US3803844A - Hydraulic transmission systems - Google Patents

Hydraulic transmission systems Download PDF

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Publication number
US3803844A
US3803844A US00324642A US32464273A US3803844A US 3803844 A US3803844 A US 3803844A US 00324642 A US00324642 A US 00324642A US 32464273 A US32464273 A US 32464273A US 3803844 A US3803844 A US 3803844A
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Prior art keywords
pump
prime mover
control
pressure
swept volume
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US00324642A
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English (en)
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A Gatiss
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Brockhouse Engineering Ltd
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Brockhouse Engineering Ltd
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/46—Automatic regulation in accordance with output requirements
    • F16H61/468—Automatic regulation in accordance with output requirements for achieving a target input torque
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/42—Control of exclusively fluid gearing hydrostatic involving adjustment of a pump or motor with adjustable output or capacity
    • F16H61/437—Pump capacity control by mechanical control means, e.g. by levers or pedals
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/46—Automatic regulation in accordance with output requirements
    • H—ELECTRICITY
    • H02—GENERATION; CONVERSION OR DISTRIBUTION OF ELECTRIC POWER
    • H02K—DYNAMO-ELECTRIC MACHINES
    • H02K9/00—Arrangements for cooling or ventilating
    • H02K9/02—Arrangements for cooling or ventilating by ambient air flowing through the machine
    • H02K9/04—Arrangements for cooling or ventilating by ambient air flowing through the machine having means for generating a flow of cooling medium
    • H02K9/06—Arrangements for cooling or ventilating by ambient air flowing through the machine having means for generating a flow of cooling medium with fans or impellers driven by the machine shaft

Definitions

  • a control system in hydraulic transmission apparatus 11/1966 Leasem, 60/431 comprising a prime mover, a variable swept volume positive displacement pump driven by said prime mover, a hydraulic machine operated by said pump, said machine being subjected, in use, to variable loading and being required to operate at different rates; the control system comprising, pump control means for varying the swept volume of the pump, prime mover control means for varying the power supplied to the prime mover, one operating element common to both the pump and prime mover control means and arranged to effect simultaneous control of both the swept volume of the pump and of the power supplied to the prime mover, said pump control means being arranged to effect reduction in the swept volume of the pump when the output pressure thereof exceeds a predetermined value dependent on the displacement of the operating element, said pump control means permitting the swept volume to be so reduced despite displacement of the operating element to, or towards, the position calling
  • This invention relates to a control system in or for hydraulic transmission apparatus, which apparatus is of the kind, herein called the kind specified, which comprises a variableswept volume positive displacement pump driven by a primemover, for example, an internal combustion engine, which pump is arranged to operate a hydraulic machine such as a positive displacement motor or ram, which machine is subjected in use to variable loading and isrequired to operate at different rates; the control system being of the type, herein called the type described, comprising pump control means for varying the swept volume of the pump and prime mover control means, includinga speed responsive governor, for varying the power supplied to the prime mover.
  • rate of energy supply for example, the rate of supply offuel in the case of an internal comefficient operating conditions for both the pump and the prime mover under varying loadings of the hydraulic machine and with minimum risk of the prime mover being overloaded.
  • a control system of the type described which is characterised by the provision of one operating element, common to both the pump and the prime mover control means and'arranged to effect simultaneous control of both the swept volume of the pump and of the power supplied to the prime mover, said pump control means being arranged to effect reduction in the swept volume of the pump when the output pressure thereof exceeds bustionengine, or the kilowattsin the case where the prime mover is an electric motor.
  • the pump may now be unable to maintain sufficient pressure to meet the load demand without requiring more driving torque than the prime mover is capable of; a condition which can only be overcome by reducing the swept volume of the pump.
  • both the pump and the prime mover are to operate efficiently carefully co-ordinated adjustment of both control means is required on the part of the operator, which calls for the exercise of skill and intelligence, so as in particular to avoid so reducing the prime mover power as, in the case of an internal com bustion engine, to cause this to stall.
  • the present invention has for its object the provision of an improved form of control system of the type described, which readily enables the operator to maintain a predetermined value dependent on the displacement of the operating element, said pump control means permitting the pump swept volume to be so reduced despite displacement of the operating element to, or towards, the position calling for maximum swept volume, said prime mover control means being arranged to control the power supplied to the prime mover by controlling the speed setting of the governor and means responsive to an' element associated with the prime mover control means for effecting reduction in the swept volume of the pump when the torque load from the pump on the prime mover commences to exceed the latters maximum available torque output.
  • a control system for an hydraulic transmission according to this invention as above defined in its widest form possesses the following advantages over previously known forms of control system of the type described of which we are aware, namely:
  • a control system embodying the present invention also reduces the risk of the prime mover being overloaded by adjustment of the pump capacity to too high a value in relation to the maximum power output of the prime mover.
  • a control system incorporating this invention can readily be so arranged that its effectiveness is not altered by changes in the rating of the particular form of prime mover provided, including for example changes in climate or altitude in the case of an internal combustion engine prime mover or loss of power due to malfunction of the prime mover.
  • the present invention is particularly applicable to the control of hydraulic transmission apparatus in which a wide range of speeds of operation on the part of the hydraulic machine are required.
  • load handling apparatus such as lifting cranes or winches and in fork trucks
  • an accurately controlled very slow speed commonly termed inching, may be needed during the manoevering of the truck to pick up or deposit a load, as well as a considerably higher transportation speed between, e.g., two widely spaced locations in a factory or warehouse complex.
  • variable swept volume positive displacement pump may be one embodying some form of eccentric control of capacity, for example, a vane type pump in which the swept volume per revolution is adjusted by varying the relative eccentricity of the stator and rotor.
  • the pump may be of the swash plate type embodying a rotatable cluster of pump cylinders, the outer ends of the pistons of which engage with the swash plate, the inclination of which to the axis of rotation of the cluster is adjustable to vary the working stroke of each piston and thus the swept volume of the pump per revolution of the cluster, so as to provide an infinitely variable adjustment between zero and maximum swept volume.
  • the single operating element may be mechanically connected to both the pump control means and the prime mover control means, and the latter may embody a mechanical connection to a power supply operating element including a movable stop element having a lost motion connection to the power supply operating element and, with the latter set to supply maximum power, the mechanical connection acting under the overriding control of the speed responsive governor to engage with a part of the pump control means to inhibit increase in the output of the pump beyond the capacity of the prime mover, the arrangement being such that decrease in prime mover speed at full power supply results in the governor acting to reduce the swept volume of the pump and thus inhibit overriding of the prime mover.
  • the operating element may comprise a foot pedal so as to provide single pedal control of both the pump and the prime mover.
  • a mechanically operated system has many practical limitations, for example, restrictions on the permitted relative disposition of pump and prime mover control means if they are both to be controlled mechanically from the one operating element, while mechanically oerated systems inherently may involve severe maintenance problems.
  • both the pump and the prime mover control means include hydraulic control valves which are simultaneously actuated, e.g., by the operating element, which may comprise a foot pedal, as in the mechanical arrangement above described.
  • the operation of the pump control valve to control the swept volume of the pump is under the control of both the pump output pressure and also of the governor which, on reduction of prime mover speed with the operating element at maximum setting, operate means controlling the swept volume of the pump to reduce the pump output and thus prevent overloading of the prime mover.
  • the operation of the prime mover control valve to control the power supplied to the prime mover is under the overriding control of a governor controlled valve, the governor being responsive to prime mover speed and operating to control the hydraulic fluid supplied to a power supply operating element to adjust the power supply; e.g., to adjust the fuel supply in the case of an internal combustion engine.
  • the governor controlled valve may include a pressure responsive element and the prime mover control valve may control the amount of hydraulic pressure supplied to the pressure responsive element to control the speed setting, as hereinafter defined. of the governor.
  • the increase in pressure supplied to the pressure responsive element consequent on displacement of the operating element serves to displace the pressure responsive element against the reaction force of the speed responsive governor thereon to increase the speed setting of the governor and hence feed an increased amount of fluid to the power supply operating element to increase the power supply to the prime mover, thus providing for simultaneous increase in engine power supply and speed setting of the governor.
  • speed setting is used herein toimean the speed at which the governor commences to be operative to inhibit increase in prime mover speed.
  • a hydraulically operated control system lends itself to controlling the operation, when required, of auxiliary machinery so as to take care of the additional power demand on the prime mover, e.g., the hydraulic lifting ram of the forks of a forktruck, namely by providing means for supplying additional pressure to the governor controlled valve to raise the speed setting thereof to give an increase in the power output of the prime mover to provide for the addtional power demand of the auxiliary machinery.
  • a hydraulically operated system readily lends itself to providing a reverse drive from the hydraulic machine which is particularly important where this is a rotary motor for driving a vehicle, namely by the hydraulic control of a change-over valve in the pump control system acting to reverse the operation of the pump control means, e.g., the direction of inclination of the swashplate in the case of a swashplate pump so that the direction of pressure flow from the pump to the motor is now reversed to reverse the direction of operation of the motor.
  • FIG. 1 illustrates a control system of a mechanical form
  • FIG. 2 is a hydraulic circuit diagram of a second embodiment of the invention
  • FIGS. 3 to 8 inclusive are sectional or part sectional views of certain details of the arrangement depicted in FIG. 2 and
  • FIG. 9 is a hydraulic circuit diagram of a third embodient of the invention, which is a modified form of the second embodiment of the invention and in which the same reference numerals are used in FIG. 9 as in FIG. 2 to refer to corresponding parts.
  • FIG. 1 of the drawings the invention is depicted as applied to hydraulic transmission apparatus of the kind specified in which the variable capacity positive displacement pump 20 is of the swashplate type earlier described, the swept volume of which pump can be varied in the known manner by adjusting the angle of the swashplate 21 between the full line position illustrated corresponding to zero pump output and the dashed line position indicated at 21 corresponding to maximum output of the pump.
  • the output port of the pump is connected to a hydraulic machine in the form of a positive displacement rotary motor 22 and the pump'is driven by a prime mover inthe form of an internal combustion engine, not illustrated but provided with the usual variable fuel supply control 23, the output speed of the engine being under the control of a governor 24.
  • the control system for this transmission apparatus comprises pump control means for varying the swept volume of the pump per working stroke and prime mover control means for varying the power supply to the engine, which two control means are, in accordance with this invention, operated by a single operat-- ing element which, as shown, is in the form of a single pivotally mounted foot pedal 25.
  • the pump control means comprises the aforementioned swashplate 21 which is biassed into the illustrated zero output position by a plunger 26 which is pressurised from the outlet or high pressure port 27 of the pump the displacement of the swashplate be yond the zero position being prevented by stop 28.
  • the swashplate is displaced from the zero position 21 towards and up to the maximum output position 21 by pumpcontrol plunger 29 forming part of the pump control means, which is engaged by compression spring 30 in engagement with spring stressing rod 31 connected through linkage 32 to the foot pedal 25.
  • pumpcontrol plunger 29 forming part of the pump control means
  • compression spring 30 in engagement with spring stressing rod 31 connected through linkage 32 to the foot pedal 25.
  • the prime mover control means comprise throttle actuating rod 33 carrying arm 34 supporting one end of bar 35, on the other end of which is slidable throttle actuating link 36, which is normally maintained by compression spring 37 in engagement with abutment 38 on the outer end of bar 35.
  • the throttle control rod 33 is supported for sliding movement in the direction of its length and connected at one end to a collar 39 engaged by one end of compression spring 40, the opposite end of which is engaged by spring abutment 41 connected through engine control linkage 42 to the foot pedal 25.
  • the arrangement is such that when the foot pedal is depressed towards the fully depressed position 25 the spring is increasingly stressed so as, through collar 39, rod 33 and arm 34, to increasingly stress spring 37 and displace the throttle actuating link 36 in a direction from the full line nearly closed position illustrated towards adjustable stop 43 which serves adjustably to limit the maximum opening of the engine throttle or fuel control valve, closing movement of which is simi larly limited by adjustable idling stop 44.
  • the collar 39 forms part of the governor 24, the weights 45 of which through the bell crank levers 46 act to slide the rod 33 against the pressure of spring 40 in a direction to reduce the fuel supply.
  • increase in the stress in spring 40 by depression of pedal 25 to increase the throttle opening increases the resistance of spring 40 against the action of the governor weights in reducing the throttle opening, i.e., depression of pedal 25 increases the engine speed at which the governor commences to'reduce the throttle opening and hence causes the throttleopening to be increased.
  • the output port 27 of the pump is connected in the usual way to the inlet port 47 of the motor of which the outlet or low pressure port 48 is connected through the usual low pressure oil reservoir 49 to the low pressure or inlet port 50 of the pump.
  • spring 40 will now be stressed so as through spring 37 to displace throttle link 36 towards the open position and at the same time spring 30 will be stressed to effect some displacement of the swashplate 21 away from its zero output position, i.e., into a working position, thus causing the pump 20 to deliver fluid to the motor under the power supplied to the pump from the engine consequent on the increase in engine speed above idling which will now have occurred.
  • the pedal 25 is merely depressed by an amount sufiicient to move the pump swashplate from the full line position 21 to provide enough pressure to the motor to enable itto move the load.
  • Such operation may result in a relative] small increase in engine speed above idling with corresponding increase in fuel supply.
  • control rod 33 of such a length that its free end 51 is adapted as shown at 51' to abut against the swashplate 21 and thus as explained below displace the swashplate in a direction away from the maximum swept volume position.
  • the pressure required of the pump is a function of the load on the motor and the relative speeds of the motor and pump and the stroke of the pump. Thus, if at one point of stability the pressure available is able to accelerate the motor, the pressure after such a period of acceleration will have dropped and hence the torque demand on the prime mover will drop and the prime mover will overspeed and retract rod 33 a little allowing an increase in stroke which in turn will raise the torque demand on the primer mover.
  • levers 21 and/or 36 may be too high to enable a satisfactory design based on the above completely mechanical system to be achieved. These difficulties are further enhanced by friction associated with these movements causing unsatisfactory hysteresis effects while the geographical position of the various elements might make mechanical connection difiicult.
  • the illustrated control system comprises a pump control means including a pump output control valve and a prime mover control means including a prime mover control valve 101 for controlling the speed of the prime mover, which is an internal combustion engine, through a governor valve 102 and hence controlling the power supply to the prime mover.
  • the governor valve 102 is responsive to the output speed of the engine, the fuel supplied to which is determined by the setting of a power supply operating element, in this case a throttle operating unit 104, under the overriding control of valves 101 and 102.
  • Both the governor 103 and its associated valve 102 are preferably constructed as a single unit (see FlG. 6).
  • the valves 100, 101 and 102 are a known type of balanced piston valve, with their respective pistons having a pair of axially spaced annular recesses 105, 106 in each case.
  • the pump output control valve 100 and the engine control valve 10] are both arranged to be simultaneously actuated by a single operating pedal 25 which, as shown, is formed as a bell crank lever and is arranged to displace a pedal bar 107, which through compression springs 108, 109 engaging opposite ends of the bar 107, provides for simultaneous displacement of the pistons of valves 100 and 101.
  • the arrangement is such that the depression of the pedal 25 is capable of increasing the compression of each of these springs and thus increasing by proportional amounts the displacing force which each of these springs exert on these valve pistons.
  • the pedal lever 25 is arranged .the latter'engages with the springs 108. 109.
  • pedal 25 The effect. of so depressing pedal 25 is to allow fluid under pressure from engine driven priming pump 110 which is being continually supplied through line 11 1 to the inlet port of each of thevalves of 100, 101 now to flow intolines 112 and 113 respectively.
  • Line 112 is connected through reverse valve 114 to one or the other, according to the position of valve 1 14, of lines 115, 116 which respectively lead to forward and reverse direction plungers 117, 118 for displacing against spring loading 119 the swashplate 120 of the pump 121, which is a known form of variable swept volume double acting swashplate pump.
  • the effect of supplying pressure to the forward or reverse swashplate displacing plungers 1 17, 118 is to displace the swashplate 120 from the illustrated full line zero output position in an anti-clockwise or clockwise direction as viewedin FIG. 2, corresponding respectively to .forward or reverse drive of the motor 122 which is a positive displacement rotary motor with the inlet and outlet ports 123, 124 thereof respectively connected to the pump delivery andsuction ports 1 25, 126. l i
  • the line 1 13 to which pressure fluid is delivered from the valve 101 is connected through a balancing passage 127 to the end face of the piston of valve 101 opposite to that engaged by,spring 109 and this line 113 is'arranged to supply fluid under pressure through line 113a to one end face of the piston 102a of governor valve 102.
  • the supply is through a two-way valve 128 which serves as later described alternatively to supply fluid to the one end face of governorvalve piston 102a at a pressure higher than that obtaining in'line 113,
  • the effect of increasing the pressure on the one end face of piston 102a is to apply to governor 103 a force acting against centrifugal force and thus increase the speed of rotation of the prime mover at which the governor displaces piston l02a downwardly in FIG. 2 against the pressure in line 113a. That is, the speed setting of the governor is increased.
  • the governor valve 102 is supplied through line 129 with fluid under pressure from priming pump 110 and the effect of supplying fluid under pressure through line 113 to the governor valve consequent-on the depression of pedal 25 is to displace the piston 102a of valve 102 against the centrifugally derived reaction force of the governor 103 thus, under the overriding control thereof, permitting fluid under pressure to flow from priming pump 110 through line 129 to throttle unit supply line 130.
  • This line 130 communicates with high pressure port 131 of the cylinder of thejthrottle unit 104 in which slides a throttle displacing plunger 132, the outer end of which is arranged to transmit opening movement to a throttle actuating lever 133 against the loading of a loading spring 133a for this lever.
  • the plunger 132 is hollow except at its outer end and receives the head 134 of the stem 135 of an anti-stalling piston valve 136, the function of which is to so act, as later described, to reduce the output of the pump when, with the engine throttle fully open, the engine is below its governed speed.
  • plunger 132 is provided with a collar 137 which is arranged to abut the underface of the stem head 134 to effect displacement of the antistalling piston valve 136 as the throttle actuating a plunger 132 is displaced'by more than a certain distance to open the throttle, i.e., to the right in FIG; 2.
  • the cylinder of throttle operating unit 104 is provided adjacent piston 136 with an anti-stalling port 140 which, through line 141, leads to return plunger 142, which when pressurised fromline 141 serves through link bar 143 (with which it is held in contact by light spring 142a) to bias the piston of pump control valve 100 in a direction for reducing the pressure suppplied to one or the other of the two swashplate displacing plungers 117, 118. 1
  • the anti-stalling piston valve 136 is provided with a pair of axially spaced annular ports 136a l36band in the position illustrated corresponding to idling position of the throttle, i.e., with the pedal not depressed, line 141 and associated port 140 is, through piston port 136a, connected to low pressure fluid tank T.
  • Port- 136b through radial passage 136a communicates with the bore of piston 136 which is of a diameter larger than that of stem 135, so that with port 13Gb in communication with anti-stalling port 140, the latter is now pressurized from port 131, subject to the overriding control of the governor valve 102.
  • piston 145 is displaced in a direction for connecting
  • This port 146 is connected through line 147 to swashplate angle reducing cylinder 148 in which works piston 149 biassed by light spring 149a to maintain its outer end in contact with link bar 143.
  • Piston 149 in like manner to piston 142 is adapted under the pressure in line 147 to displace link bar 143 in a direction to displace the piston of pump control valve 100 against the loading applied thereto from pedal 25.
  • piston 149 is arranged to act on link bar 143 at a position much nearer to the pivot or fulcrum 143a thereof than is the case with piston 142.
  • valve 144 to swashplate angle reducing piston 149 If the pump delivery pressure supplied from valve 144 to swashplate angle reducing piston 149 is great enough to overcome the force of spring 108, the piston of valve 100 will be displaced sufficiently in a direction to close the fluid supply from priming pump 110 through line 112 to the selected swashplate displacing plunger or will even be so displaced as to connect line 112 to the tank T of valve 100.
  • the result in any case is to effect reduction of the angularity of the swashplate under the loading of the appropriate spring 119 thereby reducing the output pressure of the pump 121.
  • the changeover valve 114 is a simple piston valve which may be operated by reverse drive lever 150 between forward drive and reverse drive positions in which the line supplying the fluid under pressure from pump control valve 100 is alternately connected to lines 115 and 116 leading to forward and reverse drive swashplate displacing plungers 117, 118. in each of these two positions the pressure within the one plunger which is not pressurised is relieved to low pressure tank T through the changeover valve 114 and the piston of this valve also provides an intermediate neutral position in which both of the lines 115 and 116 are simultaneously connected to the low pressure tank.
  • a return pressure line 151 extends from a lift control pilot valve 152, see FIG. 3, the purpose of which is to supply fluid through line 151 and valve 128 at a pressure to line 1 13a greater than that obtaining for the moment in line 113; as for example when pedal 25 is only partly depressed.
  • the effect of supplying fluid at this higher pressure from line 151 is to displace the valve member 128a of valve 128 from its nonnal full line position in which line 113a communicates only with line 113 to the dotted position of FIG. 2 in which line 113a now communicates only with line 151.
  • This feature is particularly advantageous as applied, for example, to fork trucks in which there is an additional demand on the engine when the lifting ram for the fork truck arms is required to be operated for the purpose of raising a load, thus putting on the engine an output demand additional to that required during the transportation of the loaded truck over the floor; it being commonly required to raise a load just picked up by the fork truck arms as the truck is being power advanced along the floor.
  • Such lift arms or other powered auxiliary devices are normally driven by a hydraulic system entirely separate from that here described but powered from the same engine, hence the necessity to provide the foregoing displacement of governor valve piston 102a to both increase engine speed and power, i.e., fuel supply.
  • the lift control pilot valve 152 is provided with a high pressure supply port 153 connected to a high pressure fluid supply, e.g., from priming pump 110.
  • the piston of the pilot valve 152 may be actuated by either or both of two side-by-side pivotally mountedlevers 154, arranged to apply pressure to the end of pilot valve actuating piston 155, which through spring 156 displaces valve piston 157 in a direction for supplying the high pressure fluid from port 153 to outlet port 158 connected to line 151 leading to a valve 128.
  • valve member 128a As the pressure supplied to line 15] becomes greater than that existing in line 113 the valve member 128a is displaced into the above mentioned dashed line position (FIG. 2) to increase the pressure in line 113a thereby raising the governor speed setting, i.e., the
  • the levers 154 when operated would also in the known way serve to actuate the separate hydraulic system above described.
  • the hydraulic transmission illustrated in FIG. 2 consisting primarily of pump 121 and motor 122 is of closed circuit configuration which is convenient for a reversingtransmission but which must be maintained full of oil. This is normally achieved by providing a priming pump, which may be the pump or as shown a second priming pump 110a and which is arranged to supply fluid from its delivery port 11% to either line connection 123 or 124, whichever is at the lower pressure, through two-way valve 159 of known form, and in a known manner.
  • a priming pump which may be the pump or as shown a second priming pump 110a and which is arranged to supply fluid from its delivery port 11% to either line connection 123 or 124, whichever is at the lower pressure, through two-way valve 159 of known form, and in a known manner.
  • valve piston 1590 of valve 159 are in communication with lines 160, 161, respectively leading to pump ports 125, 126. According to which of the latter is at the higher pressure, valve piston 159a is moved to connect the priming pump delivery port ll0b to that line 160 or 161 which is at the lower pressure.
  • the transmission is used for reversing, e.g., when embodied in a vehicle and is of closed circuit type as illustrated, it is necessary that the pressure signal conveyed in line 147 to piston 149 shall be that from the normal delivery port of pump 121 as determined by reversing valve 114.
  • pressure in line actuates plunger 1 17 to cause the pump under normal driving conditions to pump fluid from pump port 125 to motor port 124 the same pressure in line 115 will displace valve piston in valve 144 to connect line 125a with port 146 and line 147.
  • the lever 150 is moved to pressurise line 116 instead of line 115 to give reverse drive the piston 145 moves to connect line 1260 to line 147.
  • the pump control valve 100 and the engine control valve 101 which are arranged to be sirespective springs 108, 109 and pedal bar 107 be embodied in a single unit.
  • the latter may also include valve pistons 142 and 149 with their respective casings together with fulcrum bar 143 and all of these parts may be housed in a common casing so as to form a single unit, the details of which are shown in FIG. 4 and
  • the governor 103, together with the associated valve 102 and valve piston 102a may have the constructional form depicted in FIG. 6, in which these components also form a single unit with the governor, being driven at engine responsive speed through governor shaft 162 on which are mounted the radially movable governor weights l02b which engage with cam plate 163.
  • a light spring is provided for maintaining cam plate 163 in operative engagement with weights l02b when there is no pressure in line 113a.
  • FIG. 7 there is depicted in somewhatmore detail the form of the throttle operating unit 104 showing one possible constructional form thereof.
  • the high pressure lines from-the or each priming pump would be provided with the usual excess pressure relief valves, e.g., valve R in the case of pump 110.
  • the casing 166 provides a smallfchamber 170 in the pressure line in which the damperis provided;
  • This orifice 171 serves to damp out pulsations in the fluid flow within the line concerned and this is further assisted by the provision of the piston 167, which under sudden pressure increases moves inwardly of the casing 166 against the loading of springl69 so as temporarily to increase the capacity of the chamber 170 and thus relieve any sudden pressure increase.
  • Pressure relief valves are usually provided associated with and to act between the lines 125a and 126a as a protection against any excess pressure which may arise for any reason.
  • the governor valve 102 may be arranged so that the pedal 25 in the'fully up position, maintains a force on spring 109 sufficient to control the governor at a predetermined engine idling speed so that the governor permits a very small fluid flow through line to the throttle valve 104 so as slightly to open the engine throttle and provide for engine idling, any increase in speed above this being precluded by the action of governor'103 in thecompletely closing valve 102.
  • the acceleration of the truck or other vehicle can be stopped or deceleration effected by release of the pedal 25.
  • valve 100 increases the swashplate driving angle towards the maximum consistent with the driving pressure not exceeding the maximum as controlled by the spring 108 in its fully compressed condition and further restricted by the action of the throttle valve 136 which inhibits increased swashplate driving angle if the throttle is in its fully open position indicating that the load on the engine from the pump is excessive and beyond the engine capability.
  • valve 101 the maximum engine speed is demanded. As the engine reaches its maximum of speed, plunger 132 will move slightly to the left removing the inhibition of valve 136 to allow the swashplate driving angle to continue to increase up to the maximum driving pressure demand.
  • valve 102 For a given pressure in lines 113 and 113a which is controlled by the loading of spring 109, as soon as the reaction force derived from such pressure on governor piston 102a is exceeded by the oppositely acting reaction force from the governor weights, valve 102 will commence to close, thus effecting partial closure of the engine throttle from its previous fully open position.
  • the setting of governor 103 i.e., the speed at which it operates to limit engine speed by partial closure of the throttle, is dependent on the pressure applied through line 1130 to the governor valve piston.
  • piston 145 will be biassed in the direction for supplying the pump delivery pressure through line 147 to the underside of piston 149, which is in thrust engagement with fulcrum bar 143, thus through the latter, applying a thrust to the piston of valve 100, tending to displace this from the fully open position against the loading of spring 108, which is a preloaded spring and subject under the foregoing conditions to the pressure of the fully depressed pedal 25.
  • the characteristics of spring 108 are so selected as to ensure that when the pump delivery pressure in line 147 reaches a predetermined maximum, which may for example, be 3,000 P.S.l., the valve 100 is operated to adjust the pressure supplied through line 112 to the selected swashplate plunger, thus adjusting the swashplate angle, so as to prevent the output pressure of pump 121 exceeding the desired maximum value.
  • a predetermined maximum which may for example, be 3,000 P.S.l.
  • Deceleration is effected by partially or fully releasing the pedal 25. If the pedal is partially released the engine speed is reduced and this, with the vehicle moving, will cause the motor 122 to drive the pump 121 and thus there will be a severe reduction in pressure acting through the line 147 on the piston 149. Thus any load on the spring 108 will move the piston of the valve to cause the swashplate angle to increase. As mentioned above under these conditions the motor 122 will be driving the pump 121 and hence driving the engine at a speed higher than the governor setting which will therefore reduce the fuel supply to the minimum and provide a braking effort on the vehicle comparable with the torque required to motor the engine.
  • the load on the spring 108 may, if hydraulic braking is required, be arranged to reduce to zero in which case the springs 142a and 149a will cause the piston of the valve 100 to rise and reduce the swashplate angle to zero, or near to zero. Under these conditions there will be a high hydraulic resistance against the pumping action of the motor which will provide a large hydraulic braking effect limited by any pressure release valves associated with the lines and 126.
  • the motor 122 may be of the variable swept volume type. Such a provision enables the swept volume of the motor to be reduced for the purpose of obtaining a higher motor speed, with corresponding torque reduction, beyond the maximum speed obtained with full fuel supply and full swept volume of the pump; e.g., for fast vehicle running where the torque demand on the motor is below maximum.
  • FIG. 9 In order to achieve a more controlled form of hydraulic braking the embodiment described with reference to FIGS. 2 to 8 may be modified as shown in FIG. 9.
  • the pedal 25 and the pedal bar 107 are produced as a single component and in this embodiment the prime mover control valve 101 and associated element are disposed so that they are behind the pump control valve 100 and associated elements and so are not shown in FIG. 9.
  • the pedal bar 107 is arranged to act on the valves 100 and 101 in the same way as described with reference to FIG. 2.
  • the pedal 25 is provided with an extension part 200 on the opposite side of its pivot to the pedal bar 107.
  • the extension part 200 is arranged to operate on a coil compression spring 201 which acts on a nose element 202 connected through a lost motion device 203 to a spring abutment element 204.
  • the nose element 202 acts on an etension part 205 of the link'bar 143.
  • a second piston 210 in a cylinder 211 similar to the piston 149 and cylinder 148 arranged to pivot the bar 143 in the opposite direction to the'directions in which it is pivoted by the piston 149.
  • the pressure controlled piston valve 144 in this embodiment takes theform of a two port two positioned valve 220 which is again under the control'of the forward and reverse drive lever 150 and which again permits the normal drive pressure line of the main hydraulic circuit to be connected to the piston 149, through the line 147 as in the emboidment described with reference to FIG. 2 and in addition permits the return line of the main hydraulic circuitto be connected to a cylinder 211 of the piston 210 through the line 221.
  • the pedal 25 is acted upon by a coil tension spring 222 which is of such rate as to return the pedal 25 to its fully released position against any resistance applied thereto by the spring 201.
  • the spring 108 will probably be affected by the full movement of the lever 25 whilst the spring 201 will be restricted so that it is only operated as the pedal 25 approaches its fully released position.
  • the springs 108 and 201 will balance each other so that the system is set for equal pressure to exist in the main drive and return lines of the main hydraulic circuit. lf the pressure in the main drive line should predominate then the pressure therein will be fed by the line 147 to the cylinder 148 to cause the piston 149 therein to move to pivot the link bar 143 clockwise in FIG. 9 to move the piston of the valve 100 to decrease the swashplate angle and so reduce the pressure in the main drive line.
  • a control system in hydraulic transmission apparatus comprising a prime mover, a variable swept volume positive displacement pump driven by said prime mover, a hydraulic machine operated by said pump, said machine being subjected, in use, to variable loading and being required to operate at different rates; the control system comprising, pump control means for varying the swept volume of the pump, prime mover control means for varying the power supplied to the prime mover, one operating element common to both the pump and prime mover.
  • control means and arranged to effect simultaneous control of both the swept volume of the pump and of the power supplied to the prime mover, said pump control means being arranged to effect reduction in the swept volume of the pump when the output pressure thereof exceeds a predetermined value dependent on the displacement of the operating element, said pump control means permitting the swept volume to be so reduced despite displacement of the operating element to, or towards, the position calling for maximum swept volume, said prime mover control means including a speed responsive governor and being arranged to control the power supplied to the prime mover by controlling the speed setting of the governor and means responsive to an element associated with the prime mover control means for effecting reduction in the swept volume of the pump when the torque load from the pump on the prime mover commences to exceed the latters maximum available torque output.
  • the pump control means includes resilient biasing means which is increasingly stressed to effect an increase in swept volume on displacement of the operating element from the idling position, which biassing means is also capable of being further stressed under pressure derived from the pump output in such a manner as to allow the pump control means to reduce the swept volume even though the operating element may already be exerting its maximum pressure on the biassing means.
  • a control system according to claim 1 wherein the prime mover control means comprises a further resilient biassing means which is increasingly stressed to increase the speed setting of the governor on displacement of the operating element from the idling position to increase the power supplied.
  • a control system wherein the single operating element is mechanically connected to both the pump control means and the prime mover control means and the latter embodies a mechanical connection, to a power supply operating element, including a movable stop element having a lost motion connection to the power supply operating element and, with the latter set to supply maximum power, the mechanical connection acting under the overriding control of the speed responsive governor to engage with a part of the pump control means to inhibit increase in the output of the pump beyond the capacity of the prime mover, the arrangement being such that decrease in prime mover speed at full power supply results in the governor acting to reduce the swept volume of the pump and thus inhibit overloading of the prime mover.
  • both the pump and the prime mover control means each include a hydraulic control valve which are simultaneously actuated by said operating element, operation of the pump control valve to control the swept volume of the pump being under the control of both the pump output pressure and also of the governor, which on reduction of prime mover speed with the power supply operating element at maximum setting, operates means controlling the swept volume of the pump to reduce the pump output and thus inhibit overloading of the prime mover.
  • control system includes means responsive to said pump output pressure and operative on the pump control valve to cause the valve to operate to reduce the swept volume of the pump when the pump output pressure exceeds said predetermined value.
  • a control system according to claim 5 wherein operation of the prime mover control valve to control the power supplied to the prime mover is under the overriding control of a governor controlled valve, the
  • governor being responsive to prime mover speed and operating to control the hydraulic fluid supplied to the power supply operating element to adjust the power supply.
  • a control system wherein the governor controlled valve includes a pressure responsive element, the prime mover control valve controls the amount of hydraulic pressure applied to said pressure responsive element and wherein increase in pressure supplied to the pressure responsive element displaces said pressure responsive element against the reaction force of the speed responsive governor thereon to increase the speed setting of the governor to feed an increased amount of fluid to the power supply operating element, thereby to incease the power supply to the prime mover.
  • a control system including means raised to effect increase in the swept volume of the pump when the return pressure thereof exceeds a pedetermined value dependent on the displacement of the operating element.
  • both the pump and the prime mover control means each include a hydraulic control valve which are simultaneously actuated by said operating clement, operation of the pump control valve to control the swept volume of the pump being under the control of both the pump output pressure and also of the governor, which on reduction of prime mover speed with the power supply operating element at maximum setting, operates means controlling the swept volume of the pump to reduce the pump output and thus inhibit overloading of the prime mover, and the control system includes means responsive to said pump return pressure and operative on the pump control valve to cause the valve to operate to increase the swept volume of the pump when said pump return pressure exceeds said predetermined value.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Power Engineering (AREA)
  • Control Of Fluid Gearings (AREA)
  • Control Of Positive-Displacement Pumps (AREA)
US00324642A 1972-01-22 1973-01-18 Hydraulic transmission systems Expired - Lifetime US3803844A (en)

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GB317172A GB1417699A (en) 1972-01-22 1972-01-22 Hydraulic transmission systems

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Cited By (2)

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US4943756A (en) * 1989-12-05 1990-07-24 Crown Equipment Corporation Control of hydraulic systems
US20140060487A1 (en) * 2012-09-04 2014-03-06 Clark Equipment Company Utility vehicle horsepower management

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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS5638416B2 (de) * 1973-09-05 1981-09-07
IT7927506A0 (it) * 1979-11-23 1979-11-23 Linde Guldner Italiana S P A Dispositivo idraulico di sincronismo dell'alimentazione di un motore a combustione interna con la cilindrata variabile della pompa in un azionamento mediante trasmissione idrostatica.
US4400935A (en) * 1980-01-28 1983-08-30 Sundstrand Corporation Engine speed control

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US1259090A (en) * 1915-06-14 1918-03-12 Walter Ferris Control for hydraulic transmissions.
US1981805A (en) * 1932-12-31 1934-11-20 Kacer Bohuslav Driving mechanism
US3003309A (en) * 1959-01-30 1961-10-10 Dowty Hydraulic Units Ltd Single lever control apparatus for engine and hydraulic transmission
US3284999A (en) * 1966-02-01 1966-11-15 Sundstrand Corp Hydrostatic transmission

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US2148277A (en) * 1936-04-04 1939-02-21 Waterbury Tool Co Power transmission
DE905456C (de) * 1944-06-16 1954-03-01 Hanauer Pumpen Und Getriebebau Steuerung fuer Fluessigkeitspumpen
DE1400632B2 (de) * 1960-02-12 1971-12-23 Dowty Hydraulic Units Ltd , Tewkes bury, Glos (Großbritannien) Steuergeraet fuer ein aus einer kraftmaschine und einem hydrostatischen getriebe gebildetes antriebsaggregat
DE1219338B (de) * 1960-02-12 1966-06-16 Dowty Hydraulic Units Ltd Steuervorrichtung fuer ein hydrostatisches Getriebe, insbesondere fuer Kraftfahrzeuge
GB1035822A (en) * 1963-01-11 1966-07-13 Dowty Hydraulic Units Ltd Infinitely variable transmission
US3166891A (en) * 1963-07-08 1965-01-26 New York Air Brake Co Hydrostatic transmission
DE1253513B (de) * 1963-09-16 1967-11-02 Linde Ag Einrichtung zum Steuern einer Antriebseinheit, die aus einer Brennkraftmaschine und einer hydrostatischen Pumpe besteht
GB1137914A (en) * 1965-01-28 1968-12-27 Massey Ferguson Perkins Ltd Hydrostatic transmission and control therefor
DE1630137A1 (de) * 1967-07-13 1971-06-16 Bosch Gmbh Robert Hydrostatischer Fahrantrieb mit einer Steuer- und Regeleinrichtung

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Publication number Priority date Publication date Assignee Title
US1259090A (en) * 1915-06-14 1918-03-12 Walter Ferris Control for hydraulic transmissions.
US1981805A (en) * 1932-12-31 1934-11-20 Kacer Bohuslav Driving mechanism
US3003309A (en) * 1959-01-30 1961-10-10 Dowty Hydraulic Units Ltd Single lever control apparatus for engine and hydraulic transmission
US3284999A (en) * 1966-02-01 1966-11-15 Sundstrand Corp Hydrostatic transmission

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4943756A (en) * 1989-12-05 1990-07-24 Crown Equipment Corporation Control of hydraulic systems
US20140060487A1 (en) * 2012-09-04 2014-03-06 Clark Equipment Company Utility vehicle horsepower management
US9291105B2 (en) * 2012-09-04 2016-03-22 Clark Equipment Company Utility vehicle horsepower management

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DE2303014A1 (de) 1973-08-09
GB1417699A (en) 1975-12-17
DE2303014B2 (de) 1981-01-15
DE2303014C3 (de) 1981-09-03

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