GB2087050A

GB2087050A - Hydrostatic Transmission Control System

Info

Publication number: GB2087050A
Application number: GB8133420A
Authority: GB
Original assignee: Danfoss AS
Current assignee: Danfoss AS
Priority date: 1980-11-06
Filing date: 1981-11-05
Publication date: 1982-05-19
Also published as: GB2087050B; FR2493454A1; DE3041856C2; IT8168436A0; DK481981A; IT1145704B; FR2493454B1; JPS57107463A; DE3041856A1; JPS6015818B2

Abstract

Diesel engine (1) drives a vehicle transmission pump (4) and an auxiliary pump (6a) which delivers control fluid to servo-valve (19) via a unit (10) producing an engine speed signal (pst) applied via reversing valve (47) to a servo-valve control piston (34 or 35). The servo-valve controls transmission pump displacement via a servomotor (5) having a feedback link (28), and the speed signal (pst) is opposed by pump delivery pressure (pH1 or pH2) to avoid undue drop in engine speed when the load on the transmission motor (8) increases. A second auxiliary pump (6) supplies lifting or other equipment (48). During overrun, the piston (34 or 35) not subject to the speed signal is lifted clear of the servo-valve to permit hydrostatic braking responsive to the speed signal (and engine accelerator (2)). <IMAGE>

Description

SPECIFICATION Control Means for Hydrostatic Gear This invention relates to control means for hydrostatic gear in which a drive motor drives a main pump which is adjustable by a servoelement and is connected to a hydraulic motor by way of a conduit system, wherein in said control means and with the aid of an auxiliary pump likewise driven by the drive motor and of a control pressure generator a speed-dependent pressure can be produced, the servo-element is actuated by a hydraulic servo-motor to which pressure fluid can be fed by way of a servo-valve, and the servovalve is adjustable by a piston arrangement of which the oppositely acting control pressure forces can be impinged by the control pressure which can be supplied to the main pump by way of a directional switching valve for determining the supply direction.

In a proposed control means of this kind (DE OS 20 08 078), the main and auxiliary pumps are driven by a motor in the form of an internal combustion engine. The auxiliary pump has the triple purpose of feeding the hydraulic servomotor of the servo-element, to produce the control pressure and to replenish the conduit system of the hydrostatic gearing by way of check values.

Such control means have the object of preventing stalling of the drive motor If, the rotary speed of the drive motor drops as a result of excessive loading of the hydraulic motor driven by the main pump, the control pressure is also reduced, with the consequence that the main pump is set to a smaller supply volume. The torque to be applied at this supply volume can then be produced by the drive motor.

However, on heavy loading of the hydraulic motor, it can happen that a new condition of equilibrium is produced only at a very much lower rotary speed of the drive motor. Such intensive motor suppression' is undesirable for various reasons. If the drive motor also drives other pumps, for example for operating lifting or other operational equipment the operating speed is reduced when the drive motor runs more slowly.

In addition, the strong reduction in rotary speed counteracts the aim of preferably operating the drive motor at its optimum speed. If the load on the hydraulic motor exceeds a predetermined limiting value then, despite using the control means, it is not always possible to prevent stalling of the drive motor.

It is an object of the present invention to provide control means of the aforementioned kind which ensure; in a simple manner, that motor suppression, i.e. a reduction in the speed of the drive motor, is less severe upon excessive loading of the hydraulic motor.

That object is obtained according to the invention in that a counter-pressure opposing the currently effective control pressure can be exerted on the piston arrangement that is equal to the pressure obtaining in the conduit system at the supply side of the main pump.

If, in this construction, the hydraulic motor is excessively loaded, this becomes noticeable as a pressure increase in the conduit system on the supply side of the main pump. By reason of the fact that this pressure causes the effect of the control pressure to be reduced, the servo-valve is adjusted and thus the servo-element of the main motor is changed to a reduced supply volume.

One therefore obtains a new condition of equilibrium without the speed of the drive motor dropping too much. Consequently, lifting or operative equipment supplied by another pump driven by the drive motor can be operated at practically full speed even if the hydraulic motor of the hydrostatic gear is loaded up to almost a standstill. In addition, one ensures that the drive motor is always operated near its optimum rotary speed and will not be stalled even under unfavourable conditions. Since the counterpressure is derived from the high pressure in the conduit system one obtains an accurate measure for the loading torque of the hydraulic motor that is not effected by other influences, especiaily those of the main pump. Further, the use of the counter-pressure is a feature which can be implemented at a low constructional expense.

Particularly simple control means are obtained if, apart from the control pressure face, the piston arrangement comprises an oppositely acting counter-pressure face for each supply direction of the main pump and if the two counter-pressure faces are each connected to a respective side of the main pump. In particular, the piston arrangement may comprise for each supply direction a stepped piston of which the larger face is the control pressure face and the smaller face is the counter-pressure face.

With particular advantage, the piston arrangement comprises two parts each with a control pressure face and a counter-pressure face and both parts are adapted non-positively to engage the servo-valve with opposite directional effect. This takes into account difficulties during braking when the hydraulic motor operates as a pump and consequently the pressure drops at the supply side of the main pump and rises at the inlet side. Since the parts of the piston arrangement engage the servo-valve only non-positively, the high pressure obtaining during braking cannot act on the servo-valve. The latter is therefore solely adjusted by the control pressure so that a vehicle equipped with the hydrostatic gear can also be controlled by the usual accelerator pedal during braking.

The construction becomes particularly simple if two stepped pistons are adapted non-positively to engage an axially movable slide valve of the servo-valve from opposite sides.

It is also favourable if the servo-valve comprises two valve elements which are intercoupled by springs and of which one can be influenced by the piston arrangement and the other is connected to the servo-element by a feedback. In this way the forces acting on the piston arrangement are compared with the spring forces in the servo-valve. This leads to very small valve movements in comparison with a distance controlled servo-valve, so that the entire control means can be very compact.

The present invention also provides hydraulic control equipment including a prime mover for driving a main pump and an auxiliary pump, the main pump supplying fluid to a main circuit and the auxiliary pump supplying fluid to a servo control via a valve in a control circuit, the servo control regulating the capacity of the main pump and the valve being controlled in dependence on a pressure signal related to the speed of the prime mover, wherein a pressure signal related to the pressure in the supply side of the main circuit is used to control the valve in opposition to the effect of the first-mentioned signal.

Hydraulic control equipment constructed in accordance with the invention will now be described, by way of example, with reference to the accompanying drawing the single figure of which in a diagrammatic block diagram of the equipment.

Referring to the accompanying drawing, a drive motor 1 in the form of a diesel engine controllable by an accelerator pedal 2 has an output shaft 3 driving a main pump 4 of which the supply volume is variable by means of a servo-element 5, a first auxiliary pump 6 and a second auxiliary pump 6a. The main pump 4 supplies by way of a conduit system 7 a hydraulic motor 8 which can, for example, serve to drive the wheel 9 of a vehicle. The first auxiliary pump 6 serves to operate lifting or operational apparatus 48 connectible to the auxiliary pump 6 by a conventional switching valve 49.The auxiliary pump 6a has several functions:- a) With the aid of a control pressure generator 10 comprising a plate orifice 11 and a transformer unit 12, a control pressure pet is produced in an outlet conduit 13 that is approximately proportional to the quantity delivered by the auxiliary pump and thus to the speed of the drive motor 1.

b) By way of a conduit 14 safeguarded by an overpressure valve 15, pressure fluid is fed into the conduit system 7 to compensate leakage losses occurring therein. Two check valves 16 and 17 ensure that the pressure fluid is fed to the respective low pressure side of the conduit system 7.

c) By way of a conduit 18 and a servo-valve 19, a two-part hydraulic servo-motor 20, 21 is supplied with pressure fluid for adjusting an actuating element 22 of the servo-element 5.

In the present example, each servo-motor part 20, 21 is provided with a piston 23 which is biassed bya spring 24 and has an end stop 25.

On adjustment of the actuating element 22, a shaft 26 is pivoted which, in the main pump 4, brings about an adjustment by which the supply volume is changed. For example, in the case of an axial piston pump, it is the angle of inclination of the rotor plate at the end that is changed and in the case of a radial piston pump it is the eccentric position of the rotor ring that is changed. The actuating element 22 could also be made in one piece with the rotor plate or the rotor ring. On pivotal motion in the direction of the arrows 27, one valve part 29 of the servo-valve 19 is adjusted by way of a feedback 28 relatively to the other valve part 30 against the force of two springs 31 and 32.

The valve element 30 is controlled with the aid of a piston arrangement 33. The latter comprises two stepped pistons 34 and 35 which, by way of a respective piston rod 36 or 36', non-positively lie against the valve element 30. By way of a directional switching valve 47, the larger pressure face 37 of the stepped piston 34 or the larger pressure face 38 of the stepped piston 35 can be selectively supplied with the control pressure p,t.

The smaller pressure face 39 of the stepped piston 34 is connected by way of a conduit 40 to a point 41 of the conduit system 7 and the small pressure face 42 of the stepped piston 35 is connected by way of a conduit 43 to a point 44 of the conduit system.

The arrangement is so designed that, when the control pressure Pat acts on the stepped piston 34, pressure fluid is fed at the output 45 of the main pump 4. Depending on the load on the hydraulic motor 8, the high pressure PHI then obtains on-this supply side. If, however, the stepped piston 35 is subjected to the control pressure p,t, pressure fluid is discharged through the output 46 of the main pump 4. Depending on the load on the hydraulic motor 8, the high pressure PH2 then obtains on this supply side.

Consequently, it is possible to exert on the piston arrangement 33 a counter-pressure PHI or PH2 which is opposed to the respectively operative control pressure pit and which is equal to the pressure obtaining in the conduit system at the supply side of the main pump 4.

The result of this is that the effect of the control pressure P.t on the servo-valve 19 is reduced all the more rapidly as the load on the hydraulic motor 8 increases. Such a strong load therefore directly leads to.an adjustment of the servo-element 5 in the sense of producing a lower supply volume for the main pump 4. Accordingly, the speed of the drive motor 1 need not drop as much as hitherto until a new state of equilibrium is reached. The auxiliary pump 6 therefore also retains its normal speed so that, when the operational equipment 48 is turned on, this can be operated at full speed.

If the hydraulic motor 8 is turned faster than the supply by the main pump 4 demands, for example when driving downhill, a braking effect is obtained and there is a change of high pressure from the supply side of the main pump to the opposite connection. If, for example, the control pressure Pit acts on the stepped piston 34 and high pressure PHI therefore acts on the supply side 45 during normal operation, the high pressure during braking operation changes to the opposite side 46 so that the stepped piston 34 is no longer loaded by the high pressure, whereas the stepped piston 35 is subjected to the high pressure PH2 and the piston rod 36' is lifted off the valve element 30. The valve element 30 is therefore now only subjected to the control pressure Pat on the left-hand side. Consequently, during this braking effect the capacity of the pump 4 is controlled solely by the speed at which the drive motor 1 is driven, that is to say only by regulation of the accelerator pedal 2. During braking, therefore, the rotary speed can be regulated with the aid of the accelerator pedal 2 in the same way as during normal operation.

The construction of the control pressure generator 10 is optional. Different variations are known for this purpose. What is desired is that the control pressure should as accurately as possible follow the speed of the drive motor 1. The servovalve may additionally comprise two neutral position springs which are stationarily supported to define the rest position of the servo-valvq.

Claims

1. Hydraulic control equipment including a prime mover for driving a main pump and an auxiliary pump, the main pump supplying fluid to a main circuit and the auxiliary pump supplying fluid to a servo-control via a valve in a control circuit, the servo-control regulating the capacity of the main pump and the valve being controlled in dependence on a pressure signal related to the speed of the prime mover, wherein a pressure signal related to the pressure in the supply side of the main circuit is used to control the valve in opposition to the effect of the first-mentioned signal.

2. Control means for hydrostatic gear in which a drive motor drives a main pump which is adjustable by a servo-element and is connected to a hydraulic motor by way of a conduit system, wherein in said control means and with the aid of an auxiliary pump likewise driven by the drive motor and of a control pressure generator a speed-dependent pressure can be produced, the servo-element is actuated by a hydraulic servomotor to which pressure fluid can be fed by way of a servo-valve, and the servo-valve is adjustable by a piston arrangement of which the oppositely acting control pressure faces can be impinged by the control pressure which can be supplied to the main pump by way of a directional switching valve for determining the supply direction, characterised in that a counter-pressure opposing the currently effective control pressure can be exerted on the piston arrangement that is equal to the pressure obtaining in the conduit system at the supply side of the main pump.

3. Control means according to claim 2, characterised in that, apart from the control pressure face, the piston arrangement comprises an oppositely acting counter-pressure face for each supply direction of the main pump and the two counter-pressure faces are each connected to a respective side of the main pump.

4. Control means according to claim 2, characterised in that the piston arrangement comprises for each supply direction a stepped piston of which the larger face is the control pressure face and the smaller face is the counterpressure face.

5. Control means according to one of claims 2 to 4, characterised in that the piston arrangement comprises two parts each with a control pressure face and a counter-pressure face and both parts are adapted non-positively to engage the servovalve with opposite directional effect.

6. Control means according to claims 4 and 5, characterised in that two stepped pistons are adapted non-positively to engage an axially movable slide valve of the servo-valve from opposite sides.

7. Control means according to one of claims 2 to 6, characterised in that the servo-valve comprises two valve elements which are intercoupled by springs and of which one can be influenced by the piston arrangement and the other is connected to the servo-element by a feedback.

8. Hydraulic control equipment substantially as hereinbefore described with reference to, and as illustrated by the accompanying drawing.