
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Variable displacement fluid transducer
US 3895565 A
Abstract
A variable displacement fluid transducer including a plurality of radial pistons arranged in pairs on a rotating piston barrel. Cam tracks circumscribe the pistons and convert the reciprocation of said pistons into rotary motion of a shaft means, or vice-versa. Piston chambers provided in the barrel couple the pairs of pistons, and the cam tracks are adjustable to vary the displacement of the transducer by changing the relative reciprocating movement of the pistons comprising each pair.
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Description  (OCR text may contain errors)

United States Patent Schottler July 22, 1975 [54] VARIABLE DISPLACEMENT FLUID FOREIGN PATENTS OR APPLICATIONS TRANSDUCER 1,347,202 11/1963 France 91/175 [76] Inventor: Henry Schottler, 8008 Country Club Ln., North Rlverside, I11. 605 Primary ExaminerWilliam L. Freeh 22 Filed; 12, 1973 Attorney, Agent, or FirmMelvin F. Jager [21] Appl. No.: 331,549

[57] ABSTRACT 52 US. Cl. 91/492 A variable displacement fluid transducer including a 51 Int. Cl. FOlb 13/06 plurality of radial Pistons arranged in P On a rotat- [58] Field of Search 91/472, 417/215, 486 ing piston barrel. Cam tracks circumscribe the pistons and convert the reciprocation of said pistons into ro- [56] References Cited tary motion of a shaft means, or vice-versa. Piston UNITED STATES PATENTS chambers provided in the barrel couple the pairs of pistons, and the cam tracks are adjustable to vary the 237L825 Eugen et a] displacement of the transducer by changing the rela- 41962 :22;; 9l/498 tive reciprocating movement of the pistons comprising 3:046:950 7/1962 Smith 91/491 each 3,187,681 6/1965 Firth et al...... 91/492 3,241,463 3/1966 Barrett 91/498 5 7 Drawmg guns .970, Ma 60 .70 J3; 30 76 622 w a) I J 4% 37 '37 Q I C. 45 25,, J 2;: 31

I a 7 c 43,29 g

SHEET PATENTED JUL 2 2 ms VARIABLE DISPLACEMENT FLUID TRANSDUCER BACKGROUND AND GENERAL DESCRIPTION This invention relates generally to a variable displacement fluid transducer, which can be operated as a pump, motor or transmission, for imparting energy to or extracting energy from a fluid under pressure.

There is a demand for well-designed transducer units, for use as motors, pumps, transmissions and the like, which are lightweight, compact, and efficient to manufacture and operate. There is also a need for a transducer that is capable of producing an infinitely variable fluid displacement over a broad operating range. The transducer at any selected displacement within a given range also should operate with uniform and continuous fluid flow, with minimum fluid losses. Another important feature of an effective fluid transducer is the capability of being operated in a reverse direction without substantially effecting the efficiency and fluid displacement characteristics of the unit.

The purpose of this invention is to provide a fluid transducer which meets the above-described design parameters. The transducer in accordance with this invention can be utilized as a motor, pump or hydraulic transmission. The transducer is compact in design and efficient in operation. Moreover, the present transducer operates with a fluid displacement that can be infinitely varied, in either a positive or a negative direction from a zero flow condition, within a selected displacement range. The transducer therefore is sufficiently versatile to be used in applications which call for varying fluid displacement, speed and torque output requirements.

Briefly described, the transducer in accordance with the present invention includes rotatable shaft means and a cylindrical piston barrel joined for rotation with the shaft means. Sets of radial pistons comprising a plurality of piston balls are positioned uniformly spaced around the barrel. The pistons are arranged to define coupled pairs of piston balls and a piston chamber is provided in the barrel for each piston pair. In operation, the relative reciprocation of the piston balls comprising each pair displaces fluid from the associated piston chamber. Cam track means circumscribe the pistons and include a plurality of uniformly spaced lobes projecting inwardly toward the piston barrel. The cam lobes are in frictional engagement with the piston balls so that relative rotation of the barrel with respect to the cam track means orbits the balls in rolling engagement with the cam lobes. The piston balls and lobes thereby cooperate to reciprocate the pistons radially and to displace fluid from the piston chambers as a function of the relative reciprocating movement of the piston balls comprising each piston pair. Control means are provided to shift the cam track means circumferentially, to vary the relative position of the cam lobes and thereby vary the fluid displacement of the transducer by changing the relative reciprocation of the piston balls comprising each pair. Porting means are further provided for directing fluid sequentially to and from each of the piston chambers during the operation of the transducer.

EXEMPLARY EMBODIMENTS Further objects and features of the present invention will become more apparent from a description of several preferred embodiments thereof, taken in conjunction with the accompanying drawings, in which:

FIG. 1 is a cross-sectional elevational view of one embodiment of variable displacement piston pump having the features and advantages of the present invention;

FIG. 2 is a cross-sectional end view of the pump, taken along the line 22 in FIG. 1;

FIG. 3 is a cross-sectional elevational view of a second embodiment of the variable displacement piston pump in accordance with this invention;

FIG. 3A is a removed cross-sectional view of the second pump embodiment, taken along the line 3a3a in FIG. 3;

FIG. 4 is a cross-sectional end view of the second embodiment. taken along the line 44 in FIG. 3;

FIG. 5 is an elevational view, in partial cross-section. of a fluid transmission system embodying the features and advantages of the present invention; and

FIG. 6 is an elevational view, in partial cross-section, of a second embodiment of a fluid transmission embodying the features and advantages of the present invention.

PUMP

The first embodiment of the pump in accordance with this invention is generally indicated by the reference number 10 in FIGS. 1 and 2. The pump 10 includes a generally cylindrical pump housing 12 having removable end cover plates 14 and 16. A bearing 18 mounted in the end plate 16 supports an input shaft 20. A high-pressure sealing ring 22 is provided in a groove in the shaft 20 to seal the interior of the housing 12 from the compartment containing the bearing 18. This arrangement adapts the pump 10 to high pressure applications by preventing any oil loss from the housing 12 around the shaft 20. The ring 22 also prevents contamination of the fluid being pumped by leakage from the lubricant applied to the bearing 18.

The bearing 18 is held in a stationary position in the end plate 16 by means of a sealing and retaining assembly 24. A retaining ring 26 in the assembly 24 locks the assembly in the plate 16. Suitable sealing rings 27a and 27b seal the compartment for the bearing 18 from the atmosphere. The input shaft 20 is restrained from axial movement by a shoulder 20a on the shaft, which abuts against one side of the bearing 18, and a retaining ring 28, which abuts against the opposite side of the bearing.

As seen in FIG. 1, the inner end of the input shaft 20 defines a hollow cylindrical piston barrel 30. For ease in machining and assembly, the piston barrel 30 is preferably an integral part of the shaft 20. In the illustrated embodiment, the piston barrel 30 includes two sets of radial piston cylinders 32 and 34. Each set of cylinders 32 and 34 includes a plurality of cylinders uniformly spaced around the circumference of the barrel 30. In the preferred arrangement, the set 32 includes nine radial cylinders 32a-i, and the set 34 includes nine radial cylinders 34a-i. The set 32 and 34 are also arranged so that the cylinders 32a-i and 34a-i are uniformly spaced in axial alignment, as seen in FIG. 2. Moreover, a plurality of axial bores in the barrel 30 define piston chambers 31 which join the cylinders in the set 32 with the adjacent aligned cylinder in the set 34. As seen in FIG. 1, the adjacent cylinders, such as cylinders 32a and 34a, will thereby cooperate to impart energy to a charge of fluid in the piston chambers 31 during the operation of the pump 10.

In accordance with this invention, each cylinder in the sets 32 and 34 includes a ball piston which is designed to reciprocate as the input shaft 20 is rotated. Accordingly, a set of piston balls 36a-i is positioned within the cylinders 32a-i, and a second set of piston balls 38a-i is positioned within the cylinders 34a-i. In the preferred embodiment, all ball pistons 36 and 38 are the same diameter, and the diameter is selected so that the ball pistons seal the associated cylinders without the need for any piston rings or the like. In addition, the inner ends of the cylinders 32a-i and 34a-i define fluid ports 35ai and 37a-i, respectively; The ports 35 and 37 provide openings for the flow of fluid to and from the cylinders 32 and 34 during the operation of the pump 10. The design of the ports 35 and 37 further arranges the ball pistons 36 and 38 so that a portion of each ball piston extends radially outward from the periphery of the piston barrel 30 during the operation of the pump 10.

The embodiment of the pump illustrated in FIGS. 1 and 2 includes a stationary cylindrical pintle 40. The pintle is press-fit within a flange 13 provided on the left end of the housing 12. Suitable sealing rings 42a and 42b seal the juncture between the pintle and the left end cover 14. The pintle 40 is dimensioned and machined to fit, with a close tolerance, within hollow central core of the piston barrel 30. The barrel 30 thereby rotates around the pintle 40 during the operation of the pump 10. In the illustrated embodiment, the fluid being pumped through the pump 10 is used to lubricate the engaged frictional contact surfaces of the barrel 30 and pintle 40 by squeeze lubrication which is induced by the rotation of the barrel 30 and the radial forces on the piston balls 36 and 38, as described further below.

The pintle 40 includes a series of bores which define inlet and outlet passages for controlling the flow of fluid through the pump 10. The left end cover 14 defines an inlet bore 11 which communicates with a fluid collector chamber 42 provided in the adjacent end of the pintle 40. A plurality of axial inlet channels 44 extend through the pintle 40 to a location adjacent the cylinders 32 and 34 in the piston barrel 30. In the arrangement illustrated in FIGS. 1 and 2, the pintle 40 is provided with four uniformly spaced inlet channels 44a-d in communication with the fluid chamber 42. As seen in FIG. 1, four uniformly spaced radial inlet bores 46 extend through the pintle 40 and connect all four inlet channels 44a-d in fluid communication. The bores 46 are axially aligned with the first set of cylinders 32, and place the cylinders 32a-i in successive communication with the inlet channels 44a-d as the barrel 30 rotates. As seen in FIGS. 1 and 2, a second set of four radial inlet bores 48 is axially aligned with the second set of cylinders 34. The radial bores 48 are in axial alignment with the bores 46, and connect the cylinders 34a-i to the four inlet channesl 44a-d as the barrel 30 rotates.

The pump 10 also defines a fluid outlet bore which communicates with an annular fluid collector chamber 41. The pintle 40 also includes four uniformly spaced axial outlet channels 43a-d which are spaced uniformly between the inlet channels 44a-d. A set of four radial outlet bores 45 are aligned axially with the first set of cylinders 32. The bores 45 are also spaced uniformly in the pintle 40, and the cylinders 32a-i in successive fluid communication with the outlet channels 43a-d as the barrel 30 rotates. An additional set of radial outlet bores 47 is aligned with the second set of cylinders 34, and places the cylinders 34a-i in successive communication with the outlet channels 43a-d as the barrel 30 rotates. The outlet bores 45 and 47 are axially aligned in the pintle 40.

The pump 10 also includes a pair of annular cam members 50 and 60 which control the reciprocation of the ball pistons 36 and 38, respectively, during the operation of the pump 10. As seen in FIG. 1, the cam members 50 and 60 surround the piston barrel 30 and are arranged to space the cam member 50 radially above the set of piston balls 36ai and to space the cam member 60 radially above the set of piston balls 380-1". The internal periphery of the cam members 50 and 60 define concave cam tracks 52 and 62, which engage with the piston balls 36 and 38 respectively. The tracks 52 and 62 are provided with a radius of curvature slightly larger than the radius of the engaged balls 36 and 38 so that the balls will be self-seating within the cam tracks during the operation of the pump 10. The tracks 52 and 62 further define a plurality of uniformly spaced cam lobes which cause the engaged piston balls 36 or 38 to reciprocate, as the piston barrel 30 is rotated. In the preferred arrangement the cam track 52 is provided with four uniformly spaced lobes 52ad, and the track 62 is provided with four uniformly spaced lobes 62a-a'. The lobes are designed to have a uniform amplitude and wave length, and in the preferred embodiment, are contoured to produce a cycloidal displacement of the pistons 36 and 38. This arrangement produces a substantially uniform total fluid flow over the entire range of displacement for the pump 10.

The cam members 50 and 60 are held in a relatively stationary position during the operation of the pump 10. To accomplish this, the pump housing 12 includes a pair of rotatable ring gears 66 and 68 spaced around the cam members 50 and 60, respectively. The ring gears 66 and 68 are restrained from lateral movement by the housing 12. The inner periphery of the gears 66 and 68 include splines or keys 67 and 69, respectively, which mate with corresponding splines or key slots provided on the adjacent outer portion of the cam members 50 and 60. The splines or keys 67 and 69 permit the cam members 50 and 60 to move axially within the housing 12, and further cause the cam members 50 and 60 to rotate in response to the rotation of the associated ring gears 66 and 68.

The pump 10 also includes means to gradiently vary the displacement of the pump by adjusting the circumferential positioning of the cam members 50 and 60. The adjustment means includes bevelled gear teeth 66a and 68a provided on the inner portion of the ring gears 66 and 68. The teeth 66a and 68a mesh with a bevel pinion 70 mounted on the pump housing 12 by means of a shaft 72. Needle bearings 73 support the shaft 72 in the housing 12 and are selected to have a carrying capacity sufficient to absorb the reaction torque load applied to the shaft 72 during the operation of the pump 10. Suitable sealing rings are provided around the shaft 72 to maintain the housing 12 fluid-tight. A control handle 74 projects from the housing 12 and is connected to the shaft 72. As seen in FIG. 1, manual rotation of the handle 74 rotates the pinion 70, and thereby adjusts the circumferential positioning of the ring gears 66 and 68 and the connected cam members 50 and 60. In the illustrated embodiment, movement of the pinion 70 shifts the cam members 50 and 60 an equal distance in opposite directions, and varies the relative axial positioning of the cam lobes 52a-d with respect to the cam lobes 62a-d.

In accordance with this invention, the cam members 50 and 60 are positioned within the housing 12, in a selected relationship with respect to the pintle 40, to provide the pump with an infinitely variable fluid displacement within a broad range in either a positive or negative direction from a zero displacement condition. Thus, the cam members 50 and 60 initially are arranged so that the lobes 52a-d are 45 out of phase with respect to the lobes 62a-d. That arrangement places the bottom center-point of the lobes 52a-d in axial alignment with the top center-point of the lobes 62a-d, and vice versa, as seen in FIG. 2. Furthermore, the cam members 50 and 60 are circumferentially arranged in the housing 12 so that maximum port opening for the pump occurs when one of the piston balls 36 or 38 is engaged with the top center-point of the cam lobes 52a-d or 62a-d, and the other axially aligned piston ball 38 or 36 is engaged with the bottom center-point of the other cam lobe 62a-d or 52a-d. This arrangement defines the zero displacement condition for the pump 10. In the arrangement illustrated in FIG. 2, this zero condition occurs when the piston ball 38a is engaged with the top center-point of the cam lobe 62a, and the axially aligned piston ball 36a is engaged with the bottom center-point between the cam lobes 52a and 52d. At this point in the operation of the pump 10, the balls 36a, 38a and the ports 35a and 37a are radially aligned to the maximum open port condition, with the radial outlet bores 45 and 47 in the stationary pintle 40. The remaining fluid ports 35b-i and 37b-i are at varying positions with respect to the inlet and outlet bores 45-48 in the stationary pintle 40, and correspondingly produce varying conditions of port opening.

All axially aligned piston balls, i.e., 36a-38a, 36b-38b, etc. are engaged with either the top or bottom center-point of the cam lobes 52a-d and 62a-d under the zero displacement condition described above. Since each pair of axially aligned ball pistons (e.g. 36a and 38a) are reciprocating in opposite directions under that condition, the combined pumping force transmitted by each piston pair to the charge of fluid contained within the piston chamber 31 and the associated cylinders (e.g. 32a and 34a) is zero. Upon rotation of the barrel 30, the fluid in the piston chambers 31 under those conditions will be moved only between the associated cylinders 32 and 34, and the resultant fluid displacement of the pump 10 will be zero.

The operating handle 74 is rotated to increase the pump displacement from the above described Zero displacement condition to a maximum displacement condition in either direction of fluid flow. For instance, the handle 74 and the pinion 70 can be rotated to rotate the cam member 50 counter-clockwise and the cam member 60 clockwise, as viewed in FIG. 2. The cam lobes 52a-d are thereby axially displaced with respect to the lobes 62a-d. The rotation of the piston barrel 30 under such conditions will reciprocate the pistons 36 and 38 in a manner which imparts a pumping force to the fluid charge within the piston chambers 31, or imparts a suction force which draws the fluid charge into the chambers 31, depending on the arrangement of the associated porting. The maximum pumping and suction forces, and hence the maximum fluid displacement for the pump 10 for fluid flow in the positive direction, will occur when the cam lobes 52a-d are axially aligned with the cam lobes 62a-d. This maximum condition is created by rotation of the cam members 50 and with respect to each other a resultant of 45 degrees. Since each member rotates, the circumferential movement of each member 50 and 60 to produce such a condition will be 22 /2". In such a position, the aligned piston balls 36 and 38 (e.g. 36a and 38a) are reciprocating in unison and hence produce the maximum pressure flow or suction flow of the fluid in the cylinders 32, 34 and the piston chambers 31. The cam members 50 and 60 can be positioned between the above-described zero and maximum condition to infinitely vary the fluid displacement of the pump 10.

In accordance with this invention, the cam members 50 and 60 can be shifted beyond the zero condition counter-clockwise, in the direction opposite to that described above, to reverse the flow of fluid in the pump 10. The fluid flow is reversed because the sequential relationship between the barrel ports 35, 37 and the inlet ports 46, 48 and the outlet 45 and 47 is reversed. Hence, the ports 46, 48 become outlet ports, and the ports 45, 47 become inlet ports. The pump operates in a similar manner to provide an infinitely variable fluid displacement between the zero flow condition and the maximum flow condition. The zero flow condition again occurs with the pump arranged as shown in FIG. 2. The maximum reverse or negative flow condition would occur when the cam members 50 and 60 are shifted, a resultant 45 (22 /2 each), to align the cam lobes axially; with the lobe 62a aligned with the lobe 52d; 62b aligned with 520; 62c aligned with 52b; and 62d aligned with 52a.

The symmetry of the pump 10 in accordance with this invention permits the pump to operate with equal effectiveness for either clockwise or counter-clockwise rotation of the input shaft 20. The operation of the pump 10 will therefore be described only with respect to the results obtained by rotating the shaft 20 in a clockwise direction, as viewed in FIG. 2.

The clockwise rotation of the shaft 20 causes clockwise rotation of the piston barrel 30 at the same speed as the shaft 20. The rotating barrel 30 drives the piston balls 36 and 38 in a planetary fashion around the stationary pintle member 40. The piston balls 36 and 38 also roll along the annular cam tracks 52 and 62, and the cam lobes 52a-d and 62a-d impart a radial reciprocating'motion to the piston balls 36 and 38. The profile of the lobes 52a-d and 62a-d causes the balls 36 and 38 to follow a cycloidal displacement path.

The pumping effect of the reciprocation of the balls 36, 38 is a function of the relative positioning of the cam members 50 and 60. When the cam members 50 and 60 are aligned in the zero position shown in FIG. 2, the reciprocation of the piston balls 36a-i will oppose the reciprocation of the piston balls 38a-i (e.g. ball 36a will be moving radially outward as the associated ball 38a is moving radially inward.) Since the relative reciprocation of the balls comprising each pair of pistons is zero, the fluid in the common chambers 31 will merely flow between the associated cylinders 32 and 34 under such a condition, and the fluid displacement of the pump 10 will be zero.

Fluid displacement will occur in the pump 10 when the handle 74 is operated to move the cam members 50 and 6 out of the above-described zero position. Maximum positive displacement will occur, for instance, when the handle 74 is rotated to displace the cam members 50 and 60 a resultant of 45 from the aligned zero condition. Since both cam members rotate equally, but in opposite directions, upon rotation of the handle 74, maximum positive displacement will occur in the illustrated embodiment when the cam member 60 is rotated 22 /2" counter-clockwise from the zero displacement condition illustrated in FIG. 2. In such a position, the top center-point of the cam lobe 62a and the top center-point of the cam lobe 52a will be in axial alignment, at a location 22 /2 from vertical in FIG. 2. The remaining cam lobes 6212-5212, 62c-52c, and 62d-52d also will be in axial alignment. The rotation of the piston barrel 30 will thereby cause the piston balls 36 and 38 to reciprocate in unison under such conditions, and each pair of piston balls will provide maximum pumping displacement of the fluid in the associated piston chamber 31 and cylinders 35 and 37.

The operation of the pump under the maximum positive displacement condition is apparent when the pump 10 is visualized as being shifted from the zero displacement condition as shown in FIG. 2, into the maximum displacement condition described above. As the piston barrel 30 is rotated at a constant speed in a clockwise direction, each piston ball 36a-i and 38a-i roll along the cam members 50 and 60 respectively. The lobes 52ad cause the pistons 36a-i to reciprocate radially in the cylinders 32a-i, and the lobes 62a-i cause the pistons 3811-1 to reciprocate radially in the cylinders 34a-i. As the barrel 30 rotates, the pistons 36a-i and 38a-i will be moving radially outward when the associated piston ports 35a-i and 37a-i are in fluid communication with the inlet channels 44a-d through the radial inlet bores 46 and 48. Each outwardly moving piston 36a-i and 38a-i is thereby providing a suction stroke which draws fluid into the common chambers 31 and the associated cylinders 32a-i and 34a-i from the common fluid inlet collector chamber 42. Similarly, the rotation of the barrel 30 causes the pistons 36a-i and 38a-i to move radially inwardly when the associated piston ports 35a-i and 37a-i are in fluid communication with the outlet channels 43a-d through the radial outlet bores 45 and 47. Each inwardly moving piston thereby provides a pressure stroke which pumps the fluid from the chambers 31 and the associated cylinders 32a-i and 34a-i into the common fluid outlet collector chamber 41. As seen in FIG. 2, the arrangement of the pump 10 thereby causes each set of piston balls 36 and 38 (e.g. 36a and 38a) to reciprocate alternately through suction and pressure strokes so that a cycle is completed when the piston barrel 30 rotates 90. Hence, each set of pistons will produce four complete pumping cycles for each revolution of the input shaft 20. As also seen in FIG. 2, the arrangement of the pump 10 causes the piston ball sets 36, 38 (e.g. 36a and 38a) to be in different phases of the pumping cycle at any given instant in the operation of the pump 10. (e.g. one piston set is starting a pumping or suction stroke, another set is ending a pumping or suction stroke, etc.) As a result, the suction and pumping strokes will be staggered, and the output discharge of the pump 10 will be uniform and continuous.

The pump displacement can be varied infinitely between the Zero position and the maximum positive displacement position by adjusting the handle 74 to change the axial alignment of the cam members 50 and 60. The handle 74 also can be adjusted in the opposite direction to create a negative flow of fluid through the pump 10. The negative flow results from rotating the cam member 60 in a counter-clockwise direction, and the cam member 50 in a clockwise direction, as viewed in FIG. 2. By such an adjustment, the circumferential position of the lobes 52a-d and 62a-d is reversed with respect to the fluid channels 430-11 and 44a-d and the radial bores 45-48 on the stationary pintle 40. The piston balls 36 and 38 are moving outwardly through a suction stroke when the associated ports 35 and 37 are in fluid communication with the bores 46, 48 on the pintle 40, and the balls 36, 38 are moving inwardly through a pumping stroke when the ports 35, 37 are in communication with the bores 45 and 47 on the pintle 40.

Accordingly, the pump 10 operates in a manner identical to that described above, but the fluid flow through the pump is reversed, with the bore 15 becoming the inlet bore, and the bore 11 becoming the outlet bore. The maximum negative flow will occur when the cam members 50 and 60 are axially positioned with the bottom center-point of the cam lobe 62a aligned with the bottom center-point of the cam lobe 52d. Such alignment of the lobes 62a and 52d would occur in the illustrated embodiment at 22 /2 to the left of the vertical centerline of the pump 10, as viewed in FIG. 2. The negative pump flow likewise may be infinitely varied by positioning the cam members 50 and 60 at intermediate locations between the zero position and the abovedescribed maximum negative flow position.

FIGS. 3 and 4 illustrate a second embodiment of th variable displacement pump in accordance with this invention. The pump shown in FIGS. 3 and 4 operates in essentially the same manner as the pump 10 illustrated in FIGS. 1 and 2. The pump 100, however, includes side porting which eliminates the need for the stationary pintle 40 incorporated in the pump 10. The side porting design for the pump 100 permits the pump to be more compact, and to be more readily machined and assembled. The side porting also permits the fluid channels in the pump 100 to be readily enlarged, and designed to minimize fluid friction losses.

The pump 100 includes a generally cylindrical pump housing 112 having a removable end cover plate 114. A bearing 118 in the end plate 114 supports an input shaft 120. A sealing ring 122 is mounted in a retaining assembly 124, and seals the pump housing 112, 114. A retaining ring 126 locks the assembly 124 to the end plate 114. Retainer rings 127a and 1271; are mounted in grooves within the shaft 120, and thereby restrain the input shaft from axial movement. As seen in FIG. 3, the inner end of the shaft 120 is rotatably supported in a needle bearing 119 positioned within the housing 112.

The pump 100 also includes an annular rotatable piston barrel 130 which is positioned within the housing 112 intermediate the ends of the shaft 120. Splines 121 on the central portion of the shaft 120 mate with corresponding splines on the piston barrel 130. The splined connection between the barrel 130 and the shaft 120 causes the barrel and shaft to rotate in unison while allowing the barrel to shift laterally so that the pistons carried by the barrel will be self-seating.

The barrel 130 is provided with two sets of radial piston cylinders 132 and 134. Nine cylinders 1320-1 and 134ai are included in each set in the illustrated embodiment. The cylinders 132a-i and l34ai are uniformly spaced around the circumference of the barrel 130, and are arranged in axial alignment. In addition, a plurality of axial bores in the piston barrel 130 define fluid piston chambers 131 which place the adjacent aligned cylinders, such as cylinders 132a and 134a, in fluid communication, and also provide the pump 100 with axial porting. Sets of piston balls 136a-i and 138a-i are positioned within the cylinders l32ai and 134a-i, respectively. The pistons 136 and 138 in the illustrated embodiment are provided with equal diameters, and seal the associated cylinders 134 and 136. The piston barrel 130 also includes fluid ports 135a-i and 13711-1" joining the inner ends of the cylinders 132a-i and 134a-i, respectively, with the associated fluid chamber 131.

The pump 100 illustrated in FIGS. 3 and 4 also includes a porting insert 140 which contains porting channels for directing the flow of fluid through the pump. The insert 140 is annular in configuration, and is positioned around the shaft 120 at one end of the pump housing 112 (the left end in FIG. 3). Sealing rings 141 seal the junctions between the insert 140 and the housing 112, and a plurality of locking pins 142 secure the insert 140 from rotation within the housing 112. The pins 142 permit the insert 140 to move laterally a small distance so that the insert establishes a state of equilibrium when the pump 120 is in operation. A coil spring 143 is positioned around the shaft 120 and extended between the housing 112 and the insert 140 to create a preloading force which biases the insert 140 toward the rotatable piston barrel 130.

The porting insert 140 includes a series of conduits for directing the flow of fluid through the pump 100. An inlet bore 111 provided in the housing 112 communicates with an annular fluid inlet collector chamber 113. The chamber 113 in turn is in fluid communication with a set of four uniformly spaced axial inlet conduits 144a-d extending through the porting insert 140. A second set of axial inlet conduits 145a-d is also provided in the end cover plate 114 in axial alignment with the conduits 144a-d. A plurality of radial and axial fluid channels 146 extends through the end plate 114 and the housing 112 and place the conduits 145a-d in communication with the inlet collector chamber 113. The axial conduits 144a-d and 145a-d are arranged to be in communication with the piston chambers 131 in the barrel 130 as the barrel 130 rotates.

The pump housing 112 further defines an outlet bore 115 which is in fluid communication with an annular fluid collector chamber 116. A set of four uniformly spaced axial outlet conduits 147a-d extend through the insert 140 in fluid communication with the collector chamber 116 and the rotating piston chambers 131 on the pump barrel 130. A second set of axial outlet conduits l48a-d is also provided in the end cover plate 114 in axial alignment with the conduits 147a-d. A plurality of radial and axial fluid channels 149 extend through the end plate 114 and the housing 112 to connect the conduits 148a-d in fluid communication with the outlet collector chamber 116. As seen in FIG. 4, the inlet conduits 144, 145 and outlet conduits 147, 148 define inlet and outlet ports, respectively, which are alternately arranged in a uniform pattern adjacent each side of the barrel 130.

The above-described porting arrangement causes the piston chambers 131 to be placed in fluid communication sequentially with the inlet conduits 144a-d, l45a-d, and the outlet conduits 147a-d, 1480-11 as the piston barrel 130 rotates. The reciprocation of the ball pistons 136, 138, caused by the rotation of the barrel 130, will thereby pump the fluid from the inlet 111 through the outlet 115.

The pump also includes means to minimize the pressure between the rotating piston barrel and the stationary end plate 114 and insert 140. A low pressure between the barrel 130 and the adjacent stationary components is desirable so that the barrel 130 can be rotated at high speed. and the pump can be operated at high fluid pressures. As illustrated in FIGS. 3 and 3A. the interfaces between the barrel 130 and the stationary insert and end plate 114 are provided with an annular pressure relief grooves 117a and 11717. A pair of radial slots 117C (FIGS. 3A and 4) places the outer relief groove 117a in direct fluid communication with the low-pressure interior of the housing 112. The inner relief groove 117b also is in direct communication with the low pressure interior of the housing 112, as seen in FIGS. 3 and 3A. The grooves 117a and 117b reduce the area between the barrel 130, the insert 140 and the end plate 114 to a minimum. Accordingly, the pump 100 includes a minimum area exposed to the high-pressure fluid in the pump 100, so that fluid friction losses are minimized.

The fluid pressure against the rotating piston barrel 130 is further minimized by a pressure balancing device incorporated within each of the conduits 144ad and 147a-d in the insert 140. As seen in FIG. 3, a slidable pressure piston is positioned in each of the conduits 144a-d and 147a-d adjacent the end of the housing 112. The outer end of each pressure piston 180 engages with the housing 112, and a sealing ring 181 maintains the fluid seal of the associated conduit. The inner end of each piston 180 defines a circular pressure area 182 which is selectively larger than the crosssectional area of the conduits 144a-d and 147a-d. The high pressure fluid in the conduits 144a-d and 147a-d exerts an equal pressure force in both axial directions within the cross-sectional area of the conduits 144 and 147. However, the high pressure fluid also exerts an unbalanced axial force (to the left in FIG. 3) against the area A (FIG. 3A) defined by the interface between the barrel 130 and the insert 140. The pressure area 182 of each piston 180 is a selectedamount larger than the cross-sectional area of the conduits 144 and 147 so that fluid pressure against the area 182 creates a reaction force (to the right in FIG. 3) which balances the force created against the area A. The pressure relief grooves 117a and 117b reduce the effective pressure area A. The effective pressure tending to separate the barrel 130 from the insert 140 and plate 114 due to leakage is approximately one-half of the operating pressure.

Therefore, in the preferred arrangement, the balancing area 182 (in excess of the diameter of the conduits 144, 147) is slightly larger than one-half of the opposed area A (FIG. 3A) so that the operation of the pump 100 creates a small fluid-pressure preloading force (to the right in FIG. 3) which assists in maintaining sealing contact between the rotating barrel 130 and the ports in the insert 140 and plate 114 under all loading conditions. Such a fluid preloading force supplements the substantially constant preloading force of the compression spring 143. This arrangement balances the fluid pressure and reaction forces on the insert 140 and barrel 130 during the operation of the pump 100.

A pair of annular cam members 150 and 160 are included in the pump 100 to produce the reciprocation of the ball pistons 136a-i and 1380-1, respectively. The cam members 150 and 160 are annular in configuration, and surround the piston barrel 130 in radial alignment with the ball pistons 136, 138. The inner periphery of the cam members 150 and 160 define arcuate cam tracks 152 and 162 which have a radius of curvature slightly larger than the radius of the engaged piston balls 136 and 138.

The cam tracks 152 and 162 define a plurality of uniformly spaced cam lobes which engage with the piston balls 136, 138 and cause the balls to reciprocate as the barrel 130 is rotated on the embodiment illustrated in FIGS. 3 and 4, four equally spaced lobes 152a-d and 162a-d are included on the cam tracks 152 and 162, respectively. The lobes are uniform in amplitude and wave length, and are designed to produce a cycloidal displacement of the pistons 136, 138 so that the fluid output of the pump 100 is substantially uniform and continuous over the entire pump displacement range.

The cam members 150 and 160 are held in a relatively stationary position during the operation of the pump 100 by means of a pair of rotatable ring gears 166 and 168. The ring gears 166 and 168 are restrained from lateral movement by the housing 112 and the end plate 114. The inner surface of each gear includes splines or keys 169 which mate with corresponding splines or key slots provided on the adjacent outer circumference of the cam members 150 and 160. The splines or keys 169 allow the cam members 150 and 160 to move axially within the housing 112 so that the piston balls 136, 138 will be self-seating in the cam tracks 152, 162. The splines 169 further cause the cam members 150, 160 to rotate in unison with the connected ring gears 166, 168.

As described above with respect to the pump 10, the fluid displacement of the pump 100 is gradiently varied by adjusting the circumferential positioning of the cam members 150 and 160. Beveled gear teeth 166a and 168a are provided on the inner portion of the ring gears 160 and 168 to accomplish this adjustment. The teeth 166a and 168a mesh with a bevel pinion (not shown) which is connected to a control shaft 170. A control handle 172 is coupled to the shaft 170 to permit manual adjustment of the cam members 150 and 160. As set forth above, with respect to the pump 10, rotation of the control handle 172 rotates the cam members 150 and 160 an equal distance in opposite directions to change the axial relationship between the cam lobes l52a-d and 162a-d, and thereby change the reciprocation of the piston balls 136 and 138.

As also described above, the cam members 150 and 160 are initially placed in a selected relationship with respect to the porting provided in the insert 140 and end plate 114, to provide the pump 100 with an infinitely variable fluid displacement over a broad range in either a positive or negative direction from a zero displacement condition. The cam members 150 and 160 are initially arranged so that the lobes 152a-d are 45 out of phase with respect to the lobes 162a-d. Such an arrangement initially places the bottom center-point of the lobes 152a-d in axial alignment with the top centerpoint of the lobes 62a-d, and vice versa. In addition,

the cam members 150 and 160 are circumferentially arranged in the housing 112 so that maximum port opening for the pump occurs when one of the piston balls 136 or 138 is engaged with the top centerpoint of one of the cam lobes 152a-d or 162a-d, and the other axially aligned piston ball 138 or 136 is engaged with the bottom center-point of the other cam lobe 162a-d or 152a-d. Such an arrangement defines the zero displacement condition for the pump 100. At the zero point in the operation of the pump 100, the fluid conduits 147a and 148a are axially aligned to the maximum open port condition, with one of the piston chambers 131 in the rotating piston barrel 130. Since the axially aligned piston balls, such as 136a and 138a, are engaged with either the top or bottom center-point of the associated cam lobes 152 a-d or 162a-d. the aligned balls will be reciprocating in opposite directions. The resulting pumping displacement of the fluid contained within the piston chamber 131 therefore will be zero.

The control handle 172 is rotated to increase the pump displacement from the zero displacement condition to a maximum displacement condition in either di- 'rection of fluid flow. In FIG. 4 the pump 100 is illustrated in the maximum positive-displacement condition. The cams 150 and 160 are positioned with the lobes 152a-d in axial alignment with the lobes 162a-d, and the lobes 152a and 162a are positioned 22 /2 to the right of the vertical centerline of the pump 100, as viewed in FIG. 4. Clockwise rotation of the piston barrel will thereby produce maximum displacement of fluid by the pump 10 in a positive direction with the fluid entering the suction bore 111 and discharging from the outlet bore 115.

The operation of the pump 100 under the maximum positive displacement condition is illustrated in FIG. 4. Since lobes 152a-d and 162a-d on the cam members and are in axial alignment, the aligned pistons 136a, 138a, etc., will reciprocate in unison. Thus, with the piston barrel 130 rotating in a clockwise direction in FIG. 4, the aligned pistons 136i and 138i are moving radially outward on the cam members 150 and 160 in between the cam lobes 152d, 162d and 152a, 162a. Simultaneously, the piston barrel 130 is positioned to align the piston chamber 131 associated with the piston balls 136i, 1381' with the axial inlet conduits 144d, 145d. The outward motion of the piston balls 136i, 1381' thereby creates a suction force which draws fluid from the inlet collector chamber 113 into the associated piston chamber 131. This suction stroke continues as the barrel 130 rotates and continually reduces the port opening between the chamber 131 and the conduits 144d, 145 d.

The rotation of the barrel 130 next carries the pistons 136i, 1381' into the position illustrated by the aligned pistons 136a, 138a in FIG. 4. In this condition, the piston chamber 131 is axially aligned with the outlet conduits 147a, 148a, and the piston balls 136i, 1381' (as shown by the balls 136a, 138a in FIG. 4) are approaching the cam lobes 152a, 162a. The following rotation of the barrel 130 thereby engages the balls 136i, l38i with the lobes 152a, 162a, and drives the balls radially inward through a pressure or pumping stroke. The fluid in the chamber 131 is hence forced through the conduits 147a, 148a into the outlet collector chamber 1 16. The pressure stroke for the piston balls 136i, 138i is completed, and a successive suction stroke is initiated,

when the balls 136i, 138i assume the position illustrated by the balls 136b, l38b in FIG. 4.

As illustrated in FIG. 4, the four lobes l52a-a' and l62a-d will cause each set of aligned piston balls 136i, 138i, 136a, 138a, etc., to reciprocate through four complete suction and pressure strokes for each revolution of the piston barrel 130. Moreover, the arrangement of the components of the pump 100 in accordance with this invention causes the sets of aligned piston balls to be placed in different stages of the pressure or suction stroke at any given instant during the operation of the pump 100. The operation of the piston balls is thus staggered to produce a substantially uniform displacement of fluid through the pump 100.

TRANSMISSION FIG. illustrates a first embodiment of a hydrostatic transmission 200 incorporating the features and advantages of the present invention. The transmission 200 is a constant horsepower unit comprised of two units, similar to the above-described pump 100, coupled to form a variable speed transmission.

In FIG. 5 the right-hand portion of the transmission 200 embodies a variable displacement pump 100A which is substantially identical, in construction and operation, with the pump 100 described above. The lefthand portion of the transmission 200 embodies a variable speed hydraulic motor 100B. The motor 100B is also substantially identical in construction to the above-described pump 100, but operates in a reversed fashion.

Generally, the pump 100A in the transmission 200 is driven by the rotation of an input shaft 220. The pump 100A then operates to displace fluid under high pressure into the motor 100B. The motor 1008 in turn functions to drive an output shaft 230. The fluid displacement of the pump 100A can be gradiently varied in a positive or negative direction, within a selected range, by manipulating the adjustment handle 222. Similarly, the fluid displacement of the motor 1008 can be gradiently varied by manipulating the adjustment handle 232.

The arrangement of the operation components of the pump 100A and motor 100B are essentially identical to the components of the pump 100 shown in FIGS. 3-4, and have not been duplicated in FIG. 5. The fluid interface connections between the pump 100A and the motor 1008 are schematically illustrated in FIG. 5. Generally, the pump 100A includes a housing 112A and a porting insert 140A. A plurality of uniformly spaced and axial fluid outlet channels 149A are provided in the housing 112A to connect the axially aligned outlet conduits 147A, provided in the porting insert 140A, and 148A (not shown in FIG. 5; see FIG. 3). The channels 149A and conduits 147A and 148A (not shown) are in fluid communication with a common annular outlet collector chamber 116A. The chamber 1 16A also functions as the inlet fluid collector chamber for the motor 100B. As explained in detail with respect to FIGS. 3-4, the fluid displacement of the pump 100A is created by discharging fluid from the piston chambers 131A provided in a rotating piston barrel 130A. The displaced fluid is collected in the common chamber 116A for discharging into the motor 1008.

The pump 100A also includes uniformly spaced inlet channels and conduits such as illustrated by the conduits 144a-d and l45ad, and the connecting channels 146 shown in FIGS. 3 and 4. The inlet conduits and channels, not shown in FIG. 5 are alternately spaced with respect to the above-described outlet channels and conduits, and are in fluid communication with an annular common inlet collector chamber 113B. The chamber 1138 also functions as the outlet collector chamber for the motor 1008. In operation, fluid therefore is recirculated into the piston chambers 131A of the pump A from the collector chamber 113B.

The pump housing 112A also defines the outer housing for the motor 1008. A plurality of uniformly fluid inlet channels 1468 are provided in the motor 1008. in axial alignment and fluid communication with the outlet channels 149A of the pump 100A. The motor also includes a porting insert B in axial alignment with the pump insert 140A. Motor inlet conduits 1458 are uniformly arranged in the motor insert 140B in axial alignment and fluid communication with the outlet conduits 147A of the pump 100A. As illustrated in FIGS. 3 and 4, the motor 100B also includes inlet conduits 1448 (not shown) axially aligned with the conduits B and coupled to the conduits 145B by the channels 1468. The fluid is displaced into the motor 100B through the channels 1463 and conduits 144B, 145B from the annular collector chamber 116A. The fluid then flows into the piston chambers 1318 provided in a rotatable piston barrel 1308. The motor 100B further includes uniformly spaced outlet conduits and channels such as illustrated by the conduits l47a-d and 148ad, and the connecting channels 149 shown in FIGS. 3 and 4. These outlet passages for the motor 1008, not shown in FIG. 5, are in fluid communication with the collector chamber 116A, which functions both as an outlet collector chamber for the motor 100B and an inlet collector chamber for the pump 100A.

The transmission 200 also includes a compression spring 143A which functions to apply an axial preload force to the aligned porting inserts 140A and 1408. The spring 143A thereby urges the inserts 140A and 1408 toward the associated piston barrels 130A and 130B, respectively, to assist in maintaining a fluid seal at the interface between the rotating piston barrels and stationary porting inserts. A common pressure piston A is placed in fluid alignment with the conduits 147A and 145B, and operates to balance the axial pressure forces applied against the inserts 140A and 1408, in the same manner as described above with respect to the pressure piston 180 shown in FIG. 3. In the transmission 200 the piston 180A includes a central bore so that fluid will flow through the piston from the pump conduits 147A into the motor conduits 1458 during the operation of the transmission.

In other respects the construction of the pump 100A and motor 100B is the same as the pump 100 illustrated in FIGS. 3 and 4. Moreover, the operation of the pumps 100 and 100A are the same. The motor 100B operates in the reverse manner, as compared to pumps 100 and 100A, and responds to the introduction of fluid under pressure from the pump 100A to drive the output shaft 230. Both the motor 1008 and the pump 100A are gradiently variable within a selected range, in either a positive or negative direction.

During the operation of the hydraulic transmission 200 the output speed and direction of rotation of the output shaft 230 can be varied by manipulation of the adjustment handles 222 and 232. If the pump 100 A is adjusted to its zero displacement setting, no fluid will be displaced by the pump 100A to the motor 100B, and the speed of the output shaft 230 will be zero, regardless of the adjustment of the motor 1008. As the pump 100A is adjusted to increase its fluid displacement, the output speed of the motor 100B and shaft 230 increases proportionately, up to a maximum for each setting of the motor 1008. When the levers 222 and 232 are adjusted so that the pump 100A and motor 1008 have the same fluid capacity or displacement, then the output speed of the shaft 230 will equal the input speed of the pump shaft 220.

When the fluid displacement of the motor 1008 is less than the selected displacement for the pump 100A, the transmission 200 operates as a speed increaser. If the displacement of the motor 1008 is one'half of the selected displacement for the pump 100A, the motor speed and the speed of the output shaft 230 will be twice the speed of the pump and the input shaft 220. Similarly, if the displacement of the motor 1008 is set to be larger than the displacement of the pump 100A, the transmission 200 will operate as a speed reducer in the same manner. If the motor displacement is twice the pump displacement, the speed of the output shaft 230 will be one-half the speed of the input shaft 220.

Thus, the motor 1008 and the pump 100A can be gradiently adjusted to provide speed reducing or speed increasing ratios for the transmission 200 which can be varied within a broad range. Moreover, the direction of rotation of the output shaft 230 can be selected by adjusting the setting of the pump 100A into either a positive or negative flow condition. Also, if the speed of the motor 100B is doubled for any particular pump setting, the torque of the motor shaft 230 will be reduced by a factor of one-half, etc. The transmission 200 is therefore very versatile, and can be adjusted to meet the requirements of many different transmission applications.

FIG. 6 illustrates a second embodiment of a hydrostatic transmission 300 incorporating the features and advantages of the present invention. In contrast to the above-described transmission 200, the illustrated transmission 300 is designed to produce substantially constant torque output at constant fluid pressure. As seen in FIG. 6, the transmission 300 comprises a pump 100C and a motor 400 coupled to form a variable speed hydrostatic transmission.

The pump 100C in the transmission 300 is identical in construction and operation with the above-described pump 100A incorporated in the transmission 200. The pump 100C is driven by the rotation of an input shaft 320, and operates to displace fluid under high pressure into the motor 400. The motor 400 then drives the output shaft 330. The fluid displacement of the pump 100C can be gradiently varied, in a positive or negative direction within a selected range, by manipulating the adjustment handle 322. In this constant torque transmission 300, the displacement of the coupled motor 400 is fixed at a selected multiple of the displacement of the pump 100C. In the illustrated embodiment, the motor displacement is about three times as large as the pump displacement.

The operative components of the pump 100C are described in detail above with respect to the pumps 100 and 100A. Generally, the pump 100C includes a housing 112C and a porting insert 140C. Four uniformly spaced axial fluid outlet channels 149C and conduits 147C (and conduits 148A; see FIG. 3) are in communication with a common annular output collector chamber 116C. The fluid displacement of the pump C is created by discharging fluid under pressure from the piston chambers 131C provided in the rotating piston barrel C. This fluid under high pressure is collected in the chamber 116C for discharge into the coupled motor 400. The chamber 116C also operates as the inlet fluid collector chamber for the motor 400.

The pump 100C also includes four uniformly spaced inlet channels and conduits such as illustrated by the conduits 144a-d, l45a-d and the connecting channels 146, shown in FIGS. 3 and 4. These inlet conduits and channels are alternately spaced with respect to the above-described outlet conduits 147C and 148A and channels 149C, so that the rotation of the piston barrel 130C aligns the piston chambers 131C sequentially with the inlet and outlet conduits during the operation of the pump 100C. The inlet channels and conduits (not shown in FIG. 6) are in fluid communication with an annular inlet collector chamber 413. The transmission 300 is designed to recycle the fluid from the motor to the pump. Hence, the chamber 413 also operates as a fluid outlet collector chamber for the motor 400.

The motor 400 is contained within an enlarged housing 412. Four fluid outlet channels 449 and four inlet channels 446 are alternately spaced around the housing 412. The motor inlet channels 446 are coupled to the motor inlet collector chamber 116C and the outlet channels 449 are coupled to the motor outlet collector chamber 413. The outer end of the motor housing 412 is closed by an end plate 414. The plate 414 supports the output shaft 330 in a suitable bearing 419. Retaining rings 420 engage within grooves in the shaft 330 and with the bearing 419 to prevent axial movement of the shaft 330 during operation of the transmission 300. The end plate 414 further includes a plurality of radial fluid channels 449A and 446A in direct communication with the axial channels 449 and 446, respectively. Four additional axial channels 4498 and 4468 are coupled to the radial channels 449A and 446A to connect the channels 449, 446 and 449A, 446A with the interior of the motor 400. Suitable O-ring seals 401 are positioned between the end plate 414 and housing 412 to maintain the fluid seal for the channels 449 and 446.

The motor 400 also includes an annular porting insert 440 in axial alignment with the pump insert C. Four inlet conduits 444 are uniformly spaced in the insert 440 in axial alignment and fluid communication with the motor inlet channels 4468 and the pump outlet conduits 147C and 148A. The fluid displaced from the chamber interior through the pump conduits 147C is directed to the interor of the motor 400 through these inlet conduits 444 in the insert 440. The porting insert 440 in the motor 400 also includes four axial outlet conduits 147 uniformly spaced between the axial conduits 444 (conduits 147 are not shown in FIG. 6, but are evident from FIG. 3). The outlet conduits 147 are in axial alignment and fluid communication with the motor outlet channels 4493 and are connected to the motor outlet collector chamber 413 by radial passages 449C, as shown in phantom lines in FIG. 6.

The motor 400 further includes a rotatable piston barrel 430. The barrel 430 is joined to the output shaft 330 by suitable splines or keyways 421 so that rotation of the barrel imparts rotation to the shaft 330. The piston barrel 430 in the transmission 400 is an enlarged version of one side of the piston barrel 130 incorporated in the pump 100, as shown in FIGS. 3 and 4. The barrel 430 in the illustrated transmission includes nine radial cylinders 434. A reciprocating piston ball 438 is located in each of the cylinders 434. Each cylinder 434 is connected by a radial port 437 to an axial bore 431. The bores 431 are arranged in the barrel 430 to be in axial alignment with the fluid conduits 144, 447, 446B and 4498 as the barrel 430 rotates with respect to the stationary insert 440 and the end plate 414.

The motor 400 is provided with an annular cam member 450 which is arranged in rolling engagement with the piston balls 438. The cam member 450 surrounds the piston barrel 430 and defines an arcuate cam track 452 which has a radius of curvature slightly larger than the radius of the engaged piston balls 438. As illustrated in FIG. 4 with respect to the cam tracks 152 and 162, the cam track 452 includes a plurality of uniformly spaced cam lobes which engage with the piston balls 438. In the preferred embodiment, the cam track 452 is designed to include four equally spaced lobes which are uniform in amplitude and wave length, and which further are machined to produce a cycloidal displacement of the piston balls 438. The cam member 450 is held from rotation by suitable means such as a pin 451 connected to the housing 412. However, the pin 451 permits the cam member 450 to move axially so that the balls 438 and track 452 will be selfcentering.

The motor 400 operates in the opposite manner with respect to the pump 100 described above. When the pump 100C is operated in a positive direction, fluid under high pressure from the pump 100C is directed into four of the piston chambers 431 through the inlet conduits 444 and 4468. The admitted high pressure fluid enters four of the cylinders 431 and urges the piston balls 438 radially outward against the cam track 452. The lobed configuration of the cam track 452 converts the radial reciprocating motion of the piston balls 438 into rotatory motion of the piston barrel 430. The motion of the barrel 430 in turn drives the connected output shaft 330. As the barrel 430 rotates, the four cylinders 431 charged with fluid align with the motor outlet conduits 147 and 4498. The lobes on the cam tracks 452 then exhaust the fluid from the piston chambers 431 into the outlet conduits 449B and 147, and into the collector chamber 413.

If the pump 100C is operated in a negative direction, the operation of the motor 400 is reversed. The fluid from the pump 100C will flow into the motor outlet conduits 147 and 449B and will enter the alternate four piston chambers 431. Then the fluid will exhaust from the motor 400 through the inlet conduits 4468 and 444. In this negative mode of operation, the functions of the fluid collector chambers 116C and 413 also are reversed.

As seen in FIG. 6, the transmission 300 also includes a compression spring 143C positioned at the interface between the pump 100C and the motor 400. The spring 143C functions to apply an axial preload force to the alignment porting inserts 140C and 440, to urge the inserts toward the associated piston barrels 130C and 430 and thereby maintain a fluid seal between the stationary inserts and the rotating piston barrels. A common pressure piston 180C is placed in fluid alignment with each of the aligned conduits 147C, 444 and 148A, 147 and operates to balance the axial pressure forces applied against the inserts 140C and 440, in the same manner as described above with respect to the pressure pistons 180 as shown in FIG. 3. The pistons 180C includes a central bore so that fluid will flow through the pistons from the pump conduits 147C and 148C into the motor conduits 444 and 147 during the operation of the transmission 300.

The hydrostatic transmission 300 is a speed-reducing and torque-increasing unit which produces a substantially constant torque output at constant fluid-pressure. The speed reduction of the transmission 300 results from the motor 400 having a fluid displacement larger than the displacement of the pump C. In the illustrated embodiment, for instance, the motor 400 has about three times the displacement of the pump 100C. Therefore, at maximum pump displacement. the motor speed will be about one-third the speed of the pump 100C. The torque-increasing characteristics of the transmission 300 derive from the fact that theoretically (assuming no fluid friction losses), the input horsepower of the pump 100C equals the output horsepower of the motor 400 at any given fluid displacement. Hence, since horsepower is a function of torque multiplied by speed, the built-in lower speed for the motor 400, compared to the speed of the pump 100C, increases the torque output of the motor 400.

The illustrated pumps, motors and transmissions in accordance with this invention possess many inherent advantages. For instance, the performance of the fluid transducer (pump or motor) is improved by reciprocating the piston balls in a radial rather than an axial direction. With radially reciprocating pistons, the centrifugal force on the piston balls in pumps assists the necessary piston return movement. In fact, since the centrifugal force on the piston balls increases as a square of the speed of rotation of the piston barrel, the charging pressure necessary to feed fluid into the unit, when operating as a pump, is reduced proportionately.

Moreover, the use of pairs of reciprocating piston balls coupled by a common piston chamber in the piston barrel permits the fluid displacement of the fluid transducer, in accordance with this invention, to be variable. With only one reciprocating piston ball arrangement, as seen from the motor 400 illustrated in FIG. 6, the fluid displacement of the transducer is fixed. The design of the preferred embodiment also allows the fluid displacement of the transducer to be infinitely varied, within a selected range, in either a positive or negative direction from a zero displacement condition. This feature of the invention results from the use of uniformly spaced and symmetrical lobes on the adjustable cam tracks which engage with the piston balls.

The preferred cam profile which induces cycloidal movement of the piston balls also induces uniform fluid flow through the transducer. The cycloidal piston movement further reduces fluid flow velocity, and fluid noise, errosion and cavitation in the transducer, by controlling the piston balls so that the piston acceleration is approximately zero when the associated piston ports open. The preferred arrangement of 9 pairs of piston balls and 8 associated fluid ports (four inlet and four outlet ports) in combination with the cycloidal piston movement also improves the fluid flow characteristics of the transducer by causing the port openings to occur at a rate which increases more rapidly than the increase in the speed of the piston balls.

The arrangement of the piston chambers and fluid passages in the pump or motor in accordance with this invention also produces uniform flow conditions, low fluid flow velocity, minimum fluid friction losses. The invention provides a common fluid collector chamber for both the inlet and outlet sides of the transducer (e.g. chambers 113 and 116 in FIG. 3) to minimize fluid surges and pressure fluctuations. The fluid passages in the transducer are designed to facilitate uniform flow of fluid to and from the common collector chambers. In the embodiment shown in FIGS. 1 and 2, four pairs of piston balls 36 and 38 are simultaneously coupled to the common inlet collector chamber 42 by means of the passages 44, 46 and 48. Since the suction strokes of the piston balls are sequentially staggered, this arrangement connects the inlet collector chamber 42 simultaneously to four pairs of pistons which are in different stages of operation (i.e., one piston pair is at maximum suction; one is beginning a suction stroke, and two are at intermediate stages). In the same regard, in the pump shown in FIGS. 1 and 2, four pairs of piston balls 36, 38 are simultaneously coupled to the common outlet collector chamber 41 by means of passages 43, 45 and 47. Thsu, the fluid displaced into the collector chamber 41 by each connected pair of piston balls 36, 38 is likewise sequentially staggered, and the fluid flow into the collector chamber 41 will thereby be uniform and continuous.

The modified pump 100 illustrated in FIGS. 3 and 4 also includes a flow passage arrangement which creates uniform flow with minimum fluid friction losses. The pump 100 includes large axial inlet conduits 144, 145 and outlet conduits 147, 148 which align with the common piston chambers 131 sequentially as the piston barrel 130 rotates. The inlet conduits 144, 145 lead from the common inlet collector chamber 113 to the piston chambers 131, and the outlet conduits 147, 148 lead from the piston chambers 131 to the common outlet collector chamber 116. Due to this axial porting arrangement at both sides of the barrel 130, the porting area in the pump 100 can be maximized to reduce the flow velocity and fluid friction losses in the pump. The design of the pump 100 allows the conduits area to be expanded to any desired size downstream from the piston barrel 130, to minimize the flow velocity and fluid friction losses.

The embodiment of the invention illustrated in FIGS. 1 and 2 also readily allows adequate lubrication of the highly machined pintle 40 and piston barrel 30. The radial stroke of each piston ball 36, 38 creates a small, unbalanced radial load on the pintle 40, the barrel 30 and the shaft 20. Since each of the nine pistons 36, 38 in the piston pairs will experience four strokes per revolution of the barrel 30, due to the four lobes on the cam tracks 52, 62, the radial direction of the resultant small, unbalanced force will change seventy-two times for each revolution of the piston barrel 30. As a result, the stationary pintle 40 and the rotary barrel 30 will be subjected to a slight oscillating motion which creates a squeeze lubrication effect between the engaged surfaces of the barrel and pintle.

Although the invention has been described above with a certain degree of particularity with respect to several embodiments, it should be understood that this disclosure has been made only by way of example. Consequently, numerous changes in the details of construction and in the combination and arrangement of the components as well as the possible modes of utilization for the rotary engine power assembly in accordance with this invention will be apparent to those familiar with the art, and may be resorted to without departing from the scope of the invention.

What is claimed is:

1. A variable displacement fluid transducer comprising:

rotatable shaft means;

a cylindrical piston barrel joined for rotation with said shaft means;

first and second sets of radial pistons, with each set comprising nine uniformly sized piston balls positioned for reciprocation within cylinder bores spaced uniformly around said barrel, said sets of pistons arranged on said barrel to define nine pairs of piston balls;

a piston chamber defined in said barrel for each pair of piston balls so that radial reciprocation of said pair of balls displaces fluid from the associated piston chamber;

first and second cam tracks circumscribing said first and second sets of pistons, respectively; each of said tracks including four uniformly positioned cam lobes projecting inwardlytoward said barrel in frictional engagement with the balls of one set of pistons and shaped to reciprocate the engaged piston balls with a cycloidal motion so that the relative rotation of said barrel with respect to said tracks orbits said balls into engagement with the cam lobes and reciprocates said balls radially to operate said transducer with a fluid displacement from said piston chambers having a substantially uniform flow through said transducer proportionate to the relative reciprocating movement of the balls comprising each pair;

control means to move said cam tracks to shift the circumferential alignment of said lobes on said first track with respect to the lobes on said second track to vary the fluid displacement of the transducer by changing the relative reciprocating movement of the balls comprising each pair; and

fluid porting means for directing fluid sequentially to and from each of said piston chambers during the operation of said transducer.

2. A transducer in accordance with claim 1 wherein said cam tracks are arcuate in cross'section with a radius of curvature larger than the radius of the engaged piston balls and further wherein the center of curvature of said cam tracks is spaced axially inwardly from the center of the engaged piston balls so that the reaction force of said balls on said tracks urges said tracks together and thereby assists in maintaining the tracks in an axially stationary position.

3. A variable displacement fluid transducer comprising:

rotatable shaft means;

first and second sets of radial pistons, with each set comprising nine uniformly sized piston balls positioned for reciprocation within cylinder bores spaced uniformly around said barrel;

said sets of pistons being arranged on said barrel to define pairs of piston balls;

a piston chamber extending through said barrel for each pair of piston balls so that radial reciprocation of said pairs of balls displaces fluid from the associated piston chamber;

first and second axially aligned cam tracks circumscribing said barrel in radial alignment with said first and second sets of pistons, respectively; each of said tracks including four uniformly shaped and and outlet ports alternately spaced in a uniform circular pattern, with said inlet and outlet ports spaced adjacent said barrel so that rotation of said barrel places each piston chamber in sequential positioned cam lobes projecting inwardly toward communication with said inlet and outlet ports, said barrel in frictional engagement with the balls said ports and piston balls being further arranged of the aligned set of pistons and shaped to reciproso that said cycloidal motion of said piston balls cate the engaged piston balls with a cycloidal mocauses said inlet and outlet ports to open to said tion so that the relative rotation of said barrel with piston chambers at a rate which increases faster respect to said tracks orbits said balls into engagethan the increase in speed of reciprocation of said ment with the aligned cam lobes and reciprocates piston balls; and

said piston balls radially to thereby operate said fluid collector means in fluid communication with transducer with a fluid replacement providing a said inlet and outlet ports for directing fluid to said substantially uniform flow through said transducer inlet ports and receiving fluid displaced from said proportionate to the relative reciprocating moveoutlet ports.

ment of the balls comprising each pair;

4. A fluid transducer in accordance with claim 3 wherein the cross-sectional area of said piston chambers and said inlet and outlet ports is substantially the same and the inlet and outlet ports are spaced alternately so that each rotating piston chamber closes an inlet port and opens the successive outlet port substantially simultaneously.

5. A fluid transducer in accordance with claim 4 wherein said porting members include four inlet and four outlet ports.

each side of said barrel and including axial inlet

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Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US4128048 *Dec 16, 1976Dec 5, 1978Teijin Seiki Company LimitedRadial piston type hydraulic pump-motor
US4132154 *Dec 16, 1976Jan 2, 1979Teijin Seiki Company LimitedRadial piston type hydraulic pump-motor
US4700613 *Aug 22, 1986Oct 20, 1987Oy Partek AbHydraulic motor with stationary axle and rotating fluid distributor
US5080050 *Jan 29, 1990Jan 14, 1992Irving M. SmithRotary engine
US5177967 *Jul 22, 1991Jan 12, 1993Tecumseh Products CompanyHydrostatic transaxle
US7614337 *Jul 9, 2007Nov 10, 2009Gabriele PecorariRotary radial piston machine
WO1991011595A1 *Jan 3, 1991Aug 8, 1991Thomas W DaleRotary engine
Classifications
U.S. Classification91/492
International ClassificationF03C1/247, F03C1/40, F03C1/28, F03C1/30, F03C1/26, F01B13/06, F04B1/107, F04B1/113, F16H39/32, F03C1/24, F16H39/18, F03C1/32, F04B1/10, F04B49/00
Cooperative ClassificationF01B2009/045, F16H39/32, F01B13/06, F04B49/005
European ClassificationF04B49/00F, F01B13/06, F16H39/32