US3247800A

US3247800A - Pump

Info

Publication number: US3247800A
Authority: US
Inventors: John F Campbell
Original assignee: Individual
Current assignee: Geosource Inc
Priority date: 1959-07-02
Filing date: 1962-08-29
Publication date: 1966-04-26
Anticipated expiration: 1983-04-26

Description

April 26, 1966 J. F. CAMPBELL. 3,247,800

PUMP

Original Filed July 2, 1959 7 Sheets-Sheet 1 FREE WHEELING RANGE POWER RANGE INVENTOR. JOHN E CAMPBELL ATTOR NE YS April 26, 1966 J. F. CAMPBELL 3,247,800

PUMP Original Filed July 2, 1959 7 Sheets-Sheet 3 5 FIG. 5

INVENTOR. JOHN E CAMPBELL ATTORNEYS April 26, 1966 J. F. CAMPBELL 3,247,800

PUMP

Original Filed July 2, 1959 7 Sheets-Sheet 4 FIG. 7

I88 I y 39 205 fl I87 1'85 \t '97 IBTA 1 FIG I4 ee \L 5? FIG. I5

22 INVENTOR.

JOHN E CAMPBELL M, MI QM "M ATTORNEYS April 1966 J. F. CAMPBELL 3,247,800

PUMP

Original Filed July 2, 1959 '7 Sheets-Sheet 5 F|G.|8 FIG. 20

FIG. l6

ATTOR N EYS April 26, 1966 J. F. CAMPBELL 3,247,800

PUMP

Original Filed July 2, 1959 7 Sheets-Sheet a FIG. 2|

INVENTOR. JOHN E CAMPBELL AT TO R N EYS April 26, 196$ J. F. CAMPBELL 3,247,800

PUMP

Original Filed July 2, 1959 '7 Sheets-Sheet 7 so pnsssm CHSBB was so X mess, zvs

o moo zcoo 5000 4000 5050 ENGINE RPM PTA FIG. 27

VAC, ms. H6 1 ENGINE RPM VAC INS. HG

FIG. 28

mess, PSI

INVENTOR. JOHN F. CAMPBELL will drums, are not required.

United States Patent PUMP John F. Campbell, Beech Knoll, Timheridge Trail, Gates Mills, Ohio Original application July 2,1959, Ser. No. 824,506, now Patent No. 3,059,416, dated Oct. 23, 1962. Divided and this application Aug. 29, 1962, Ser. No. 220,243 13 Claims. (Cl. 103-417) This application represents a divisional application of my copending parent application entitled Fluid Drive and Brake System, Serial No. 824,506, filed July 2, 1959, now Patent No. 3,059,416, issued October 23, 1962.

The present invention relates generally, as indicated, to a pump and more particularly to a pump assembly for a fluid drive and brake system for a power transmission and braking in motor vehicles such as passenger cars, trucks, buses, etc.- More particularly, the present invention relates to an all-fluid power transmission system characterized in that the conventional transmission, differential, live axle, and the separate hydraulic brake system, with its master cylinder, wheel cylinders, brake shoes and brake In such fluid power transmission system, a prime mover, be it an internal combustion engine, a diesel engine, a steam or gas turbine, a dynamoelectric machine, or the like, drives such fluid pump assembly and, in turn, such pump assembly drives a fluid motor system which, in the case of a motor vehicle, comprises rotary fluid motors mounted in the rear wheels, or in the front wheels, or preferably in all the wheels.

Numerous attempts have been made heretofore to provide such hydraulic power transmission wherein an enginedriven pump supplies wheel mounted motors, but to my knowledge, these all have had the serious shortcoming of either lack of proper control of capacity or speed, or of poor overall operating efiiciency owing to a variety of dif ferent reasons such as hydraulic slip, restriction of the pump inlet to achieve control, changing of engine load by changing of power absorbed by the pump, lack of relation of pump capacity to the variables attuned to eflicient engine operation, lack of coordination of engine rpm. and carburetor throttle position or manifold pressure for eflicient engine operation, etc.

It is accordingly a principal object of this invention to provide such fluid drive and brake system which comprises an engine-driven variable delivery positive displacement pump and fluid motors driven by the pump, the pump being provided with control means effective to provide for eflicient operation of the engine under all conditions of engine speed and load.

Other objects and advantages of the present invention will become apparent as the following description proceeds.

To the accomplishment of the foregoing and related ends, the invention, then, comprises the features hereinafter fully described and particularly pointed out in the claims, the following description and the annexed drawings setting forth in detail certain illustrative embodiments of the invention, these being indicative, however, of a few of the various ways in which the principle of the invention may be employed.

In said annexed drawings:

FIG. 1 is a side elevation view of my power transmission 1pgmpgassembly as viewed from the right-hand end of FIG. 5 is a cross-section view, on somewhat enlarged scale, taken substantially along the line 5-5, FIG. 4;

FIGS. 6 to 15 are detail cross-section views taken respectively along the lines 66, FIG. 5; 7-7, FIG. 4; 8 -8, FIG. 5; and 99, 101tl, 11-11, 1212, 1313, FIG. 4; 14-14, FIG. 3; and 1515, FIG. 4;

FIG. 16 is a central diametrical cross-section view showing a preferred form of wheel-mounted fluid motor;

FIG. 17 is a transverse cross-section view taken substant-ially along the line 17-17, FIG. 16;

FIGS. 18 to 20 are detail cross-section and developed views of the preferred form of valving employed in connection with the pump of the engine-driven pump assemy;

FIGS. 21 to 24 are cross-section views and developed views illustrating the preferred form of valving employed in each wheel mounted fluid motor;

FIG. 25 illustrates a further modification wherein manifold vacuum is employed to provide certain control characteristics; and

FIGS. 26 to 28 are graphs with engine r.p.m. plotted against pressure p.s.i. (FIG. 26) and presssure p.s.i. plotted against manifold vacuum in inches mercury (FIGS. 27 and 28).

I. The fluid drive and brake system as a whole 1 (FIGS. 1 and 2 Referring to FIG. 1, there is shown therein in dot-dash lines a typical passenger car 1 with which the present invention may be used to drive and to brake the four wheels 2, said car being powered as by a spark-ignition, four-cycle engine 3 and being steered as by a steering wheel 4. Operatively connected between the engine 3 and fluid motors (not shown in FIG. 1) mounted in the wheels 2 is the hydraulic unit 5 which transmits engine power to the wheels and which is controlled as by the foot-operated pedal 6 in the car 1. The pedal 6, as later described, also serves as the brake pedal to effect braking of the wheels 2. Although the unit 5 is herein shown mounted on the rear end of the engine block it may be mounted at any convenient place for driving by the engine crankshaft. Because the engine power is transmitted to the wheels '2 by way of fluid lines there is no need for the usual differential and live axles, nor for the universal joints and conventional drive shaft. Thus, substantial economies in cost and weight are effected and the floor of the passenger compartment need not have the usual longitudinal tunnel.

Referring to FIG. 2, one of the main elements of the present system, i.e. the hydraulic unit 5, the wheel motors, and the control elements for the engine 3 and unit 5 is the engine-driven variable delivery'positive displacement pump 1%, the capacity of which may be varied from zero to maximum in accordance with the fluid pressure in the pump control line 11 leading thereto. The pump inlet line 12 communicates with a vented sump 13 and the pump output line 14 leads to the respective motors 15 mounted in the vehicle wheels 2 by way of a check valve 16, a forward and reverse control valve 17, and either

line

19 or 20 depending on whether the wheel motors 15 are to be driven to propel the vehicle 1 in a forward or a backward direction. The other of said. lines 19 and20 is the return line through which the spent fluid is returned to the sump 13 through valve 17, return line 18, brake valve 21, and heat exchanger 22. Ta 23 or the like may be provided so that the front wheels may be operated by one pair of

lines

19 and 20 and so that the rear wheels may be similarly connected with another pair of lines 19 and 211 (not shown in FIG. 2). As is later described,

a) the delivery capacity of pump is preferably directly proportional to the fluid pressure in the line 11.

Downstream of the check valve 16 is a relief valve 24 which is normally closed and set to open only when the discharge pressure from the pump ll? is greater than the maximum operating pressure of the system.

The forward and reverse control valve ll7is normally in the forward position when the solenoid 25 thereof is in deenergized condition. However, when the lever 26 of the ignition switch 27 is in the On position and the lever 29 of the direction switch 28 is turned to the R (reverse) position, the solenoid 25 will be energized to actuate the valve 17 so as to reverse it and thus the wheel motors will be driven in reverse direction.

As shown in FIG. 2, each wheel 2 is mounted on a strut assembly 3.!) which allows the wheel shaft 31 to move up and down as the vehicle 1 travels on rough road. In the case of the front wheels 2, the mounting strut 32 may be rotated about its axis by a steering lever 33 adapted to be operatively connected to steering wheel 4 by conventional means. Thus, when the valve 17 is in forward F position, fluid under pressure will enter the fluid motor by way of the line 19 and the spent fluid will be returned to the sump 13 by way of the other line and return line 18; and, of course, when the solenoid of said valve 17 is energized, as aforesaid, fluid under pressure will be supplied to each motor 15 through the line 29 to cause the motor 15 to be driven in a backward or reverse direction, and again the spent fluid will be returned to the sump 13 by way of the other line 19 and return line 18.

Referring now to the control system, there is provided a pivotally mounted foot pedal 6 which, through linkage 34, actuates a cam shaft 35, the pedal 6 in FIG. 2 being shown in its'intermediate Free Wheeling Range. When it is desired to increase the engine r.p.m. and power output, the pedal 6 is depressed to the Power Range and when it is desired to brake the vehicle 1, the pedal 6 is swung upwardly to the Brake Range. Operated by the cams 36 on the cam shaft 35 is the engine throttle control assembly 37 which, as hereinafter explained, opens the throttle valve 38 according to a predetermined schedule as the pedal 6 is depressed through the Power Range. The second cam 39 operates the brake valve 21 when the pedal 6 is swung counterclockwise to the Brake Range position and, in essence, the brake valve 21 is a variable restrictor which progressively blocks the return of fluid from the wheel motors 15. Because the fluid is heated thereby, especially during quick stops, the heat exchanger 22 is provided ahead of the sump 13.

The third cam 40 on the cam shaft 35 controls operation of the pump it) as is presently to be explained, and the fourth cam 41 controls energization and deenergization of the solenoid 42 of a pump unloading valve 43.

Also driven by the engine 3 is an r.p.m. pump 45 which has its intake port in communication with intake line 12 and which has its delivery port connected to output line 46 leading to a chamber at one end of valve 47 and from the chamber into a preload check valve 53 which has an orifice or restriction 49 associated therewith.

As later explained, the capacity of the pump 45, the size of the orifice 49, and the size of the seat, the contour of the valve member, and the preload and rate of deflection of a spring within the check valve 48 are so designed as to achieve the desired program of pressure in the chamber of valve 47 versus the movement of the pedal 6. This is represented by the curve 124 in FIG. 25. In essence, the pressure in the valve chamber aforesaid may be said to be an indication of the actual r.p.m. of the engine drive shaft while the pressure in another chamber at the other end of said valve 47 is an indication of the desired r.p.m. of the engine drive shaft.

The engine drive shaft also drives a servo pump 51 which has its intake and delivery ports connected to intake line 12 and delivery line 52, a relief valve 53 serving to establish a constant pressure fluid supply to the variable restrictor valve 54 which is operated by cam 46. When the actual r.p.m. of the engine drive shaft is greater than that desired, as established by the variable restrictor valve 54 and orifice the pressure in chamber 139 will be greater than in the other chamber 133. The lower pressure leading into the right end of valve 47 through line 57 will render .the predominant pressure in the left end through line as effective to cause the valve member in valve 47 to move in one direction so as to allow fluid to flow under pressure through the line 58 and the unloading valve 43 to the variable capacity pump line 11 from the servo pump output line 52. Accordingly, the amount of fluid pumped by the pump 10 will increase to thereby decrease the speed of the engine drive shaft.

Conversely, when the actual r.p.m. of the engine drive shaft is less than that desired as indicated by the setting of valve 5 5, the pressure in chamber 138 will build up whereby the then predominating pressure in the right chamber of valve 47 will cause the valve member to move in the opposite direction. This allows the fluid undr pressure in the pump capacity control line 11 to be bled into the return branch 59 of the return line 18. The lowering of the pressure in the line H causes a decreased amount of fluid to be pumped by the pump It) thereby increasing the r.p.m. of the drive shaft.

In general, the operator sets up the desired r.p.m. through the position of the foot pedal 6 while simultaneously valve 47 operates automatically to increase or decrease the torque on the engine drive shaft thereby making the actual r.p.m. agree with the desired r.p.m.

When the solenoid 42 of the unloading valve 43 is energized either by shifting the lever 28 of direction switch 29 to neutral position N or by shifting the pedal 6 to Free Wheeling Range with cam 41 closing switch 60, capacity control line 11 is vented through branch return 61 to return line 18 whereby pump llt) operates at zero capacity. Displacement of the fluid motors 15 is accommodated by the free-wheeling check valve 62 connected between intake line 12 and delivery line 14 which leads to forward and reverse control valve 17.

Having thus described the general structure and operation of the system herein, reference will be made under appropriate headings of the details of the several components of the system.

Engine-driven pump assembly (FIGS. 3, 4 and 5) and components thereof (FIGS. 6 to 15 and FIGS. 18 to 20) As best shown in FIGS. 3, 4 and 5, the engine-driven pump assembly or hydraulic unit 5 has a shaft 65 journalled therein, as in needle bearings 66 or the like, and adapted to be coupled with the crankshaft of the engine 3 in the case of an internal combustion engine and, of course, with a comparable or equivalent output or drive shaft of an electric motor, gas turbine, diesel engine, etc. This unit 5 comprises a two-part housing including a main casting 67 and a cap part 68 secured together by screws 69, the cap part 63 being formed with a vented filler cap 70 leading into the hot oil gallery '71.

Formed in the hydraulic unit 5 is the sump 13 of annular form with fluid returned thereto through a filter 72 and heat exchanger 22 also preferably annular in form, and upstream of the filter 72 is a hot oil gallery 71 to which the spend fluid in heated condition, is returned for flow through the filter 72 and heat exchanger 22 into the sump. The heat exchanger 22 as best shown in FIGS. 5 and 15 may have cooling fluid circulated therethrough by way of the

openings

73 and 74 formed in casting 67.

The variable deliver positive displacement pump 10 (FIG. 5) and pump inlet valvz'ng (FIGS. 5 and 18 to 20) shaft, actuates a plurality of spring-biased radial plungers 82, herein six such plungers being employed. Each plunger 82 is guided for radial movement in a bushing 83 which preferably is shrunk fit in a radial bore in the housing part 67 and retained by snap ring 84. The bushing is helically grooved, as shown, to retain the tension spring 85. The plunger 82 is a slip fit in the bore of a capacity regulator sleeve 86 and the outside diameter of the sleeve is a slip fit in the radial bore of housing part 67. The'capacity regulator sleeve 86 also is helically grooved to retain the other end of the tension spring 85.

In FIG. 5 the sleeve 86 is shown in a position for 50% pump delivery capacity by reason of fluid pressure in the chamber 87 acting on the radially outer end of sleeve 86, thereby elongating the tension spring 85 and urging the inner end of sleeve 86 to a position such that when the plunger 82 has moved in one-half of its stroke against eccentric 80 by biasing spring 88, the spill groove 89 thereof will be inward of the inner end of sleeve 86. Thus, during the first half of the outward pumping stroke of the plunger 82 the fluid in the chamber 90 will flow through the spill groove 89 into chamber 91 and thence through passage 92 into the sump 13. When the pressure in the chamber 87 and in the line 11 (passage 11A, manifold 11B, and passage 11C) leading thereto from solenoid valve 43 and passage 58 is at a high value, the sleeve86 will contact the stop surface 93 and will thereby assume a position for 100% pump delivery capacity, that is, the spill groove 89 is within sleeve 86 during the entire outward pumping stroke of plunger 82. On the other hand, when the pressure in the chamber 87 and line 11 (11A, 11B, and 11C) is at a low value, the sleeve 86 will move outward under the influence of the tension spring 85 so as to contact inner end of bushing 83 whereby it is then in the zero pump delivery capacity position with the spill groove 89 inward of the inner end of the sleeve 86 during the entire pumping stroke. Accordingly, the position of sleeve 86 and, therefore the pump delivery capacity, is directly proportional to the pressure in the line 11 and chamber 87.

The volume of fluid pumped by each plunger 82 is a constant amount during each revolution of the drive shaft 65, however, the portion of the volume pumped through the delivery check valve 16 is dependent upon the position of the associated sleeve 86 and thus, on the pressure in the line 11. For example, with the sleeve in the 50% capacity position, as shown in FIG. 5, fluid in each plunger chamber 90 is in common communication with chamber 91 when the plunger 82 is at any position between the bottom end of its stroke and 50% of the maXimum outward stroke, and thus the spill groove 89 will be uncovered and the fluid in the cavity 90 will pass freely into chamber 91 and thence through the passage 92 which leads into sump 13. This is true whenever spill groove 89 is uncovered by the inner end of sleeve 86 because the resistance through the above circuit to the sump 13 is much less than through the associated delivery check valve 16 which is always biased shut by the combination load of the spring 93 and the operating pressure in the delivery line 14 (manifold 14A and the pressure passages 14B leading to forward and reverse control valve 17). According to the above, it can be similarly reasoned that when the capacity sleeve 86 is in the outermost zero capacity position, that is, in contact with the end of bushing 83, the spill groove 89 will be uncovered throughout the entire stroke of the plunger 82 and, therefore, all the fluid which is displaced during the inward stroke will return to the sump 13. On the other hand, when the capacity sleeves 86 are in the 100% capacity position, that is, in contact with the stop surfaces 93, the spill grooves 89 thereof will be covered throughout the entire pumping strokes of the respective plungers 82 and, therefore, all of the fluid displaced by respective plungers 82 will pass through the respective check valves 16 and into the manifold 14A and thence through passage 14B to the forward and reverse control valve 17.

The pump inlet valve 95 is drivenv by the drive shaft 65 and as best shown in FIGS. 5 and 18 to 20 the inlet valve port 96 is open to the inlet passages 97 only during the suction strokes of the respective plungers 92. Fluid is supplied through the port 96 without restriction from line 12 (opening 12A in valve 95 which opens into sump 13).

The pump inlet valve 95 as shown in detail in FIGS. 18 to 20 has the port 96 through the wall thereof and rotates in a sleeve 98 which may be shrunk fit in the housing part 67, said valve 95 being driven through the splined connection of the valve drive shaft 99 which, in turn, has a splined connection with the pump drive shaft 65. A development of the outside diameter of the sleeve 98 is shown in FIG. 19 and a development of the outside diameter of the valve 95 is shown in FIG. 20. Multiple passages 100 registering with pump intake passages 97 are connected to multiple static holes 101 by grooves 102 respectively. The grooves 102 connect passages 100 and holes 101, which are spaced 180 apart around from the passage. The area of each pair of holes 101 is equal to the area of the diametrically opposite passage 100. By reason of that construction, the hydraulic forces on the surface of the valve 95 are opposed by equal and opposite forces and therefore, are balanced out. This allows the valve 95 to rotate freely in the bore of the sleeve 98 regardless of the pressure in the ports 97 and in the passages 100. The chamber 12A in the valve 95 receives fluid from the sump 13 and this fluid is drawn through the port 96 and flows through the multiple passages 100 and related ports 97 to the respective pump chambers as the valve rotates to align the opening 96 with successive passages of sleeve 98.

The solenoid valve 43 (FIG. 7; also FIGS. 2-5) This valve 43 is formed as a part of the housing member 68 including a ported sleeve 105 and solenoid 42. held therein as'by means of snap rings as shown, the sleeve 105 and the solenoid retainer 106 being provided with suitable packing rings such as O-rings, to prevent leakage. The ported sleeve 105 is intersected by a pair of passages 58 and 11A, of which the passage 11A, best shown in FIG. 5, extend through the housing part 67 to an annular manifold 11B, and thence through radiating passages 11C to the respective chambers 87 associated with capacity regulating sleeves 86, whereby such passages 11A, 11B, and 11C constitute the pump capacity control line 11 referred to in connection with FIG. 2. The other passage 58 communicates with a port of the valve 47 as shown in FIG. 2 and as shown in detail in FIG. 11.

Reciprocable in the ported sleeve 105 is the valve spool 107 which is biased to the position shown by the spring 108, the stem 109 of the valve spool extending into the solenoid 42 so as to constitute an armature which is pulled toward the right when the solenoid 42 is energized, as hereinafter explained. When the solenoid 42 is deenergized, fluid under pressure entering from passage 58 flows around the groove of the spool 107 to the passage 11A.

When the solenoid 42 is energized the valve spool 107 is pulled toward the right to block such flow of fluid from passage 58 to passage 11A and to open the passage 11A to the hot oil gallery 71 by way of the open end of the ported sleeve 105. The solenoid 42 is thus energized when the direction switch lever 28 (FIG. 2) is moved to the neutral N position and is also so energized when the pedal 6 is moved to Brake Range position or to Free Wheeling Range, whereby the cam 41 on the cam shaft 35 closes the switch 60 which is wired in parallel with the aforesaid neutral N position of the direction switch 29.

The check and restriczor valve 4849 (FIG. 10; also FIGS 2,4 and 8) This valve is in the nature of a pressure regulating valve also formed in the housing part 68 and including a ported sleeve 115 held in place by a sealing plug 116 and snap ring 117 and provided with outlet ports 118 leading into the hot oil gallery 71 and an inlet 46 from r.p.m. pump 45 and valve 47 (FIGS. 2, 8, 10, and 11). The valve member 11? is provided with a guide stem 12! and a head 121 which has a clearance in sleeve 115 to form the orifice 49. The valve member 119 is biased by means of a spring 122 which, through a headed wire 123 secured to the valve member, normally tends to urge the latter to the position shown in FIG. 10. As the pressure in the passage 46 builds up the valve member 119 is urged to the left to bleed off an increasing volume of the fluid to the hot oil gallery '71 so as to maintain a prescribed pressure in the line 46 and in the chamber at the left end of valve 47 (FIG. 11). The capacity of rpm. pump 45, the size of the orifice 49, the size and contour of the valve head 121, and the preload and deflection rate of the spring 122 are selected to establish the program of pressure in line 46 versus movement of pedal 6 as represented by line 124 in FIG. 25.

The valve 47 (FIGS. 10, 11; also FIGS. 2, 3, 4 and 8) This valve 47, as the others, is formed as a portion of the housing part 68 and has a ported sleeve 136 therein which is formed with openings in register with passage 46 from the r.p.m. pump 45, with passage 52 from the servo pump 51, with passage 58 to valve 43, with return branch 59 to the hot oil gallery 71, and with passage 57 to the variable restrictor 54.

The ported sleeve 139 is closed at its ends by the spring abutment members 131 that are held in place together with sleeve 130 as by means of snap rings 132. Centered in the ported sleeve 130 by springs 133 is the valve spool 134 provided with three

lands

135, 136, 137 and a pair of intervening grooves.

When the spool 134 is in its centered position as in FIG. 11, the passage 58 is blocked by the middle land 136. When the pressure in the passage 57 and chamber 138 acting on the right-hand end of the spool 134 is greater than the pressure in the passage 46 and chamber 139 acting on the left-hand end, the spool 134, the latter will be shifted toward the left to establish metered flow (by reason of the metering portions of the middle land 136) from passage 58 to passage 59 thereby bleeding the control line 11 (11A, 11B, and 11C). On the other hand, when the pressure in the chamber 135 acting on the left-hand end of the spool 134 is greater than the pressure in the chamber 138 acting on the right-hand end, the spool 134 will be shifted toward the right to establish metered flow from passage 52 to passage 58 to build up a desired control pressure in line 11.

The operation of this valve 47 in establishing the pressure difl'erentials aforesaid in chambers 138 and 13) at the opposite ends of the spool 134 is controlled by the cam actuated regulating valve or variable restrictor 54, which is next to be described, and by the valve 48.

The cam actuated regulating valve 54 (FIG. 12; also FIGS. 2 and 8) This valve comprises a ported sleeve 145 held in place in the housing part 68 by means of snap rings as shown and mounted on the pedal operated cam shaft 35 is the cam 40 which engages the end of a metering valve 146 which is biased by spring 147 against the cam 40. Said metering valve 146 is provided with an intermediate neck and adjoining conical surface which cooperates with the passage 14% to vary the size thereof according to the position of the cam 40 and metering valve 146. Thus, when the cam 40' is swung in the clockwise direction, as viewed in FIG. 12, the tapered metering portion of the metering valve 146 will gradually restrict the fiow of fluid from passage 57 through chamber 56 and through passage 143 into the hot oil gallery 71.. The passage 57 just referred to is the same passage 57 that communicates 8 with the chamber 138 at the right-hand end of the spool 134 of the valve 47 in FIG. 11.

The passage 52 from valve 47 and servo pump 54 leads to the chamber 55 upstream of the orifice 149 and therefore fluid flows through the chamber 56 and passage 148 into the hot oil gallery 71 through ports 156 of the sleeve when the metering valve 146 is partly or fully open as shown in FIG. 12. Of course, when the metering valve 146 is only open a slight amount such that the flow area of passage or variable orifice 148 is less than the flow area of the orifice 149, fluid pressure will build up in the chamber 56 and passage 57 which leads to the right-hand end of the spool 134 of the valve 47 in FIG. 11 to thus urge the spool 134- toward the left. This will be described in detail.

The pumps 45 and 5] and the regulating valve 53 associat d with the latter (FIGS. 2, 5, 8 and 9) The r.p.m. pump 4-5 and the servo pump 51 are gear pumps arranged in tandem and driven by a shaft which is coupled to the rotary pump inlet valve drive shaft 99, which, in turn, is coupled to the engine-driven pump drive shaft 65. The shaft 155 is keyed to one of the pair of meshing gears 156 of the pump 45 and to one of the pair of meshing gears 157 of the pump 51.

These pumps 45 and 51 have

intake ports

158 and 159 respectively which communicate with the sump 13 and

delivery ports

160 and 161 which by way of the previously referred

passages

46 and 52, communicate respectively with the valves 47 and 48-49, and the

valves

47 and 53, as best shown in FIG. 8.

The regulating or relief valve 53 associated with the servo pump 51 is in the nature of a ball relief valve connected in parallel with the pump 51 as clearly shown in FIGS. 2, 8, and 9. The ported sleeve 165 of valve 53 is held in place by the plug 166 and snap ring 167 and has a seat for the ball 168 which is biased by spring 169 to closed position and which is adapted to be unseated to open communication between the delivery passage 52 from pump 51 and the return branch 59 that leads to the hot oil gallery 71 to maintain a prescribed pressure of the fluid delivered by the pump 51. In other words, the ball 168 is in seated position except when the pressure of the fluid delivered by the pump 51 is greater than a predetermined maximum as determined by the bias of the spring 169. Such spring bias may be increased or decreased as desired by substituting spring backup disks 170 of different thicknesses between the spring 169 and the aforesaid plug 166.

The direction control valve 17 (FIGS. 2 t0 6) The body of this valve is formed in the housing part 68 including a bore therethrough intersected axially therealong by a pressure inlet passage 143; a pair of service passages 20; a return passage 18; and another pair of service passages .19.

Reciprocable in the bore 175 is a valve spool 176 biased to the right as viewed in FIG. 6 by means of the spring 177 that is compressed between the solenoid assembdy 25 and the valve spool 176. In this position, which is the forward drive position, fluid under pressure delivered through

passages

14A and 14B from the variable capacity, positive displacement pump 10 flows through the service passages 20 to the wheel-mounted fluid motors 15 to drive the wheels 2 to propel the vehicle 1 forwardly. When the solenoid 25 is energized by turning the direction switch lever 28 to the reverse R position, the spool 176 will be pulled to the left, as viewed in FIG. 6, thereby closing communication between the inlet and the

forward passages

14B and 20 and establishing communication between the inlet passages 14B and the reverse passages 19 by way of the chamber 178 at the left, axial cross-over passage 179 from chamber 178 to the chamber 180 at the right-hand end of the valve 17 to passages 19 whereby the wheel motors 15 and wheels 2 will be driven in the opposite direction to propel the vehicle 1 rearwardly.

When the spool 176 is in the forward position (solenoid 25 deenergized) the fluid displaced by the fluid motors returns through the service passages 19 into the return passage 18 to the sump 13 by way of the open brake valve 21 which is shown in FIGS. 2 and 13, and similarly, when the spool 176 is in the reverse position (solenoid 25 energized) the displaced fluid from the wheel motors 15 flows through the service passages 20 to the return passage 18 and thence to the sump 13 by Way of the open brake valve 21.

The ends of the valve bore 175 are closed by

plugs

181 and 182 which are retained in place by snap rings or the like.

The brake valve 21 (FIGS. 2 and 13) This valve as described in connection wth FIG. 2, is

operated by the cam 39 on the cam shaft responsive to movement of the pedal 6 to the Brake Range. When the pedal 6 is in the Free-Wheeling Range and Power Range, as represented in FIG. 2, the pressure balanced

brake valve spool

185, 188 will be biased by spring 186 against cam 39 to assume the position shown in FIG. 13, whereby the return fluid flowing in passage 18 from the wheel motors 15 and from the return passage 18 of the direction control rvalve 17 will flow through the

passages

187 and 187A into the hot oil gallery 71, but as the pedal 6 is swung into the Brake Range (cam shaft 35 rotates counterclockwise as viewed in FIG. 13) the spring 186 will cause the brake valve spool 185 to progressively decrease the cross-section size of the

passages

187 and 187A thereby restricting the return of fluid from the wheel motors 15 and applying a braking action in proportion to the size of the

openings

187 and 187A. When the brake is fully applied, the brake valve spool 185 substantially closes the passages 187 so that full braking elfect is exerted on the wheel motors 15 except for slight leakage through the annular gap between the land 188 and the bore 189 of the ported sleeve 190.

It is to be noted that the heating of the fluid due to the braking action is remote from the wheel motors 15, that is, it is at the throttled openings 187 and 187a in the brake valve 21 and as heat is generated, it is promptly dissipated by the heat exchanger 22 which is closedly adjacent to the hot oil gallery 71.

The pilot operated relief valve 24 (FIGS. 2, 3, 4 and 14) The valve body is formed in the housing part 68 and the valve includes a spring-biased main valve member 195 which closes communication between the pump 10, delivery line 14 (or 14a and 14b) and the return line 13, except when the pressure in the delivery line exceeds a predetermined maximum safe value which, for example, may be in the vicinity of 12,000 psi. In order that such main relief valve member 195 may employ a relatively soft spring 196 and effect a prompt and large opening for such relief of excess pressure, there is provided in tandem therewith a pilot valve assembly 197 which includes a thimble portion 198 having an orifice 199 therethrough and normally closed by the springbiased pilot valve member 2110. The main valve195 also has an orifice 2G1 therethrough and when the pilot valve member 200 is in seated position, the pressures in the

chambers

202 and 203 are equalized through the main valve orifice 201 so that the main valve 195 is held in seated position by the relatively weak spring 196. However, when the pressure in the chamber 203 acting on the small exposed area of the pilot valve 200 exceeds the bias effect of the pilot valve spring 204, the pilot valve 200 is urged to open position to vent the chamber 203 by way of the passage 205 at a rate which is greater than that at which fluid can be replenished through orifice 201 into the chamber 2133 from the main valve chamber 202. When this occurs there will be a pressure differ- 1Q ential between the

chambers

202 and 203 such that the biasing effect of the main valve spring 1% is overcome, whereby the main valve is urged by the predominating pressure in the chamber 2112 toward the right, as viewed in FIG. 14, to relieve such excess pressure building up in the delivery system 14A and 1413 from the variable capacity pump 11) through

passages

206 and 205 into the hot oil gallery 7 1.

The free-wheeling check valve 62 (FIGS. 2 20 6) As best shown in FIGS. 5 and 6, this valve has an inlet port 210 communicating with sump 13, a seat 211, a check valve member 212 urged by spring 213 to closed position (and also by fluid pressure in the pressure inlet passage 14B of the direction control value), and an outlet port 214 leading to passage 14B. Thus, whenever the wheels 2 are turning faster than the fluid from pump 10 flows into the motors 15, the deficiency is made up by opening of the valve member 212 under the influence of negative pressure in

passages

14B and 214 which permits the then predominating pressure of the fluid in the sump 13 to force the valve member 212 away from seat 211. This occurs also when the pedal 6 is moved to the Free-Wheeling Range while the vehicle 1 is in motion.

Throttle operation (FIGS. 1 and 2) As shown in FIG. 2, the engine throttle valve 33 is linked to a cam member or lever 37 which has a cam groove of configuration as shown to open the throttle in the manner represented by the curve 215 in FIG. 25, which, in conjunction with the other controls effects a pressure increase in accordance with the curve 124 and an rpm. increase in accordance with the curve 216. Also, the braking curve 213 shown in FIG. 25 is preferably a straight line function with percentage of braking power correlated with the degree of movement of the pedal 6. Of course, as already mentioned in connection with FIG. 2, the cam link 37 is actuated by the cam 36 on the cam shaft 35 in accordance with the movement of the operating pedal 6.

II. The wheels 2 and wheel motors 15 (FIGS. 16, 17, and 21 I024!) As already described in connection with FIG. 2, the wheels 2 are rotatable on the generally horizontally ex tending portions 31 of the mounting struts 32 and in the case of the front wheels, the struts are: provided with steering levers 33. Referring now in detail to the Wheel and wheel motor construction, there is provided a wheel rim 220 which may be of conventional form, that is, it may be of the drop-center type including side flanges 221 and base flanges 222 constituting supports for the beads 2235 of a pneumatic tire 224, such tire herein being shown as a tubeless tire. Said rim 220 also has a drop-center well 225 operative in well-known manner to facilitate mounting and demounting of the tire 224. Welded, or otherwise secured, to the wheel rim 22% is a mounting ring 226 formed with a plurality of holes through which extend the wheel mounting studs 227, said studs accommodating the nuts 228 which are operative to mount the wheel rim 220) in place on the body 229 of the wheel 2. The, body 229 of the wheel, in this case, is the body of a hydraulic motor 15 which has a plurality of radial plungers 234 therein which are biased inwardly by springs 231 against the periphery of the eccentric 232, the latter being supported on needle bearings 233 on the fixed shaft 31 through which fluid under pressure enters

passage

20A or 19A on one side or the other of the partition 2% and through which fluid returns back to the sump 13 as aforesaid through the

passage

19A or 20A on one side or the other of the partition 234. When pressure enters through passage 211A, the wheel 2 will be driven by motor 15 in forward direction and when pressure enters through passage 19A the wheel will be driven by motor 15 in reverse direction.)

' pump 10 inlet in that it is pressure-balanced.

sear/n The valving arrangement is similar to that used at The body 229 is formed with radiating passages 239 that register with passages 240 formed in the ported sleeve 241 that is press-fitted or keyed in the body 229, said passages 239 leading to the respective plunger chambers 242 to force the plungers inward against the eccentric 232 when the passages are communicated with a fluid pressure source and to displace the fluid from the chambers 24-2 and through

passages

239 and 2% to sump 13 when the plungers are forced outwardly by the eccentric 232. To so communicate

successive passages

239 and 240 with the pressure source and with the sump, the fixed shaft 31 is formed with-

passages

243 and 244 leading to passages 19A and 212A therein.

As best shown in the developed views, FIGS. 23 and 24, the shaft 31 has two sets of slots 245 on either side of the passages 243 and 244- which are connected by passages 246 to

opposite passages

243 and 244, and the sleeve has two sets of openings 24-7 of aggregate area of the respective passages 2 50. Therefore, the sleeve 241 will turn freely on the shaft 31 irrespective of the magnitude of the fluid pressure.

The body 229 and closure 249 assembly is rotatably supported on shaft 31 by needle bearings 250 and by ball bearing 251. The bearings and idle motor chambers are vented through passage 252 which is adapted for connection with sump 13.

When the wheel 2 is to be driven in the forward direction, fiuid under pressure enters certain ones of the

passages

240 and 239 leading to the respective plunger chambers 242 to force the plungers 23h inward when they are at a position on one side of dead center of the eccentric 232. When the wheel 2 is to be driven in reverse direction the fluid under pressure enters through the certain ones of the passages 24th and 239 to act on the plungers 23%) that bear on the eccentric 232 on the other side of dead center to force the plungers 2350 inwardly and by reaction with the eccentric 232 cause reverse rotation of the wheel.

As aforesaid, when it is desired to brake the wheels 2, the return passage 18 (FIG. 2) is blocked by the brake valve 21 and the extent of blocking or restriction of the return flow determines the braking effect to gradually slow down or quickly stop the vehicle 1.

III. Alternative structures-Positioning of the valve 54 by manifold vacuum instead of by the operators foot pedal 6 (FIGS. 25 to 28) As shown in FIG. 25, the structure is basically the same as FIGS. 2-5, with the exception that the valve 269 is positioned by manifold vacuum applied in the chamber 261 against a spring-biased diaphragm 262 and that the throttle valve 263 is positioned by direct mechanical linkage 264 with the operators foot pedal through the pedal operated member 255, there being no need for the cam 46 shown in FIGS. 2 and 12. The orifice 266, and

chambers

267 and 268 correspond respectively with orifice 142 and

chambers

55 and 55 shown in FIGS. 2 and 12.

In FIG. 25, the brake valve 269 is operated by member 255 which is linked to the operators pedal in such manner that when member 265 moves to the left, the fluid from the direction control valve 17 connected with port 2749 is constricted as it flows past valve 269 to the sump port 271.

The curve 275 in FIG. 26 shows the desired relationship between engine r.p.m. and the servo pressure in the chambers at the ends of the valve 47. The curve 276 in FIG. 27 shows the relationship between engine rpm. and manifold vacuum. The curve 277 in FIG. 28, is a crossplot of the curves 275 and 27a, and shows the relationship desired for manifold vacuum versus control pressure in the chamber 268 (chamber 56 in FIG. 2). This relationship is obtained through proper selection of the following variables as indicated on FIG. 25, viz., the

area of the diaphragm 262, the rate and load of the spring 278, the size of the orifice 266, and the contour of the valve 260 and the size of the valve seat.

Referring to FIG. 27, the point 280 corresponds to zero throttle valve opening, the point 281 corresponds to approximately 50% throttle valve opening, the point 282 corresponds approximately to throttle valve opening and the point 283 corresponds to throttle valve opening, the curve 284 illustrating the relationship of manifold vacuum versus engine r.p.m. for a constant throttle opening of approximately 50%. Curves similar to curve 284 may be plotted for other throttle valve openings. Assuming steady operation at point 281, the 50% throttle valve opening to be at 50 mph. on level road under this condition, the pressures in the chambers of valve 47 are equal and the valve therein is centered. Assuming that the vehicle starts up a hill, but that the operator does not change the position of the pedal 6, as the vehicle starts up the hill the extra load imposed thereby will cause the engine r.p.m. and manifold vacuum to decrease along the constant throttle line together with a reduction in vehicle speed. The decrease in r.p.m. reduces the pressure in the chamber 267 of valve 47 in accordance with the curve 275 while the reduction in manifold vacuum increases the pressure in the chamber 268 in accordance with the curve 277. These pressure changes cause the valve 47 to move to the right in FIG. 25 and thereby reduce the capacity and power of the engine driven pump 10. This allows the engine to increase r.p.m. and manifold vacuum along the curve 284 until the pressures are again equalized at the point 281 on curve 276. At this time, the vehicle speed will be less than 50 mph. by an amount proportional to the grade of the hill.

On the other hand, if the operator wants to maintain a speed of 50 mph. while climbing the hill, all that he would have had to do was to depress the pedal 6 and thereby open the throttle valve 263 a suficient amount to obtain the extra power required. By way of example, it is assumed that the throttle valve opening giving the manifold vacuum at point 282 would have been correct and that the throttle 263 is now opened to this point. Concurrently, with the opening of the throttle from point 281 to point 282 the manifold vacuum decreases to the value shown at point 282. The lower vacuum permits the spring 278 to urge the diaphragm 262 to the right and thereby results in the valve 260 moving in a direction to decrease the flow through its seat thereby raising the pressure in the chamber 268 in accordance with the curve 277. Thus, momentarily the pressure in the chamber 268 is greater than in the other chamber 55 and the valve 47 is moved to the right and reduces the capacity and power of the engine driven pump 14 This allows the engine rpm. to increase thereby concurrently increasing the pressure in the chamber 55 and when the engine r.p.m. has increased to the value corresponding to point 282 the pressure in chamber 55 will have increased to an amount which equals the pressure in the chamber 268 and the valve 47 again will be centered. Under this new condition of pressure balance, the power delivered by the engine will be greater by an amount proportional to the decrease in manifold vacuum and the increases in r.p.m. This increase in power is that which was required to maintain the vehicle speed at 50 mph. while climbing the hill.

Power for acceleration is obtained in a similar manner and it can be seen that by flooring the foot pedal 6, the engine power associated with full throttle and rated r.p.m. is instantly available for acceleration at any vehicle speed thereby providing more power for acceleration than it attainable with any known transmission system in use.

Other modes of applying the principle of the invention may be employed, change being made as regards the details described, provided the features stated in any of 13 the following claims or the equivalent of such be employed.

I, therefore, particularly point out and distinctly claim as my invention:

1. A variable capacity pump comprising a housing; radial plungers reciprocable in said housing; a drive shaft journalled in said housing; an eccentric on said drive shaft engaging said plungers to move them; said housing being formed with an intake port adapted to be connected to a sump and a delivery port; rotary intake valving driven by said drive shaft successively to communicate said plungers with said intake port when said plungers move in one direction; an outlet valve in said housing successively to communicate said plungers with said delivery port when said plungers move in an opposite direction; a spill passage adapted to be connectedto such sump, and fluid pressure actuated capacity varying means in said housing effective to control fluid flow through said spill passage to such sump during predetermined portions of the strokes of said plungers according to the magnitude of the fluid pressure acting on said capacity varying means.

2. The pump of claim 1 wherein said rotary intake valving is fluid-pressure balanced.

3. The pump of claim 1 wherein said capacity varying means comprises a sleeve-type valve for each plunger which is spring-biased to zero capacity position.

4. In a power transmission system for a prime mover having an associated control means to vary the speed and power output of the prime mover; a variable capacity pump comprising a housing; radial plungers reciprocable in said housing; a drive shaft journalled in said housing adapted to be driven by the prime mover; an eccentric on said drive shaft engaging said plungers to move them; said housing being formed with an intake port adapted to be connected to a sump, and a delivery port; rotary intake valving driven by said drive shaft successively to communicate said plungers with said intake port when said plungers move in one direction; an outlet valve in said housing successively to communicate said plungers with said delivery port when said plungers move in an opposite direction; a spill passage adapted to be connected to such sump, and fluid pressure actuated capacity varying means in said housing effective to control fluid flow through said spill passage to such sump, during predetermined portions of the strokes of said plungers according to the magnitude of the fluid pressure acting on said capacity varying means.

5. The power transmission system of claim 4 wherein said capacity varying means comprises a sleeve-type valve for each plunger which is spring-biased to zero capacity position.

6. The power transmission system of claim 5 wherein said prime mover comprises an internal combustion engine including an intake manifold, and the position of said sleeve-type valve is responsive to the manifold vacuum of said prime mover.

7. The power transmission system of claim 4 wherein said rotary intake valving is fluid-pressure balanced.

8. A variable capacity pump comprising a housing, radially extending plungers reciprocable in said housing, a drive shaft journalled in said housing, an eccentric on said drive shaft operative to reciprocate said plungers, said housing being formed with an intake port adapted to be connected to a sump, and a delivery port, rotary intake valving driven by said drive shaft successively to communicate said plungers with said intake port when said plungers move in one direction; a spill passage adapted to be connected to such sump, and fluid pressure actuated capacity varying means in said housing operative to control fluid flow through said spill passage to such sump for a predetermined portion of the strokes of said plungers according to the magnitude of fluid pressure acting on said capacity varying means.

9. A pump as set forth in claim 8 wherein said capacity varying means comprises a movable sleeve surrounding each plunger, a spill groove in each plunger communicating with said spill passage for a portion of the stroke of said plunger depending on the position of said sleeve.

10. A pump as set forth in claim 9 including fluid pressure means acting on said sleeve tending to move said sleeve to shorten the portion of the stroke of said plunger wherein said spill groove will be exposed to said spill passage to increase the capacity of said pump.

11. A pump as set forth in claim 10 including spring means acting on said sleeve opposing said fluid pressure means tending to move said sleeve to increase the portion of the stroke of said plunger wherein said spill groove will be exposed to said spill passage to decrease the capacity of the pump.

12. The pump of claim 8 wherein said rotary intake valving comprises a sleeve, multiple passages in said sleeve operative to communicate said plungers with said intake port when said plungers move in one direction, pairs of static 'holes in said sleeve diametrically opposite each passage, and groove means in said sleeve connecting each passage and the pair of holes diametrically opposite thereto to balance out the fluid. forces on said sleeve.

13. A variable capacity pump comprising a housing, radially extending plungers reciprocable in said housing, a drive shaft journalled in said housing, means operative to reciprocate said plungers, said housing being formed with an intake port adapted to be connected to.a sump, and a delivery port, a spill passage, capacity varying means in said housing actuated by fluid pressure imposed directly thereon operative to control fluid flow through said spill passage to such pump for a predetermined portion of the strokes of said plungers according to the magnitude of fluid pressure acting directly on said capacity varymg means.

References Cited by the Examiner UNITED STATES PATENTS 1,440,428 1/ 1923 Wigelius 103-37 1,508,054 9/ 1924 Hoppins 103-37 1,642,103 9/ 1927 Daubenmeyer -66 1,657,841 1/1928 Peris 251-283 2,251,783 8/1941 Davis 103-41.1 2,319,566 5/ 1943 Sunderman 103--41.1 2,372,523 3/ 1945 Sinclair 103-174 2,418,123 4/1947 Joy 180-66 2,433,222 12/1947 Huber 103-173 2,494,505 1/1950 Bouchrad 180-54.3 X 2,524,235 10/1950 Schenk 103-41 2,535,617 12/ 1950 Westbrook 103-12 2,545,220 3/1951 Wolcott 251-283 2,640,372 6/1953 Dodge 74-472.1 2,650,573 9/1953 Hickman 60-53 X 2,786,424 3/ 1957 Raymond 103-174 2,875,635 3/ 1959 Fleck et a1. 74-472.] 2,889,780 6/1959 Binford 103-12 2,945,451 7/ 1960 Griswold 103-174 3,059,416 10/1962 Campbell 180-60 X FOREIGN PATENTS 893,213 1/1944 France.

578,985 6/1933 Germany.

459,578 1/ 1937 Great Britain.

LAURENCE V. EFNER, Primary Examiner. G. HARRY LEVY, Examiner.

M. L. SMITH, Assistant Examiner.

Claims

1. A VARIABLE CAPACITY PUMP COMPRISING A HOUSING; RADIAL PLUNGERS RECIPROCABLE IN SAID HOUSING; A DRIVE SHAFT JOURNALLED IN SAID HOUSING; AN ECCENTRIC ON SAID DRIVE SHAFT ENGAGING SAID PLUNGERS TO MOVE THEM; SAID HOUSING BEING FORMED WITH AN INTAKE PORT ADAPTED TO BE CONNECTED TO A SUMP AND A DELIVERY PORT; ROTARY INTAKE VALVING DRIVEN BY SAID DRIVE SHAFT SUCCESSIVELY TO COMMUNICATE SAID PLUNGERS WITH SAID INTAKE PORT WHEN SAID PLUNGERS MOVE IN ONE DIRECTION; AN OUTLET VALVE IN SAID HOUSING SUCCESSIVELY TO COMMUNICATE SAID PLUNGERS WITH SAID DELIVERY PORT WHEN SAID PLUNGERS MOVE IN AN OPPOSITE DIRECTION; A SPILL PASSAGE ADAPTED TO BE CONNECTED TO SUCH SUMP, AND FLUID PRESSURE ACTUATED CAPACITY VARYING MEANS IN SAID HOUSING EFFECTIVE TO CONTROL FLUID FLOW THROUGH SAID SPILL PASSAGE TO SUCH SUMP DURING PREDETERMINED PORTIONS OF THE STROKES OF SAID PLUNGERS ACCORDING TO THE MAGNITUDE OF THE FLUID PRESSURE ACTING ON SAID CAPACITY VARYING MEANS.