US3095708A

US3095708A - Variable displacement hydraulic assembly

Info

Publication number: US3095708A
Application number: US141555A
Authority: US
Inventors: Lee T Harris
Original assignee: Individual
Current assignee: Individual
Priority date: 1957-07-09
Filing date: 1961-08-24
Publication date: 1963-07-02
Anticipated expiration: 1980-07-02

Description

July 2, 1963 1-. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Original Filed July 9, 1957 '7 Sheets-Sheet l S mm m 2 m m A M w I V m m X m H \X\ v T a "w E Q $3 w 5 a N a VW mm B mm b mN .TN m i I M/ ll H n ll R 1 1 lh1l|al m S 3 H I. MN Q: I t .8 3 Ev mwm mm m w 2 mm M NN OH W. v? ni|| H 8 8 July 2, 1963 L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY 7 Sheets-Sheet 2 Original Filed July 9, 1957 S R. m O m R w A 0 mm mmfiw A H .FLTW Q Q T E W mm om L 2 E Q Y \N. Mm B a m mm Fm 5 IN ,I 1 IHMI d I m fl nnh m N m R: T l w l n u l I E. 2 3 mv mv NM d N WM mm NQ mm Om w a NM mm 2 m" E mm g .2 ON 3 \F I 0 o 0 0 o \L m m July 2, 1963 1.. T. HARRIS 3,

VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Original Filed July 9, 1957 7 Sheets-Sheet 5 ifiijmg g r INVENTOR.

LEE T HARRIS July 2, 1963 L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY 7 Sheets-Sheet 4 INVENTOR. LEE T HARRI 8 Original Filed July 9, 1957 July 2, 1963 Y -r. HARRIS 3,095,708

VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Original Filed July 9, 1957 7 Sheets-Sheet 5 INVENTOR.

LEE T. HARRIS July 2, 1963 L. T. HARRIS 3,095,703

VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Original Filed July 9, 1957 XXII F19. 21 15 7 Sheets-Sheet 6 I I l INVENTOR.

a LEE T HARRIS July 2, 1963 Original Filed July 9, 1957 TO IGNITION 5W.

T0 BAND BRAKE SOLENOID SERVO-VALVE Fig.52

L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY 7 Sheets-Sheet 7 TO ENGINE VACUUM TO BAND BRAKE SOLENOID SERVO-VALVE M LL w 120 BRUSH BRUSH ARM COMMUTATON ROTOR .TO TO RELAY R RELAY F INVENTOR.

BY LEET HARRIS United States Patent Q 3,095,708 VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Lee T. Harris, 511 William St., Rome, N.Y. Original application July 9, 1.957, Ser. No. 670,814, now Patent No. 3,044,409, dated July 17, 1962. Divided and this application Aug. 24, 1961, Ser. No. 141,555

9 Claims. (Cl. 60-53) This application is a division of my Patent 3,044,409 dated July 17, 1962.

(This invention relates to hydraulic pump mechanisms and more particularly to a variable displacement pump mechanism that may be used in a driving and driven configuration to comprise a unique hydraulic transmission system.

For some years now variable displacement pumps have been used in hydraulic transmission systems for coupling a prime mover to a load such as in automobiles and other industrial applications. These transmissions and pumps have generally been satisfactory for various applications but have had certain disadvantages such as limited inputoutput speed ratios requiring special gear trains, high frictional and hydraulic losses resulting in low efiiciency transmissions, costly and expensive machining of highly stressed and critical parts resulting in high cost limited installations. According to the present invention I have discovered a variable displacement hydraulic pump and transmission system that permits infinite variation in input-output speed and torque ratios that is dynamically and statically balanced so as to minimize frictional forces and that still may be manufactured largely by castings in a cheap and simple manner.

Accordingly it is an object of the present invention to provide a variable displacement pump mechanism that is infinitely variable in displacement. It is another object of the present invention to provide a transmission system that is infinitely variable in input-output speed ratio throughout the design limits thereof. It is another object of the present invention to provide a variable pump mechanism that is dynamically and statically balanced so as to minimize frictional forces therein. It is another object of the present invention to provide a transmission system that is extremely flexible in application and efficient in operation. It is another object of the present invention to provide a transmission system wherein the driving pump may be located remotely from one or more driven pump mechanisms. It is another object of the present invention to provide a variable displacement hydraulic pump mechanism that may be combined with a corresponding mechanism to provide a transmission that rotates together at a one-to-one speed ratio without external gearing or locking. It is another object of the present invention to provide a transmission system that requires no gear train or other speed control mechanism to connect it between a prime power source and a load. It is a still further object of the present invention to provide a variable displacement pump and transmission mechanism that is highly efiicient, extremely simple to operate and economical to construct. These and other and further objects will be in part apparent and in part pointed out as the specification proceeds.

In the drawings:

FIGURE 1 is an axial section of the transmission with the rotatable inner portion shown at a position wherein the pitch axis of the driving and driven pump divider disks are perpendicular to the plane of the drawing and wherein Ice FIGURE 3 is a side elevation of the bearing member in the same aspect relationship as shown in FIGURE 1;

FIGURE 4 is a top plan partially broken away of the valve and bearing member of FIGURE 1;

FIGURE 5 is a vertical section of. the valve and hearing member along the line VV of FIGURE 3;

FIGURE 6 is an axial section of the ducting core which fits into the bearing member in the same aspect relationship as shown in FIGURE 4 with certain parts shown in full lines for clarity;

FIGURE 7 is a view similar to FIGURE 6 taken on line VII-VII of FIGURE 6.

FIGURES is a partial sectional view on line VIIIVIII of FIGURE 7;

FIGURE 9 is an end elevational view of the disk support assembly as viewed from the left hand side of FIGURE 2 and FIGURE 10;

FIGURE 10 is a side elevational view of the disk support assembly in the same aspect relationship as shown in FIGURE 2;

FIGURE 11 is a top plan View of the disk support assembly of FIGURE 10;

FIGURES 12 and 12A are from left to right an end and side view of either of the two bushings that fit onto the transverse projections of the disk support assembly;

FIGURE 13 is a diagrammatic drawing of the transmission main oil ducting described by 360 counterclockwlse rotation as viewed from the left hand end of FIG- URE 1 of the ducting core, starting at top center in respect to FIGURE 1;

FIGURE 14 is a side View of the pump chamber end piece in the same aspect relationship as shown in FIGURE 1;

FIGURE 15 is a top plan view of the chamber end piece of FIGURE 14;

FIGURE 16 is a vertical transverse section of the pump chamber end piece along the line XVIXVI of FIG- URE 14;

FIGURE 17 is a right half end view of the frontal end plate of the inner housing as viewed from the right hand side of FIGURE 1;

the transmission is at an input-output forward speed ratio ofone-to-one;

FIGURE 2 is an axial section on line IIII of FIG-- FIGURE 18 is an axial section of the frontal end plate of the inner housing in the same aspect relationship as shown in FIGURE 1;

FIGURE 19 is an axial section of the frontal end plate of the inner housing in the same aspect relationship as shown in FIGURE 2;

FIGURE 20 is an end View of the intermeshed rotor and divider disk of the driving pump as viewed from the right in FIGURES 1 and 2;

FIGURE 21 is a side view of the driving pump rotor and divider disk of FIGURE 20 shown with the latter in section along line XXIXXI of FIGURE 20;

FIGURE 22tis a partial transverse section of the drivmg pump ro or and divider dis a XXII-XXII of FIGURE 21; k 310mg the hue FIGURE 23 is an enlarged view of a rotor vane and a partial circumferential section of the intermeshed divider disk showing the relationship of the two to each other at a phase of the rotation cycle where the maximum angle of incidence occurs in respect to the normal;

FIGURE 24 is a partial radial section of the dividerdisk along the line O-XXIV of FIGURE 22 with the rotor removed;

FIGURE 25 isa partial radial section of the divider disk along the line O-XXV of FIGURE 22 with the rotor removed;

FIGURE 26 is an enlarged exploded perspective view of adivider disk insert and the corresponding urging spring;

FIGURE 27 is a schematic representation of a method for forcibly maintaining a relatively uniform speed relationship of the divider disk in respect to the rotor;

FIGURES 28, 29, and 30 are enlarged top, side, and end views respectively of a rotor vane showing a further alternative method of forcibly maintaining a relatively uniform speed relationship of the divider disk in respect to the rotor; and

FIGURES 31 and 32 are schematic drawings of a possible control system for automatically controlling the input-output speed ratio of the transmission for automotive application.

Referring now to FIGURES l and 2, the transmission includes an input shaft 1; an output shaft 2; an outer housing comprising end plate 3 and upper and

lower casings

4 and 5 respectively; an inner housing comprising end plate 6, front and

rear barrel sections

8 and 9, central section 10, and rear section 12; ducting core 13; a first pump chamber bounded by pump chamber end piece 7, bearing member 11, and disk support assembly 16-16; a second pump chamber bounded by member and rear section 12; a driving pump including a rotor 14 and divider disk 15-15; a driven pump or motor including rotor 17, and

divider disk 18-18; over-running clutch 19; and a band brake including band 21 and drum 20.

The driving pump is infinitely variable in displacement from design maximum forward output through zero output to design maximum reverse output as will be subsequently described. The rotor 14 is mechanically coupled to input shaft 1 and has therein six radial vanes 22 (see FIGURE circumferentially spaced 60 apart at the centers with fiat sides and peripheral surfaces which are as surfaces of a common sphere whose center is the geometric center point of the pump chamber, and a core section whose outer surface 23 is spherically contoured with the sphere center coinciding with the same aforementioned geometric center point of the pump chamber, said surface 23 being the innermost surface or boundary of the pump chamber. Front face surface 24 of chamber end piece 7, rear face surface 25 of bearing member 11 and peripheral surface 26 of assembly 16-16 define the other principal boundaries of the first pump chamber and are shaped so as to provide an oil sealing contact to the end and peripheral surfaces of revolution of the vanes 22 of rotor 14. Thus the six vanes 22 circumferentially divide the pump chamber into six separate and equal volumes of fixed displacement.

Divider disk 15-15 comprises two main sections (see FIGURES 20, 21, and 22); six slots 27 for accommodation of the vanes 22 of rotor 14, recesses 28 for accommodating inserts 29, valve chambers 30 for accommodation of shuttle balls 31, oil channels 32, recess grooves 33, oil channels 34, and means for securely fastening the two

main parts

15 and 15 to each other. Said divider disk 15-15 intermeshes with rotor 14 as shown in FIGURES 20, 21, and 22 and serves to longitudinally divide each fixed volume between adjacent vanes 22 into two parts. As visualized in FIGURE 21 divider disk 15-15 would be assembled on rotor 14 by sliding the part 15 onto said rotor 14 from the left and sliding the other part 15 on from the right, one part, either 15 or 15 containing the inserts 29 with corresponding springs and shuttle balls 31, and the two parts 15-15 being joined by screws 36 as shown in FIGURES 20' and 24 or in some other suitable manner. Divider disk 15-15 is supported in axial position by contact or proximity of the outer face bearing surfaces 37 of said divider disk 15-15 with the recessed mating surfaces 37 of disk support assembly 16-16 (FIGURES 2 and 10) and in radial position by rotor 14. Disk support assembly 16-16 contains two projections, 33 and 38, which are journaled in mating bearing surfaces contained in the

barrel sections

8 and 9 of the inner housing (FIGURE 2), said disk support assembly 16-16 thereby being pivotally mounted on a transverse axis defined by an imaginary line connecting the two diametrically opposite bearing centers, said line 4 intersecting the geometric center of the pump chamber. Disk support assembly 16-16 is subject to control in its movements about the axis as will be described herein.

In operation, oil sealing action between divider disk 15-15 and rotor 14 is provided by contact or proximity of the spherically contoured inner contact surfaces 23 and outer slot contact surfaces 27 of said divider disk 15-15 with the mating spherically contoured contact surfaces 23 and 22 of said rotor 14 in any supported position of said divider disk 15-15 within the limits of movement of disk support assembly 16-16 about its transverse axis. Oil sealing contact between the slots 27 and the radial surfaces of vanes 22 is maintained by inserts 29 which are slidably mounted in recesses 28, said inserts 29 being continuously but yieldably urged against vanes 22 by springs 35 or hydraulic pressure channeled into recesses 28 by shuttle balls 31 or a combination of the two forcing means. Shuttle balls 31 actually serve two purposes: (1) by being exposed to the pressure appearing on both sides of divider disk 15-15, between any two adjacent vanes 22, through orifices 39 said shuttle balls 31 are automatically seated against the orifices 39 which are facing the lowest pressure, thereby leaving the orifices 39 which are facing the highest pressure open, and permitting the highest available pressure to be transmitted into valve chambers 30 and then into recesses 28 via oil channels 32 wherein said high pressure acts against the backsides of inserts 29 thereby counteracting the highest pressure appearing on the exposed face surfaces 40 of said inserts 29 (FIGURES 23 and 26); (2) by leaving one orifice 39 open at all times, the oil trapped in recesses 28 by inserts 29 is provided an escape when said inserts 29 are forced back into said recesses 28 from extended positions.

In operation, it will be understood that when the input shaft 1 is rotated, as for example by a prime power source, that rotor 14 likewise rotates by virtue of its direct coupling to input shaft 1 and that since divider disk 15- 15 is mechanically intermeshed with rotor 14, it will likewise rotate with said rotor 14 but will be permitted some circumferential movement relative to rotor 14, said relative movement being restricted to the limits corresponding to the excess clearance between the face contact surfaces 40 of the opposing inserts 29, on each side of a particular slot 27 when said inserts 29 are in their fully retracted positions, as compared to the thickness of vanes 22. Said excess clearance should be at least as great as the maximum phase deviation of corresponding points of rotor 14 and divider disk 15-15 when calculated from the equation tan B=tan A cos P, in which P is the pitch angle of divider disk 15-15, A is any angle of the complete 360 cycle of rotation of rotor 14 with either side of the pitch axis of divider disk 15-15 as the zero degree reference and B is the angle of correspondence of the angle A on said divider disk 15-15 when translated to the plane of rotation of rotor 14. Therefore, angle B subtracted from the angle A represents the phase difference of slots 27 in respect to corresponding vanes 22 for various angles of rotation and for various pitch angles at which divider disk 15-15 may be set between zero and maximum design limits. In order to arrive at the minimum design clearance for vanes 22 in slots 27, angle A should be taken at angles of 45, 225 and 315, and angle P at the maximum design pitch angle of divider disk 15-15'. Maximum angle P is 15 in the illustrated embodiment which causes the maximum phase difference between rotor 14 and divider disk 15-15 to be approximately :1"; therefore the total clearance between face contact surfaces of opposing inserts 29 when in their fully retracted positions in grooves 28 should be at least equal to the thickness of vanes 22 plus r sin 2, where r is the radius from the geometric center of the pump housing.

Although inserts 29 will perform most of the yielding to the phase variation between rotor 14 and divider disk 15-15', the frictional resistance to rotation against divider disk -15 will tend to cause said divider disk 15-15 to lag behindrotor 14 to the limits imposed by the excess clearance between opposed inserts 29 which will tend to cause cyclic variations in the speed of said divider disk 15-15 in respect to the speed of rotation of rot-or 14. FIGURE 27 illustrates a means whereby a more nearly constant speed rotation of divider disk 15-15 in respect to rotor 14 may be assured, when the transmission is operating at an input-output speed ratio of one to one. A groove 41 is provided in the peripheral surface 22 of each vane 22, the sides of which are shaped to provide contact to at least one point of a mating projection 42, located at the peripheral center of each slot 27 in divider disk 15-15, at all phase angles of rotation between 315 to 45 and 135 to 225, with the zero degree reference considered to be in coincidence with one side or the other of the transverse pitch axis of divider disk 15-15.

A further alternative means whereby more nearly constant speed rotation of divider disk 15-15 in respect to rotor 14 may be assured (FIGURES 28, 29, and 30) is to shape the sides of vanes 22 so as to keep one or both of the opposing inserts 29 in any one slot 27 fully retracted into their respective recesses 23 at all phases of cyclic rotation between the angles of 315 to 45 and 135 to 225 with the zero degree reference again chosen to coincide with either side of the transverse pitch axis of divider disk 15-15.

By inspection of the drawings, it will be apparent that divider disk 15-15 operating in a complementary manner with rotor 14 essentially divides the pump chamber into two pumping sections (one of each side of divider disk 15-15) which operate in'a diametrically opposite fashion to each other with each section having its own input-output ports, represented by the

numbers

43 and 43 for the front pump section and

numbers

44 and 44 for the rear pump section as shown in FIGURE 4 and as shown schematically in FIGURE 13. The

ports

43, 43, 44 and 44 should each have a circumferential length of approximately 120, leaving an uncut-away circumferential section of approximately 60 in length between each port which will serve to trap the fluid between ad-' jacent vanes 22 at maximum of minimum points as the case may be and thereby to prevent unrestricted circulation of oil between input and output ports which would tend to bypass the pump. If the rotor employs more or less than six vanes as shown in the drawings, the ports should of course be of appropriate circumferential length corresponding to the number of vanes.

Referring now to FIGURE 1 it is obvious that when the plane or rotation of divider disk 15-15 is completely vertical (zero pitch angle) represented by alignment with the letters N-N, the volume of fluid entrapped between adjacent vanes 22 and the surfaces of the pump chamber remains constant when rotor 14 and divider disk 15-15 rotate and that therefore all of the fluid contained within the chamber rotates with the rotor with no fluid input or output through

ports

43, 43, 44 and 44. It should be equally obvious that if the pitch angle of divider disk 15-15 is at any valve other than zero that the individual fluid volumes between adjacent vanes 22 on each side of divider disk15-15 will vary in a sine manner as motor 14 progresses from 0 through 360 for each rotation cycle, the total changebetween minima and maxima volumes being a direct function of the pitch angle of divider disk 15-15. Therefore a low compressibility fluid such as oil when contained therein will be forcibly circ-ulated into and out of the pump chamber through

ports

43, 43, 44 and 44 in amounts proportional to the pitch angle of divider disk 15-15. In the illustrated embodiment, maximum forward circulation is achieved when the plane of rotation of divider disk 15-15 is aligned with the letters F-F (FIGURE 1); i.e. the pitch angle of divider disk 15-15 isat the forward maximum of 150. Likewise, the maximum 6 reverse circulation will be achieved when the plane of rotation of divider disk 15-15 is aligned with the letters R-R. Specific circulation paths for the fluid will be described in detail herein.

Examination of the underlying principles of operation will disclose that insofar as rotor 14 is concerned, no appreciable unb-alanced forces in respect to the housing occur but that the hydraulic forces on divider disk 15-15 would be severelyunbalanced if not compens-ated for. It is for said purpose of compensation that recess grooves 33 are provided around each peripheral side of divider disk 15-15 (FIGURES 20' and 21) with oil channels 34 providing hydraulic communication between said recess grooves 33 and the opposite side of divider disk 15-15 in the pump chamber. The effective area of recess groove 33 is chosen so that said area multiplied by the radius of the effective center of fluid pressure of said recess groove 33 is equal to the effective area of the part of divider disk 15-15 contained inside the pump chamber between adjacent vanes 22 multiplied by the radius of the efiiective center of fluid pressure of said second area. In this way any hydraulic forces appearing on the inside areas of divider disk 15-15 will be dynamically counteracted by proportionate forces created by transmission of pressure through oil channels 34 to recess grooves 33 located on the opposite sides. The obvious purpose of providing hydrostatic balance for divider disk 15-15 is to reduce the frictional resistance to rotation of said divider-disk 15-15 in disk support assembly 16-16 occurring as a result of hydraulic loading, and thereby to enhance the efliciency of power transfer and decrease mechanical wear. Another method for reducing the friction, in lieu of hydrostatic balance as described above, is to support divider disk 15-15 on ball or tapered roller bearings within disk support assembly 16-16. A third method might be a combination of the above two methods.

The driven pump or motor, hereinafter to be referred to as the driven pump, in the illustrated embodiment is essentially identical to the driving pump except that divider disk 18-18 has a fixed plane (pitch angle of 15) of rotation about its axis by virtue of its being supported in the inner housing and accordingly is a fixed displacement pump or motor. It will be apparent, however, that the driven pump may be of the variable displacement type also and that said driven pump whether of the fixed or variable displacement type may :be similar to or different from the driving pump in basic design. The rotor 17 with six vanes 22A, divider disk 18-18 and associated pump chamber acting in cooperation with each other form the essential elements of the driven pump, the operation of which will be fully understood by reference to the above description of the driving pump. 7

Ducting core 1-3 provides means of fluid communication between the driving pump and driven pump by providing fluid channels '45 and 45 as shown in FIGURES 1 and 2 and as shown schematically in FIGURE 13. Channel 45 connects ports 43' and 44 of the driving pump to ports 436 and 47 of the driven pump and channel 45 connects

ports

43 and 44 of the driving pump to

ports

46 and 47 of the driven pump.

It will be apparent that when input shaft 1 is. rotated clockwise when viewing FIGURES 1 and 2 from the left, and that when divider disk 15-15 is set at a pitch angle greater than zero toward alignment with the letters F-F of FIGURE 1, fluid will be expelled from the chamber of the driving pump through

ports

43 and 44 into channel 45 from which said fluid will be forced into the chamber of the driven pump through ports 46* and 47 thereby causing rotor 17 to rotate in the same clockwise sense as rotor 14, causing said rotor 17 at the same time to expel a like quantity of fluid from the chamber of the driven pump through ports 46' and 47 into channel 45 from which said like quantity of fluid will be returned to the chamber of the driving pump through

ports

43 and 44. Fluid flow under the above conditions is shown schematically by the solid arrows in FIGURE 13. When divider disk 15-15 is set at a pitch angle greater than zero toward alignment with the letters R-R of FIGURE 1 and rotation of input shaft '1 is in the same clockwise sense, the flow of fluid is reversed and would be as shown schematically by the dashed arrows in FIGURE 13, in which case output shaft 2 would rotate counter-clockwise or in the opposite direction to the direction of rotation of input shaft \1.

It is also apparent that regardless of the direction of rotation of input shaft 1, with divider disk 15-15 pitched toward alignment with the letters F-F in FIGURE 1, the output shaft 2 will rotate in the same direction as input shaft 1, and with divider disk 15-15 pitched toward alignment with the letters R-R in FIGURE 1, rotation of output shaft 2 will be opposite to that of input shaft 1.

It will be understood that the ratio of speed of the input shaft 1 to the speed of the output shaft 2 will be inversely proportional to the ratio of the displacement of the driving pump to the displacement of the driven pump and that the output-input torque ratio will be directly proportional to the input-output speed ratio, neglecting frictional losses. In the illustrated embodiment of the transmission, a one to one forward ratio of speed between input shaft 1 and output shaft 2 occurs when the plane of rotation of divider disk 15-15 is in alignment with the letters F-F in FIGURE 1.

Pitch control of divider disk 15-15 is effected by control of the movement of disk support assembly 16- 16' about its transverse axis by permitting oil under pressure to enter either control chamber A through channel 48 or control chamber B through channel 49 and to exit from the other (see FIGURE 1). Disk support assembly 16-16 has projections or vanes 50 (FIGURES 9, 10, 11) which together with cylindrical surfaces 51 (FIG- URES 17, 18, 19) in end plate 6 and the spherical peripheral surfaces 51' of control chambers A and B (FIGURE l1) forms a fluid seal to prevent the unrestricted flow of oil between control chambers A and B. The axial center line of said cylindrical surfaces 1 coincides with the transverse axis of rotation of disk support assembly 16-16. Vanes 52 are similar to vanes 50 but are not necessarily designed to provide oil sealing contact with any surfaces of the housing, the essential purpose of said vanes 52 being to balance the weight of vanes 50 in respect to the longitudinal axis of the transmission when the inner housing is rotating, although said vanes 52 also serve to strengthen the disk support assembly 16-16 at points where needed.

To increase the input-output forward speed ratio or to decrease the input-output reverse speed ratio, oil under pressure is permitted to enter control chamber B through channel 49 from the control pressure source and a like amount of fluid is permitted to exit from control chamher A through channel '48 and to return to the sump or inlet of the control pressure source, which action causes disk support assembly 16-16 to move in the direction toward alignment of the transverse center plane of said divider disk 16-16 with the letters F-F of FIGURE 1. To decrease the input-output forward speed ratio or to increase the input-output reverse speed ratio, oil under pressure is permitted to enter control chamber A through channel 48 and to exit from control chamber B through channel 49 and to return to the sump or inlet of the control pressure source, which action causes disk support assembly to move in the direction toward alignment of the transverse center plane of said divider disk 16-16 with the letters R-R of FIGURE 1. Annular grooves 48 and 49' are provided in end piece 3 of the outer housing as a means of effecting hydraulic communication between a control system located externally to the inner housing and

channels

48 and 49 respectively under conditions when the inner housing is rotating as well as when said inner housing is held motionless in re- 8 spect to the outer housing as shown in FIGURES 1 and 2.

In order to obtain a mechanical indication of the transmission input-output speed ratio which might be required for a follow-up type of control system, a ratio indicating linkage is provided as shown in FIGURE 6. This linkage starts with a pair of

gears

53 and 54 fixed on shaft 55 and mounted so that gear 53 meshes with a circular gear segment 56 (FIGURE 11) on disk support assembly 116-16 and gear 54 meshes with gear rack 57 mounted in sleeve 58 inside dusting core 13. When disk support assembly 16-16 rotates about its transverse axis a proportionate angular rotation, corresponding to the gearing ratio between said gear '53 and said circular gear segment 56, is induced in gear 54 causing gear rack 57 to move longitudinally together with sleeve 58 to which it is mechanically fixed. Shaft 59 is threaded into sleeve 58 at one end (and therefore moves longitudinally with said sleeve 58) and has fixed in position on the other end of shaft 59 by retaining nut 61 a bearing assembly which is slidably mounted in output shaft 2 (FIGURE 2). An indicating sleeve 62 having thereon a rim 66 is slidably mounted on output shaft 2 and mechanically fastened to bearing assembly 60 by screws 63 which are free to move longitudinally in slots 64, said indicating sleeve 62 thereby being caused to move longitudinally as a unit with bearing assembly 60. A lever assembly consisting of two

arms

65 and 65, the extremities of which bear against rim 66 of indicating sleeve 62 fixed on a shaft 67 to which a third arm 68 is rigidly fastened extends the linkage away from the output shaft to the control mechanism. Shaft 67 is trunnioned on a transverse axis so that its center line is the pivot point for the lever assembly which is caused to pivot about its axis to follow the movement of indicating sleeve 62 by spring 85 or other equivalent means. A rod 69 is pivotally mounted on the end of arm 68 so as to translate the rotative movement of said arm 68 into longitudinal motion for utilization by a transmission control means. Rotative motion could just as well be imparted to the transmission control means by appropriate gearing arrangement from shaft 67 of the lever assembly if said rotative motion were more suitable for use by the control means.

When disk support assembly 16-16 rotates about its transverse axis a proportionate angular rotation, corresponding to the gearing ratio between said gear 53 and said circular gear segment 56, is induced in gear 54 musing gear rack 57 to move longitudinally together with sleeve 58 to which it is mechanically fixed. This movement is in turn transmitted through the linkages described until rod 69 moves the indicating mechanism to the corresponding indication.

Bypass control piston 70 (FIGURES 6 and 7) forming the end of shaft 59 which attaches to sleeve 58, contains an annular groove 71, the purpose of which is to bypass a nominal amount of oil from the high pressure channel in ducting core 13 via

orifices

72 and 72 to the low pressure channel in said ducting core 13 when a load connected to output shaft 2 is being accelerated from the motionless state in either the forward or reverse direction. Bypass control piston 70 acts in cooperation with orifices 72 and 72' to form a progressively smaller restriction to the bypass oil path as divider disk 15-15 is increased in pitch in either the forward or reverse direction from the zero pitch angle up to a predetermined pitch angle at which point the bypass path will be completely obstructed by the smooth cylindrical surface of the bypass control piston. By this means, torque in varying degrees may be applied to a motionless load without stalling the power source, thereby permitting smoother starts. For some forms of application of the transmission a bypass valve obviously would not be required.

Shuttle valve 73 (FIGURES 6 and 7) may be mounted in the ducting core 13 by threaded means as shown and comprises a valve housing 74 and shuttle spool 75. The

valve on one end is exposed to fluid channel 45 which is normally the high pressure channel and on the other end is exposed to fluid channel 45' which is normally the low pressure or return channel. The normally high pressure and return channels 45 and 45' may alternate with each other pressure-wise, depending upon the mode of operation of the transmission, however, the shuttle spool 75 is always forced against the valve seat on the high pressure side thus closing off the high pressure from the center of ducting core 13; on the oher hand the low pressure side is always in hydraulic communication with the center of the ducting core 13 so that except for the pressure drop due to oil flow through the communication channeling, the low pressure or return

prime channel

45 or 45 will he at the same fluid pressure as the center of ducting core 13. By this device a changing pressure may be applied to the

return fluid channel

45 or 45 through channel 76, located at the center of input shaft 1 (FIGURES l and 2), togdiscourage any tendency toward cavitation by pumping action of the driving and driven pumps and also to make up any leakage of oil from the inner housing of the transmission. Annular groove 77 is provided in end piece 3 of the outer housing as a means of effecting hydraulic communication between an oil pressure source located externally to the inner housing and channel 76 under conditions when the inner housing is rotating as well as when said inner housing is held motionless in respect to the outer housing as shown in FIGURES 1 and 2.

Referring to FIGURE 2 it will be apparent that spring loaded piston 78 may be a pressure regulating device which will maintain a certain changing pressure depending upon the spring characteristics and certain other factors when oil is forcibly circulated through channel 79 at the center of shaft 59. Oil flow would be through oil channel 76 from annular groove '77 into the center of ducting core 13 from where it would be transmitted through channel 79 located at the center of shaft 59 to the face of piston 78, which would cause the piston to move toward the right in FIGURE 2 until orifices 80 were exposed to the extent necessary to pass the volume of oil being circulated. Oil would flow through orifices 89 into the hollow space in output shaft 2 partially occupied by bearing assembly 60 from where said oil would then pass through slots '64 and eventually along the lower side of the outer housing where it would he returned to the sump through channel 81 in end piece 3 (FIGURE 1). Circulation of oil in this manner may also provide a means of transferring heat from the transmission proper to the sump or to a heat exchanger; however, under most conditions the power loss in the transmission will he sufficiently low as to render the use of a large capacity heat exchanger unnecessary.

The inner housing comprising parts aforementioned is journaled at the input shaft end on hearings 82 and at the output shaft end on hearings 83, said inner housing being disposed to rotate under certain conditions as described herewith. When operating at an input-output forward speed ratio of one to one, the reaction forces imparted to said inner housing through divider disk 15-15 are equal and opposite to the reaction forces imparted to said inner housing through divider disk 18-18', said forces effectively cancelling each other, thus permitting internal frictional forces to rotate said inner housing without the necessity for fluid locking. Assuming some leakage of fluid which bypasses the normal circulatory paths provided in the transmission, output shaft 2 will experience a slippage in speed relative to the speed of input shaft 1, in which case it is obvious that the inner housing cannot rotate at synchronous speed with both input shaft 1 and output shaft 2. Dynamic forces acting on divider disk 15--15' of the driving pump and divider disk 1818 of the driven pump when the two are in asynchronous motion cause the inner housing to rotate at a speed intermediate between the speeds of the input and output shafts due to the uniformity of pressure in directly communicative spaces.

Circulation of oil with the inner housing in synchronous rotation will be near zero; however, this will not impair torque transfer inasmuch as the torque transfer is a function of the net resultant hydraulic forces acting on vanes 22 and 22A of the rotors 14 and 17 and not a function of the oil transfer.

When the input-output forward speed ratio is greater than one-to-one, and when the power source is supplying torque to input shaft 1, greater torque is applied to driven pump rotor 17 then is applied to driving pump rotor -14, resulting in a net reverse reaction force acting on the inner housing'through the respective divider disks 18-13 and -1616 which tends to cause said inner housing to rotate in the opposite direction to the rotation of the input and output shafts. The function of over-running clutch 19, shown diagrammatically in FIG- URES 1 and 2, is to prevent the reverse rotation of the inner housing but to permit free forward rotation.

-When the input-output reverse speed ratio is greater than one-to-one, and when the power source is supplying torque to input shaft 1, greater torque'is applied to driven pump rotor 17 than is applied to driving pump rotor 14, and the net reaction force acting on the inner housing in this case tends to cause said inner housing to rotate in the forward direction which is undesirable inasmuch as said forward rotation would cause an effective reduction in the output-input torque ratio. Band 21, when caused to contract around the drum 20 with adequate force by hydraulic servo means or other suitable means, prevents forward rotation of the inner housing.

Under conditions when the impelling force is transmitted via output shaft 2 to rotor 17, thence to rotor 14 and then via input shaft 1 to the power source whichin this case becomes the load, as for example when a vehicle is descending a grade, and it is desired to increase the load by forcing the power source to a higher speed by increasing the transmission input-output speed ratio,

hydraulic forces tend to rotate the inner housing in,

the forward direction, which rotation if permitted to occur would lower the input-output speed ratio; therefore it is again desirable for the inner housing to be held against forward rotation by constriction of the hand 21 around the drum 20.

Shield 84 is provided for the purpose of reducing windage losses when the inner housing is rotating and is designed to cover the circumferentially asymmetrical portion of the inner housing.

Possibilities for many different configurations of control systems exist for adapting a transmission of this type to automotive, industrial or other applications. FIG- URE 3-1 is a simplified schematic drawing of a control system which might be used to automatically control the input-output speed ratio of the transmission in automotive vehicles as a function of output shaft speed and torque demand, or input shaft speed under certain conditions. The control system herein described is jointly electrical, mechanical and hydraulic in operation.

The hydraulic portion includes an oil pressure source and a follow up valve assembly 101 containing a valve spool 102 which is axially slidable in valve sleeve MP3, said valve sleeve 103 being slidably mounted in cylinder 164. The

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105 and 106 on the valve spool 102 are of such width and spacing so as to completely occlude oil inlet channel 107 and oil return channel 108' when aligned 5.? shown in FIGURE 31, but to permit circulation of oil from the pressure source 106 through the control passages and back to the sump when said valve spool 102 is displaced axially in either direction in respect to valve sleeve 103. When valve spool 102 is moved to the left in FIGURE 31, oil under pressure from inlet channel 187 is permitted to enter control chamber B through oil channel 109 and oil is permitted to return from control chamber A to return channel 108 through oil passage 110 as indicated by the solid arrows. By this means disk support assembly 1616' moves in the direction of increasing forward pitch or decreasing reverse pitch causing arm 68 of the ratio indicating linkage to move in the direction of the solid arrow. This action continues until inlet channel 107 and return chan nel 108 are realigned with

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105 and 106 at which time oil circulation into control chamber B and out of control chamber A is blocked. If valve spool 102 is moved to the right in FIGURE 31 oil is permitted to enter control chamber A under pressure and to exit from control chamber B as indicated by the dotted arrows. This causes disk support assembly 16-16 to move in the direction of decreasing forward pitch or increasing reverse pitch causing arm 68 in this case to move in the direction of the dotted arrow. This action continues until inlet channel 107 and return channel 108 are again realigned with

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105 and 106 at which time oil circulation into control chamber A and out of control chamber B is blocked. It is therefore apparent that when valve spool 102 is moved axially in either direction, that disk support assembly 16-16 is caused by hydraulic means to follow said axial movement of valve spool 102 by a corresponding angular amount in the direction which will tend to maintain alignment of

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105 and 106 with inlet channel 107 and return channel 108.

An electric motor 111 is provided to effect axial movement of valve spool 102 through cooperation of the threaded parts of shaft 112 and spool shaft 113. Spool shaft 113 should be designed for axial movement only. Motor 111 is controlled electrically by commutator-brush assembly 114 containing four pairs of commutator-brush rotors as shown schematically in FIGURE 32; viewing from left to right, one pair for Drive, one pair for Neutral, one pair for Low and one pair for Reverse. Essentially the purpose of commutator-brush assembly 114 is to compare the actual input-output speed ratio of the transmission at any instant with what the ratio should be as a function of transmission output shaft speed, torque demand, and selected speed range and to control motor 111 in such a way as to cause said motor 111 to drive valve spool 102 to the position which will correct any discrepancy.

Shaft 115 is geared at one end to spool shaft 113 and is directly coupled to the Drive, Neutral, and Low brush arms or rotors and is coupled through reverse gearing (not shown) to the Reverse brush arm and thereby causes angular movement of said brush arms when spool shaft 113 moves axially. Shaft 116 is geared at one end to control rod 117 and is coupled by suitable means to the Drive, Low, and Reverse commutation rotors so as to provide angular rotation in direct proportion to axial movement of said control rod 117 up to certain predetermined limits for each commutation rotor. For example, the limit of travel for the Drive commutation rotor might be adjusted to correspond to an input-output speed ratio of one-to-one for the transmission, whereas the limit of travel for the Low and Reverse commutation rotors might arbitrarily be adjusted to correspond to an input-output speed ratio of two-to-one for the transmission. The Neutral commutation rotor is fixed in position corresponding to an input-output speed ratio of infinity, or zero output for the transmission driving pump.

Selector switch 118 may be designed to permit manual selection of any one of the four switch contacts of the corresponding pairs of commutator-brush rotors and to provide electrical contact with the brush arm of the one selected. Only the selected pair of commutator-brush rotors may effect control of motor 111. In operation, when control action causes the ring F of the selected pair of commutator-brush rotors to come in contact with the brush, a closed circuit is established between the positive and negative poles of the vehicle electrical system via the selector switch 118, brush arm, ring F and the winding of relay F. Relay F then establishes electrical connection to motor 111 to cause said motor 111 to rotate in the direction which will cause valve spool 102 to move to the left viewing FIGURE 31, until the gear 119 and shaft 115 are rotated counter-clockwise by an amount which will cause the brush arm to move the brush to the gap position (FIGURE 32) between the rings F and R thus opening the electrical circuit which action causes the motor to stop. When control action causes the ring R of the selected pair of commutator-brush rotors to come in contact with the brush, a closed circuit is established between the positive and negative poles of the vehicle electrical system via the selector switch 118, brush arm, ring R and the winding of relay R. Relay R then establishes electrical connection to motor 111 to cause said motor 111 to rotate in the direction which will cause valve spool 102 to move to the right viewing FIGURE 31 until the gear 119 and shaft 115 are rotated clockwise by an amount which will cause the brush arm to move the brush to the gap position 120 between the rings F and R thus opening the electrical circuit which action causes the motor to stop.

Since the Drive, Neutral, and Low brush arms are directly coupled to shaft 115, rotation of said brush arms is naturally in the same direction as for shaft 115; however, inasmuch as the Reverse brush arm is coupled to shaft 115 through reverse gearing, rotation of said Reverse brush arm is in opposition to the rotation of shaft 115. The reverse gearing of the Reverse brush arm together with the oppositely oriented F and R rings permits the Reverse commutation rotor to move counterclockwise in respect to the Neutral reference position the same as for the Drive and Low commutation rotors and to provide the same transmission input-output ratio control characteristics in the Reverse speed as in the forward Low speed. The Neutral reference position for the commutation rotors is defined by vertical orientation of the gap position 120 as shown by the commutation rotor for Neutral in FIGURE 32. FIGURE 32 shows the selector switch 118 in the Low position in which a forward input-output speed ratio for the transmission is indicated by a counter-clockwise displacement of the Drive, Neutral, and Low brush arms in respect to the vertical Neutral reference. With the commutation rotors in the same position, it is apparent that if selector switch 118 was put in the Reverse position that the servo action previously described would cause valve spool 102 to move to the position representing a reverse transmission speed ratio equal to the forward speed ratio as indicated in FIGURE 32 and the Reverse brush arm would then be at the same counter-clockwise displacement angle in respect to the Neutral reference as for the Drive, Neutral and Low brush arms shown in FIGURE 32 and that the latter brush arms would then be at the same relative position as shown for the Reverse brush arm in FIGURE 32.

From the foregoing description it will be understood that input-output speed ratio control of the transmission in either forward or reverse output speeds is effected by control rod 117. As a means of causing said control rod 117 to move as a function of output shaft speed and torque demand or input shaft speed, an output speed governor 121, a vacuum modulator 122 and an input speed governor 123 are provided.

Governor 121 may be a centrifugally operated device geared to or otherwise coupled to the transmission output shaft 2 and coupled to control rod 117 as shown in FIGURE 31 to cause said control rod 117 to move axially to the left, viewing FIGURE 31, as a function of output shaft speed. Vacuum modulator 122 may be coupled to governor 121 in a suitable manner to oppose the action of said governor 121 as a function of engine manifold vacuum and thus provide higher input-output transmission ratios with decreasing manifold vacuum or increasingly open car-bureator throttle positions. With output shaft 2 motionless such as would occur when a vehicle is at a standstill, governor shaft 124 may be set at such a position as to cause the transmission pump to be at the zero output or neutral position when spring 125 is holding control rod 117 toward the right, in FIGURE 31, to the limit imposed by shoulder 126 resting against the lip of sleeve 12'7.

Governor 123 may be a centrifugally operated device geared or otherwise coupled to the transmission input shaft 1 and coupled to control rod 117 as shown in FIG- URE 31 to cause said control rod 117 to move axially to the left viewing FIGURE 31 as a function of input shaft s'pee Governor 123 and associated coupling linkage may be designed such that at a nominal input shaft speed, as for example the idling speed of a gasoline engine, shaft 128 will remain at the extreme right hand position, in FlGURE 31, but that at increasingly higher speeds of input shaft 1, shaft 128- will move to increasingly more left hand positions, viewing FIGURE 31, causing control rod 117 to move leftwardly with said shaft 128 against the tension of spring 125 until a predetermined maximum left hand position is reached, for instance corresponding to a transmission input-output ratio of four-to one, at which point shaft 128 will be held against further leftwardly movement; however, control rod 117 will not be restrained to further leftwardly movement resulting from leftwardly movement of sleeve 127. This arrangement will permit smooth starts for vehicles at a standstill, when selector switch 118 is in the Drive, Low, or Reverse positions by permitting governor 122 to exercise control up to a predetermined limit, for instance corresponding to a vehicle road speed of five miles per hour, and permitting governor 121 in conjunction with a torque-demand sensing element such as vacuum modulator 122 to effect transmission control at higher road speeds. As described heretofore, bypass control piston 70 will prevent positive circulation of oil between the driving pump and driven pump of the transmission when the input-output speed ratio is above a predetermined value in forward or reverse output speeds such as would be the case when an automotve vehicle is being accelerated from a standstill. I

A control means (not shown) of any conventional form may be provided for constricting band brake 20 around drum 21 when either the Low or Reverse positions are selected on the selector switch .118. An interlock switch may be provided to prevent starting of the engine except when disk support assembly 1616 is at the zero pitch or zero output position.

The control system herein described is suited for automatic operation of the transmission for automotive application without any additional fluid coupling interposed in the power train. The design of the transmission and control system for use with a slippage type of fluid coupling interposed between input shaft 1 and the engine crankshaft could dilfer in several details.

An obvious constructional variation in the illustrated embodiment would be the physical separation of the driving pump and driven pump with each having its own housing and hydraulically connected by lengths of tubing or conduit. Such a variation would be equivalent to cutting the transmission at line X-X ofFIGURE 1 into two parts, each part then being adapted to make appropriate connections with the interconnecting high pressure and return tubes or conduits and removing the outer housing. For instance, by locating the driving pump near the engine and the driven pump near the rear axle of an automobile, the need for the conventional drive shaft and associated universal joints would be eliminated thus making possible a fiat floor design for automobiles. By providing two driven pumps or one for each rear wheel, the conventional dilferetnial gearingcould be also eliminated.

A further possibility would be to integrally combine the driven pumps with the wheels.

C-onstructional variations such as four wheel drive and front wheeldrive in lieu of rear wheel drive could more readily be implemented with a hydraulic system than with conventional mechanical power coupling means. The possibility also exists for combining the braking means with the propelling means.

While there is given above a certain specific example of this invention and its application in practical use, it should be understood that this is not intended to be exhaustive or to be limiting of the invention. On the contrary, this illustration and explanation herein are given in order to acquaint others skilled in :the art with this invention and the. principles thereof and a suitable manner of its application in practical use, so that others skilled in the art may be enabled to modify the invention and to adapt and apply it in numerous forms each as may be best suited to the requirement of a particular use.

I claim:

1. A variable speed transmission of the hydraulic type comprising a housing having therein a driving pump chamber and a driven pump chamber; input and output shafts respectively mounted in said chambers; driving and driven pump assemblies respectively mounted in said chambers, said assemblies including a rotor connected to said input and output shafts and a plurality of radial vanes mounted on said rotors; a divider disk supporting assembly mounted in said housing and having therein an annular groove, a :divider disk having a plurality of radial slots therein mounted in said groove and adapted to intermesh with said vanes, a plurality of peripheral chambers on each side of said divider disk in the portion thereof disposed in said annular groove, each of said peripheral chambers being connected by a duct to a corresponding pump chamber formed between adjacent vanes on the opposite sides of said divider disk, said disk support assembly being pivotally mounted about a diameter thereof in said housing about said driving pump assembly; a ducting core extending from said driving pump chamber to said driven pump chamber, said ducting core having at least a pair of channels out therein to sequentially connect together corresponding pump chambers of said rotor assembly whereby oil may flow from one to the other; and control means for varying the pitch of said driving pump divider disk assembly.

2. A device as described in claim 1 wherein the outer edge of each vane of said driving pump rotor has a slot cut therein and said divider disk has at the center of each of said radial slots a cooperating pin adapted to engage in said slot, said slot being contoured to maintain said disk in proper phase relationship to said rotor.

3. A variable speed transmission of'the hydraulic type comprising a housing having therein a driving pump chamber and a driven pump chamber; input and output shafts respectively mounted in said chambers; driving and driven pump assemblies respectively mounted in said chambers, said assemblies including a rotor connected to said input and output shafts, and a plurality of v-anes mounted on said rotor; a divider disk supporting assembly mounted in said housing and having therein an annular .groove; a divider disk mounted in said groove; a plurality of radial slots cut in said disk and adapted to intermesh with said vanes mounted on said rotor; said disk supporting assembly forming with said rotor and disk a plurality of pump chambers on each side of said disk; a plurality of peripheral chambers on each side of said divider disk in the portion thereof disposed in said annular groove, each of said peripheral chambers being connected by a duct to a corresponding pump chamber; said disk support assembly in said driving pump assembly being pivotal-1y mounted in said housing about a diameter thereof disposed at right angles to the axis of said housing, a ducting core extending from said driving pump chamber to said driven pump chamber; a plurality of channels cut in said ducting core to sequentially connect together corresponding pump chambers of said rotor assemblies whereby oil may flow from one to the other; control means for varying the pitch of said driving pump divider-disk assembly; an outer casing surrounding said housing, driving and driven pump assemblies; bearing means within said outer casing to permit rotation of said driving and driven pump assemblies therein; overrunning clutch means mounted between said outer casing and said housing to prevent reverse rotation thereof and hand brake means connected between said outer casing and housing to selectively prevent forward rotation of said housing.

4. A variable speed transmission of the hydraulic type comprising a housing having therein a driving pump chamber and a driven pump chamber; input and output shafts respectively mounted in said chambers; driving and driven pump assemblies respectively mounted in said chambers, said assemblies including a rotor having six radial vanes, connected to said input and output shafts, and a divider disk having six radial slots therein adapted to intermesh with said vanes mounted on said rotor; a divider disk supporting assembly having therein an annular groove adapted to receive the periphery of said divider disk, said disk supporting assembly forming about said rotor and disk an inner chamber Within said housing, having twelve pump chambers; six peripheral chambers on each side of said divider disk in the portion thereof disposed in said support-ing assembly annular groove, each of said peripheral chambers being connected by a duct to a pump chamber on the opposite side of said divider disk; said driving pump assembly also having said disk support assembly pivotally mounted in said housing about a diameter thereof, a ducting core positioned within said driving rotor assembly and extending from said driving pump chamber to said driven pump chamber, said ducting core having at least a pair of channels out therein to sequentially connect together corresponding pump chambers of said rotor assemblies whereby oil may flow from one to the other; control means for varying the pitch of said driving pump divider disk assembly; an outer casing surrounding said housing, driving and driven pump assemblies; bearing means fixed in said outer casing and having journaled therein said driving and driven pump assemblies; overrunning clutch means operatively mounted between said outer casing and said housing to prevent rotation thereof in one direction and hand brake means operatively connected between said casing and housing to selectively prevent rotation of said housing in the other direction.

5. A variable speed transmission of the hydraulic type comprising a housing having therein a driving pump chamber and a driven pump chamber; input and output shafts mounted respectively in said chambers; driving and driven pump assemblies respectively mounted in said chambers, said assemblies comprising a rotor having six radial vanes mounted about and connected to said input and output shafts, a divider disk having six radial slots therein adapted to intermesh with said vanes mounted on said rotor, a divider disk supporting assembly mounted in said housing and having therein an annular groove adapted to receive the periphery of said divider disk, said disk supporting assembly forming about said rotor and disk an inner chamber Within said housing, s-ix peripheral chambers formed on each side of said divider disk in the portion thereof disposed in said supporting assembly annular groove, each of said peripheral chambers being connected by a duct and valve mechanism to a corresponding pump chamber formed between adjacent vanes on the opposite side of said divider disk; said disk support assembly in said driving pump assembly being pivotally mounted in said housing about a diameter thereof disposed at right angles to the axis of said housing being contoured on the outer surface to form an oil sealing contact with the inner surface of said housing; a pair of vanes mounted in the plane of the axis of said support assembly and said shaft to form oil sealing chambers above and below the axis of said support disk assembly on at least one side thereof, duct means for introducing and withdrawing oil from said chambers formed between said housing and said supporting disk assembly whereby the pitch of said assembly may be controlled, a ducting core positioned within the shafts of said rotor assemblies and extending from said driving pump chamber to said driven pump chamber, said ducting core having at least a pair of channels cut therein to sequentially connect together corresponding pump chambers of said rotor assemblies whereby oil may fiow from one to the other; control means for varying the pitch of said driving pump divider disk assembly; angle indicating means geared to said driving pump disk supporting assembly and extending through said ducting core member to a movable collar on said output shaft, lever means interconnecting said collar to said control means; an outer casing surrounding said housing, driving and driven pump assemblies, bearing means mounted in said outer casing having said driving and driven pump assemblies journalled therein; overrunning clutch means mounted between said outer casing and said housing to prevent reverse rotation thereof and band brake means connected between said housing and inner chamber to selectively prevent forward rotation of said housing.

6. The device of claim 5 wherein said angle indicating means includes a sleeve, an annular channel in said sleeve, a plurality of orifices in the high and low pressure channels of said ducting core, said sleeve being positioned about said orifices when said driving pump dividerdisk assembly is near the zero pitch position to at least partially interconnect through said channel said orifices whereby a slight bypassing of fluid is obtained to prevent stalling of the prime power source when starting the driven pump under heavy loads.

7. A device as described in claim 5 wherein said ducting core has therein a valve and port assembly interconnecting opposite pressure channels thereof to the center of said ducting core, said valve mechanism being arranged to close oif the port leading to the high pressure side of said ducting core at any given time whereby oil may be added to the low pressure side of said ducting core through the center thereof.

8. A variable speed transmission of the hydraulic type comprising a housing having therein a driving pump chamber and a driven pump chamber; input and output shafts respectively mounted in said chambers; driving and driven pump assemblies respectively mounted in said chambers, said assemblies including a rotor connected to said input and output shafts and a pluarlity of radial vanes mounted on said rotors; a divider disk supporting assembly mounted in said housing and having therein an annular groove, a divider disk having a plurality of radial slots therein mounted in said groove and adapted to intermesh with said vanes, said disk support assembly being pivotally mounted about a diameter thereof in said housing about said driving pump assembly; a ducting core extending from said driving pump chamber to said idriven pump chamber, said ducting core having at least a pair of channels cut therein to sequentially connect together corresponding pump chambers of said rotor assembly whereby oil may flow from one to the other; and control means for varying the pitch of said driving pump divider disk assembly.

9. In a hydraulic variable displacement transmission of the type having a driving pump connected to a prime power source and a driven pump connected to the source to be moved, control means comprising a pitch control mechanism connected to said driving pump means, said pitch control mechanism having forward, neutral, and reverse positions; a feed back loop for indicating the pitch of said driving pump in said pitch control mechanism; motor means operatively connected to said pitch control for moving said pitch control to the desired position; a commutator brush assembly having a plurality of commutation stators and rotors; a load demand sensing element connected to said commutator brush assembly; switch means for selectively connecting to the desired commutation rotor; battery means connected through said switch means to said commutation rotors and load demand sensing elements connected to said commutator brush assembly whereby variations in load demand or switch control will cause said pitch control mechanism to vary the pitch of said driving pump so that the desired output will be delivered to the output shaft.

References Cited in the file of this patent UNITED STATES PATENTS 951,064 Erickson Mar. 1, 1910 Erickson Mar. 12, Cuny May 13, Haines May 4, McGill July 13, Dodge Mar. 13, Jakobsen Nov. 18, Kraft June 8, Cuny Oct. 12, Paulsrn-eier Oct. 12, Marshall Apr. 1,

FOREIGN PATENTS France June 4, Germany May 29,

Claims

1. A VARIABLE SPEED TRANSMISSION OF THE HYDRAULIC TYPE COMPRISING A HOUSING HAVING THEREIN A DRIVING PUMP CHAMBER AND A DRIVEN PUMP CHAMBER; INPUT AND OUTPUT SHAFTS RESPECTIVELY MOUNTED IN SAID CHAMBERS; DRIVING AND DRIVEN PUMP ASSEMBLIES RESPECTIVELY MOUNTED IN SAID CHAMBERS, SAID ASSEMBLIES INCLUDING A ROTOR CONNECTED TO SAID INPUT AND OUTPUT SHAFTS AND A PLURALITY OF RADIAL VANES MOUNTED ON SAID ROTORS; A DIVIDER DISK SUPPORTING ASSEMBLY MOUNTED IN SAID HOUSING AND HAVING THEREIN AN ANNULAR GROOVE, A DIVIDER DISK HAVING A PLURALITY OF RADIAL SLOTS THEREIN MOUNTED IN SAID GROOVE AND ADAPTED TO INTERMESH WITH SAID VANES, A PLURALITY OF PERIPHERAL CHAMBERS ON EACH SIDE OF SAID DIVIDER DISK IN THE PORTION THEREOF DISPOSED IN SAID ANNULAR GROOVE, EACH OF SAID PERIPHERAL CHAMBERS BEING CONNECTED BY A DUCT TO A CORRESPONDING