Torque limiting means for variable displacement pumps

Abstract

The loading of rotor support journals, of variable displacement pumps, which is proportional to the theoretical driving torque of the pump, is converted to a fluid pressure which is proportional to the theoretical pump driving torque. The fluid pressure is applied to a plunger in opposition to an adjustable load, to operate a servo valve to regulate the pump displacement. The loading may be varied to provide variations in the pump driving torques to correspond to the torques available from a prime mover under a variety of power settings.

Description

FIELD AND BACKGROUND OF THE INVENTION

This invention relates to hydraulic pumps and motors of the kind comprising a body, and a rotor surrounded by and mounted in an external support, the body having an inlet and outlet for liquid to be pumped and the rotor carrying a plurality of equiangularly spaced pistons which are disposed parallel to or inclined to the rotor rotational axis and are reciprocated as the rotor rotates in order to pump liquid. Such a pump will be referred to hereinafter as being of the variable displacement piston pump type.

One form of piston pump to which the invention may be applied comprises a rotary cylinder block assembly in which the pistons are equally spaced around the cylinder block axis with the cylinder bores parallel to the axis. The cylinder block is linked to the driving mechanism in such a way that the driving mechanism and cylinder block rotate at the same speed. As the shaft rotates, the distance between any one piston and the porting surface changes continually. Each piston moves away from the porting surface during one half of the revolution and toward the porting surface during the other half, the inlet port being in line as the piston moves away, and the outlet port being in line as the piston moves closer.

An object of the invention is to provide a variable displacement piston pump, with torque limiting means in a convenient form, as to ensure a simplified and cheaper production of the torque means, which are easily adjusted on site, or controlled to suit any available torque from a power supply unit.

The invention consists in torque limiting means for limiting the driving torque of a variable displacement piston pump comprising a hydrostatically balanced slipper bearing, supporting the pump rotor, and pressure responsive means, responsive to the hydrostatic bearing pressure, to regulate the pump's displacement.

When variable displacement piston pumps are employed for power transmission purposes, it is usually necessary to provide a manual control to regulate the maximum pump displacement and to provide overriding devices, one to reduce the pump displacement if the transmission pressure exceeds a predetermined value, in order to protect the transmission components against overload, and another to reduce the pump displacement if the driving torque exceeds a predetermined value, in order to prevent overloading of the engine or electric motor which drives the pump.

The theoretical pump driving torque is the product of the pump displacement and the pressure difference, and the torque limiting means are based on this fact. Hitherto the torque limiting means have including a plunger subject to the pump delivery pressure to obtain a servo valve to regulate the pump displacement through a servo motor, the plunger movements being opposed by a spring loading which is varied by a cam responding to the movements of the servo motor, so that the spring loading is inversely proportional to the pump displacement. Such torque limiting means are complex and costly to produce and they can be adjusted correctly only on special test rigs incorporating torque dynamometers. Such torque limiters cannot be adjusted in a proportional manner to suit varying available torques, such as the varying torques of petrol engines operating over a range of throttle openings.

SUMMARY OF THE INVENTION

The invention is based on an appreciation that the rotor support journal loading of variable displacement pumps is the product of the pump displacement and the operating pressure, so the journal loading is proportional to the theoretical driving torque. According to the invention, the torque responsive journal loading is converted to a fluid pressure, so that the fluid pressure is proportional to the theoretical pump driving torque. This pressure is applied to a plunger in opposition to an adjustable load to operate a servo valve to regulate the pump displacement. The invention includes novel means for varying the loading, to provide variations in the pump driving torques to suit the torques available from a prime mover under a variety of power settings.

The invention is described as applied to a particular form of axial piston pump, which is controlled overcenter for forward and reverse drive, but it can equally well be applied to other forms of variable displacement pumps.

The invention is described with reference to a particular type of hydro-statically balanced slipper bearing, but it can equally well be applied to operate from any design of hydrostatically balanced bearing, where the hydro-static pressure supports the torque responsive loading. In order to illustrate the function of the invention in relation to the other controls, a suitable complete control system is described, but the invention is not limited to the type and details of the control system or of servo motor employed.

BRIEF DESCRIPTION OF THE DRAWINGS

An embodiment of the invention is described with reference to the accompanying drawings.

FIG. 1 is a longitudinal section through a typical axial piston pump of a kind to which the invention may be applied.

FIG. 2 is an end view of the stationary port face for the pump shown in FIG. 1.

FIG. 3 is a diagrammatic representation of a pump control system according to one embodiment of the invention.

FIG. 4 shows a preferred form of servo control valve.

FIG. 5 is a graphical representation of the performance of an engine speed responsive torque limiter according to the invention.

FIG. 6 shows a method for achieving the control characteristics shown in FIG. 5.

FIGS. 7, 8 and 9 are sectional views taken along the lines 7--7, 8--8 and 9--9 respectively of FIG. 4.

Referring to FIG. 1, the pump cylinder barrel 1 contains nine axial cylinders with pistons 2 having ball and socket joints 3 to flat faced slippers, bearing against a variable angle oblique surface 4 and held in contact with the surface by the slipper retaining plate 5.

The cylinder rotor is loaded against its port face 6 by a number of springs 7, reacting through the drive shaft 8 to the thrust bearing 9. A suitable prime mover, such as an electric motor or an internal combustion engine, is arranged to be coupled to drive shaft 8.

The oblique surface 4 is on a part cylindrical element 10, bearing in the housing end cover 11 and the angle of the oblique surface is adjustable by movements of the servo motor piston 12 through a gear tooth rack 13 engaging teeth formed on the part cylindrical element 10. Movement of the servo motor piston is capable of rotating the cylindrical element 10 to an oblique angle of say 20.degree. in either direction, thereby varying the strokes of the pistons and the direction of flow through the portings.

When the element 10 is in the position shown and the pump is driven clockwise, as viewed from the end of the drive shaft 8, the delivery will be from port A shown in FIG. 2, which will be referred to as the forward high pressure port. When the element 10 is tilted in the opposite direction, the delivery will be through port B, which will be referred to as the reverse high pressure port.

The drive shaft 8 is provided with a high degree of flexibility, so that the cylinder barrel 1 is free to align itself against the flat port face surface 6 and the cylinder barrel is located by four part spherical slipper pads. The journal loading at the cylinder barrel is applied at the slipper 14, when the oblique surface 4 is tilted in the direction shown. When the oblique surface is tilted in the opposite direction, the journal loading would be at the slipper 15. There are no theoretical sideways loadings at the cylinder barrel, but a slipper 16 as shown in the partial section of FIG. 2, is placed at each side for location purposes only.

The slippers 14 and 15 are hydro-statically balanced against the journal loadings. They may for example be connected through the connection points marked C and D to the highest porting pressure through restricted orifices. The sealed area at the spherical faces of the slippers is sufficiently great for the requirements at maximum pump displacement. The required hydro-static balance pressure becomes progressively less than the porting pressure as the pump displacement is reduced. Under these conditions the slipper operates with a small clearance at its spherical surface, the resulting leakage providing an appropriate pressure drop at the restricted orifice.

The inside of the pump housing is maintained at the same pressure as that in the low pressure port, so the journal loading is proportional to the product of the porting pressure difference and the tangent of the angle of the oblique surface. The pump driving torque is also theoretically proportional to the product of these two variables, so the hydro-static pressure at the slipper is proportional to the theoretical driving torque. According to the invention, this hydro-static pressure at the loaded slipper bearing is employed as a signal for the torque limiting device.

The two ends of the servo cylinder are shown with connection points marked E and F and the various connection points identified by the letters A,B,C,D,E, and F in FIGS. 1 and 2 are similarly identified in the control system diagram FIG. 3.

Referring to FIG. 3, a low pressure pump 17, draws from the reservoir 18 and its delivery pressure is limited by the setting of the low pressure relief valve 19. It delivers directly to the inside of the variable displacement piston pump housing, from which it flows to whichever is the low pressure port through one of the non return valves 20 or 21. Under clockwise forward conditions port A is at high pressure and port B is charged from the low pressure pump through the non return valve 20.

Ducts from the pump ports A and B have non return valves 22 and 23, the downstream sides of which are inter-connected, supplying the selected highest pressure through the restrictors 24 and 25 to the hydro-statically balanced slippers 14 and 15, respectively, through the connection points C and D as described with reference to FIG. 1.

The selected highest pressure from the non return valves 22 and 23 is supplied to the center gallery of a manually operated reversing valve 26, which also contains a small gallery and a central drilling to one end, which is ducted to the reservoir 18.

Two galleries 28 and 29 formed in the housing 27 of the manually operated reversing valve 26 are ducted through a servo valve 30 to the cylinders of the transmission pump servo motor 12 via the connection points E and F, respectively. The manually operated reversing valve 26 is shown in the position to supply the highest pressure to the servo motor connection point E, with the connection point F ducted through the gallery 29 to the reservoir. This causes the servo piston to move to the position of maximum pump displacement as shown at FIG. 1. Movement of the reversing valve 26 outwards reverses the pressures at the connection points E and F and causes the servo piston 12 to move to the other extreme position.

The servo valve 30 contains a spool valve 31 having an enlarged piston 32 at one end, sliding in a cylinder 33 and loaded by a spring 34 in the opposite direction to the position shown. The spring chamber and the remote end of the spool valve 31 are ducted to the reservoir 18. The delivery port of the low pressure pump 17 is ducted to the cylinder 33 through the restrictor 35 to move the spool valve 31 against the loading of the spring 34 to the position shown, where it does not influence the operation of the servo motor in response to the movements of the manually operated reversing valve 27.

The galleries 36 and 37 of the servo valve 30 are ducted to the transmission pump ports A and B respectively. If the cylinder 33 is exhausted, the spool valve 31 will move in the opposite direction to that shown, to isolate the galleries ducted to the manually operated reversing valve 27 and to provide communication from the pump ports A and B to the connection points F and E, respectively, of the servo motor cylinder.

Port A is the forward drive high pressure port, so movement of the spool valve 31 causes the servo motor piston 12 to reduce the displacement of the pump under forward drive conditions, but it cannot cause movements beyond zero pump displacement, because this would cause port B to become the high pressure port. The operation is similar when the manually operated valve 26 is moved to the reverse position.

Under overrun conditions, the pump operates as a motor and, for forward setting of the reversing valve, port B becomes the high pressure port. Movement of the spool valve 31 then causes the servo piston to increase the pump displacement, as desired under overrun conditions.

For counter clockwise rotation of the pump, it is necessary only to reverse the connections from ports A and B to the galleries 36 and 37.

In most applications of variable displacement transmission pumps, it is necessary to provide means to stall the pump off stroke at a predetermined maximum pressure to prevent overloading of the pump, and also at a predetermined maximum pump driving torque, to prevent overloading of the power plant driving the pump.

A pilot valve device 38 contains two small plungers 39 and 40 bearing against valves 41 and 42 which are loaded against their seatings by adjustable springs 43 and 44 to close galleries which are ducted to the cylinder 33 of the servo valve. The pilot valve discharge chambers are ducted to the reservoir.

A duct supplies whichever is the higher pressure from ports A and B to the end of the plunger 39, so that if this pressure is excessive the valve 41 opens to vent the cylinder 33, causing the pump to stall off stroke.

The pressures at the slippers 14 and 15 are ducted through non return valves 45 and 46 to the end of the plunger 40, so this plunger is responsive to the hydro-static pressure at whichever slipper is supporting the journal loading. The force at the plunger 40 is therefore responsive to the pump driving torque. If this pressure is excessive, the valve opens to vent the cylinder 33, in the servo valve 42, so limiting the pumps driving torque by variation of the pump's displacement.

Pump servo systems operated by the high pressure from the pump delivery ports are prone to instability problems, particularly under high gain conditions, i.e. at small pump displacement. The invention includes a stabilising device which is a function of the rate of change of pump displacement. A typical embodiment of this part of the invention in shown in FIG. 3.

A piston 47 is shown with a cam roller 48, bearing against a cam 49 formed on the servo motor piston. For preference the cam form is such that the movements of the piston 47, relative to the servo motor 12, is an inverse function of the pump displacement. The outer end of the cylinder for piston 47 communicates through a central drilling to the inside of the pump housing and is therefore subject to the pressure from the low pressure pump 17 to hold the cam roller in contact with the cam. The annular cylinder chamber communicates with the spring chamber of the pilot valve 38 and is ducted to the reservoir through the restricted orifice 51. The pressure difference across the restricted orifice 51, resulting from movements of the piston 47, acts on the sealed pistons 52 and 53 to vary the loading at the valves 41 and 42.

When one of the valves 41 or 42 opens to limit the pump pressure or the pump's driving torque, the pressure difference at orifice 51 provides a force to close the valve. A reduction in pressure at the plungers 39 or 40, allowing the associated valve to close, results in a pressure difference at orifice 51, providing a force to open the valve. Thus the stabilizing device provides a leading term, to stabilise the automatic control system, from the rate of change of the pump displacement.

In many applications of variable displacement pumps, it is desired to vary the maximum pump displacement by a manual control, with maximum pressure and maximum torque overriding devices. FIG. 4 shows a valve to replace valves 27 and 30 of FIG. 3 and which provides for manual displacement control.

A manually operated control lever 54 rotates a ported sleeve 55 in a housing 56. A rotary valve element 57 inside the ported sleeve has a gear tooth quadrant 58 driven by the gear 59 attached to the element 10 of FIG. 1, or by rack teeth formed on the servo motor piston 12. In either case, the gear tooth chamber communicates with the inside of the pump housing and is subject to the pressure from the low pressure pump 17. This pressure acts on an annular area of the gear driven spool 57 and it is applied to a cylinder 60 on the opposite side, through a restricted orifice 61. Normally the piston in the cylinder 60 is moved to the position shown by the pressure acting on the difference in areas, but if the cylinder 60 is vented, the gear driven rotary spool moves to the left as viewed on the drawing. This corresponds to the cylinder 33 of FIG. 3, the venting valves may be similar to those described for the pilot valve 38 of FIG. 3, and the connection K is identified in FIG. 3.

Directional control selector rotary valves of this type are well known. As shown in FIGS. 4, 7, 8 and 9 they usually have galleries for each different pressure with portings through the sleeve in diametrically opposite pairs in order to avoid side loadings on the valves.

The ports marked A and B are ducted to the transmission pump ports similarly identified. The ports marked E and F are ducted to the ends of the servo motor cylinder similarly identified. The port marked G is ducted to the selected highest pressure identified similarly in FIG. 3. The port marked H is connected to the reservoir, 18, which is permanently ported to the end of the cylinder containing the valve 57 and, through a central drilling, to an annular chamber having an outer diameter equal to the outside diameter of the sleeve 55.

In its normal operating position as shown, rotation of the valve sleeve provides communication of the servo cylinder ports E and F with the selected highest pressure port G or the reservoir port H. The resulting movement of the servo motor piston 12 closes the ports G and H, so that normally the valve operates as a position responsive control. Under these conditions the ports A and B are closed.

If either of the overriding valves 41 or 42 in FIG. 3 is opened it vents the cylinder 60 causing the spool valve 57 to slide axially to the left. This closes the ports H and G from communication with the valve flutes and opens the ports A and B. Port A then communicates with one servo cylinder of the servo motor through port F, and port B communicates with the other servo cylinder through port E. Thus this valve operates in the same manner as the servo valve 30 described for FIG. 3.

The spring which loads the valve 42 for torque limitation is shown with an adjusting screw. Obviously this screw may be replaced by a push rod operated from a camshaft coupled to the controls of a prime mover, so that the pump driving torque would be a function of the setting of the prime mover power controls.

A preferred method of controlling the pump driving torque as a function of the prime mover power control settings is to cause the controlling torque to increase as a controllable function of prime mover speed until the torque required line intersects the available torque line from the prime mover. This is exampled by the curves in FIG. 5. The curves a1, to a5 represent prime mover torque outputs for given throttle openings or given fuel pump discharges for diesel engines. The curves b define a controlled torque area increasing with the square of engine speed, with a negative constant, so that there is zero torque loading for engine speeds below 15 percent and the torque loading increases to 100 at 40 percent of maximum engine speed. The actual controlled engine conditions are at the intersection of the a curves and the b curves.

The curve b.sub.1 represents a minimum control setting and the curve b.sub.2 represents a maximum control setting for the torque control, and any other curve may be provided between these limits, for different prime mover power requirements.

A method of achieving these desired control characteristics is shown in FIG. 6 which illustrates the torque limiting device, the low pressure pump 17, the reservoir 18 and the low pressure relief valve 19.

The lettered connection points to the torque limiter are J, the torque responsive hydro-static pressure from the slipper bearing, K, the pilot pressure identified at the servo valve device 30 in FIG. 3, or the device illustrated in FIG. 4, and L, stabilising the pressure from the cylinder 50 in FIG. 3.

A variable orifice 62 is fitted in the discharge passage from the low pressure pump 17. The pressure drop at this orifice is applied to the piston 63, against the loading of the spring 64, to close the valve. The spring 64 provides the desired negative constant. The pump 17 is driven by the prime mover, so the pressure at J, which is proportional to the pump driving torque, will be K.sub.1 N.sup.2 -K.sub.2, where N is the engine speed, K.sub.2 is the negative constant and K.sub.1 is varied with the setting of the prime mover power controls, to provide controlled torque lines such as b.sub.1 and b.sub.2 in FIG. 5. The arrangement shown in FIG. 6 can replace the equivalent arrangement 38 shown in FIG. 3.

Claims

The claims defining the invention are as follows:

1. In a variable displacement piston pump including a driven rotor, torque limiting means, for limiting the driving torque of the pump, comprising, in combination, a hydrostatically balanced slipper bearing supporting the pump rotor, pressure responsive means operable to regulate a pump displacement, and means transmitting the hydrostatic bearing pressure to said pressure responsive means to effect operation of said pressure responsive means.

2. Torque limiting means according to claim 1, wherein said pressure responsive means comprise a pilot valve device including a moving element responsive to the bearing pressure in opposition to a resilient loading, a servo motor controlling the pump discharge and a servo control valve operated in consequence of the operation of said pilot valve and controlling said servo motor.

3. Torque limiting means according to claim 2, in which the pump is driven by a prime mover, and means controlling the torque as a function of the speed of the prime mover including means applying a fixed negative load to said moving element in combination with means applying, to said movable element, a load varying as a function of the prime mover speed and power control setting.

4. Torque limiting means according to claim 2, including a stabilising device comprising a piston in a cylinder device connected to the servo motor to produce a fluid pressure varying as a function of the rate of change of the pump displacement and acting on the pilot valve loading to enhance the stability of the control system.

5. Torque limiting means according to claim 2, wherein operation of said pilot valve overrides the effect of the other control means acting on the servo valve so as to limit the pump driving torque by variation of the pump displacement irrespective of the setting of said other control means.

6. Torque limiting means according to claim 5, wherein said servo valve is a direction control selector valve, having an axially moveable spool such that, at one limit of the spool axial travel, the servo valve controls the servo motor in response to the setting of said other control means whereas, at the other limit of the spool axial travel, the servo valve causes the servo motor to reduce pump capacity; the spool being moved to said other limit by a pressure change in consequence of operation of said pilot valve.

7. Torque limiting means according to claim 6, wherein said servo valve is a combined manual and automatic displacement and directional control selector valve, comprising a manually operated, ported sleeve and a rotary spool driven by the servo motor and subject to axial displacement in consequence of operation of said pilot valve.

8. Torque limiting means according to claim 5, wherein said pilot valve comprises a bearing pressure responsive plunger, bearing against a springloaded valve element, the springload being mechanically adjustable to regulate the maximum torque value.