WO2012031970A1 - Entraînement hydrostatique pour un véhicule automobile - Google Patents

Entraînement hydrostatique pour un véhicule automobile Download PDF

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Publication number
WO2012031970A1
WO2012031970A1 PCT/EP2011/065091 EP2011065091W WO2012031970A1 WO 2012031970 A1 WO2012031970 A1 WO 2012031970A1 EP 2011065091 W EP2011065091 W EP 2011065091W WO 2012031970 A1 WO2012031970 A1 WO 2012031970A1
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WO
WIPO (PCT)
Prior art keywords
pressure
hydraulic
accumulator
shaft
drive
Prior art date
Application number
PCT/EP2011/065091
Other languages
German (de)
English (en)
Inventor
Siegfried A. Eisenmann
Original Assignee
Eisenmann Siegfried A
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Eisenmann Siegfried A filed Critical Eisenmann Siegfried A
Priority to EP11749441.9A priority Critical patent/EP2614274A1/fr
Publication of WO2012031970A1 publication Critical patent/WO2012031970A1/fr

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Classifications

    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60K—ARRANGEMENT OR MOUNTING OF PROPULSION UNITS OR OF TRANSMISSIONS IN VEHICLES; ARRANGEMENT OR MOUNTING OF PLURAL DIVERSE PRIME-MOVERS IN VEHICLES; AUXILIARY DRIVES FOR VEHICLES; INSTRUMENTATION OR DASHBOARDS FOR VEHICLES; ARRANGEMENTS IN CONNECTION WITH COOLING, AIR INTAKE, GAS EXHAUST OR FUEL SUPPLY OF PROPULSION UNITS IN VEHICLES
    • B60K6/00—Arrangement or mounting of plural diverse prime-movers for mutual or common propulsion, e.g. hybrid propulsion systems comprising electric motors and internal combustion engines ; Control systems therefor, i.e. systems controlling two or more prime movers, or controlling one of these prime movers and any of the transmission, drive or drive units Informative references: mechanical gearings with secondary electric drive F16H3/72; arrangements for handling mechanical energy structurally associated with the dynamo-electric machine H02K7/00; machines comprising structurally interrelated motor and generator parts H02K51/00; dynamo-electric machines not otherwise provided for in H02K see H02K99/00
    • B60K6/08—Prime-movers comprising combustion engines and mechanical or fluid energy storing means
    • B60K6/12—Prime-movers comprising combustion engines and mechanical or fluid energy storing means by means of a chargeable fluidic accumulator
    • 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
    • F16H61/4078—Fluid exchange between hydrostatic circuits and external sources or consumers
    • F16H61/4096—Fluid exchange between hydrostatic circuits and external sources or consumers with pressure accumulators
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60Y—INDEXING SCHEME RELATING TO ASPECTS CROSS-CUTTING VEHICLE TECHNOLOGY
    • B60Y2304/00—Optimising design; Manufacturing; Testing
    • B60Y2304/07—Facilitating assembling or mounting
    • B60Y2304/072—Facilitating assembling or mounting by preassembled subunits
    • 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
    • F16H61/4078—Fluid exchange between hydrostatic circuits and external sources or consumers
    • F16H61/4139—Replenishing or scavenging pumps, e.g. auxiliary charge pumps
    • 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
    • F16H61/421—Motor capacity control by electro-hydraulic control means, e.g. using solenoid valves
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/60—Other road transportation technologies with climate change mitigation effect
    • Y02T10/62—Hybrid vehicles

Definitions

  • the invention relates to a hydrostatic drive for a motor vehicle, in particular a passenger car, according to the preamble of claim 1.
  • Range of be 230 grams per KW and hour.
  • Range of double namely at about 450 grams per KW and hour.
  • the aim is to suppress the engine speed and to keep the engine torque high at higher vehicle speed in the plane and relatively low traction values at the vehicle.
  • Construction costs in the automatic transmissions and the switching effort in the manual transmissions is correspondingly high. However, the vehicle then acts in the high gears in the
  • ISG integrated starter generator
  • On-board voltage for example 42 volts. This will be the
  • Interposed energy storage are proposed in the prior art on the one hand electrical connections and on the other hydraulic connections with intermediate storage in the form of accumulators or pressure accumulators.
  • Hydraulic pressure increased from approx. 200 bar to 500 bar, so that the hydromotors for the wheel drive deliver the torque required in extreme cases. Apart from that, these
  • the object of the invention is therefore one of a
  • Motor vehicle comprises at least one low-pressure hydraulic accumulator for receiving a working fluid, in particular hydraulic fluid, and at least one high-pressure hydraulic accumulator also for receiving the working fluid, in particular hydraulic fluid, and at least one high-pressure hydraulic accumulator also for receiving the working fluid, in particular hydraulic fluid, and at least one high-pressure hydraulic accumulator also for receiving the working fluid, in particular hydraulic fluid, and at least one high-pressure hydraulic accumulator also for receiving the working fluid, in particular hydraulic fluid, and at least one high-pressure hydraulic accumulator also for receiving the
  • the hydraulic accumulators can be any pressure-tight hydraulic containers which are suitable for use in a motor vehicle.
  • the hydraulic accumulators can in turn be composed of a plurality of smaller accumulators.
  • An accumulator in particular a gas reservoir, for example in the form of a nitrogen cylinder, is connected downstream of the high-pressure hydraulic accumulator for biasing with compensating pressure in the gas region.
  • the pressure accumulator is used when filling the high-pressure hydraulic accumulator with
  • Pressure accumulator pressed and thus stored under pressure.
  • the separation between the gas and the working fluid can be done for example by means of a plastic membrane, as is the case with known from the prior art bladder or diaphragm accumulators, as which the high-pressure hydraulic accumulator can be formed.
  • a liquid hydraulic pump is with a
  • the fluid hydraulic pump is in pressure communication between the low-pressure hydraulic accumulator and the high-pressure hydraulic accumulator for filling the high-pressure hydraulic accumulator with the working fluid under working pressure against the equalizing pressure of the
  • the pressure connection can via a line system, in particular an arrangement of
  • Hydraulic hoses and valves to be created the liquid hydraulic pump is used to fill the high-pressure hydraulic accumulator with working fluid from the low-pressure hydraulic accumulator under the working pressure, the energy required for this purpose
  • At least one hydraulic drive motor is coupled with at least one drive wheel of the motor vehicle directly or indirectly, in particular rotationally rigid.
  • the hydraulic drive motor is designed as a wheel drive, wherein in particular two drive wheels - in one possible From leadership form all driving wheels of the motor vehicle - each associated with a hydraulic drive motor.
  • Transfer case is arranged and the hydraulic
  • Drive motor optionally drives a plurality of drive wheels.
  • the at least one hydraulic drive motor has a high-pressure connection and a low-pressure connection.
  • the high pressure port provides the pressure input and the
  • Low pressure port is the pressure output.
  • a forward mode of the hydrostatic drive is the
  • High pressure port connected to the high pressure hydraulic accumulator, with the low pressure port connected to the
  • Low-pressure hydraulic accumulator is connected. Since the pressure in the high-pressure hydraulic accumulator is higher than in the low-pressure hydraulic accumulator, the interposed
  • the at least one hydraulic drive motor with the working fluid from the high-pressure hydraulic accumulator for driving the at least one drive wheel is operable.
  • To control the drive is in one possible
  • the drive motor in particular a flow control valve, arranged, by means of which a limited flow of current, that is a metered amount of working fluid, the drive motor can be supplied.
  • the working fluid flow regulator in the extreme case, a sufficiently large flow rate or has a parallel bypass valve.
  • Throttle is controlled. This system requires a certain pressure difference, which generates hydraulic losses. In other words, this type of regulation of
  • Discharge speed of the hydraulic accumulator and thus the driving speed is not possible without pressure loss between hydraulic accumulator and hydraulic motor.
  • an electric motor driven metering pump without pressure difference between the input and the output.
  • the dosage is controlled by the speed of an example electronically commutated DC motor, similar to a hydraulic power steering, in which the metering amount for the steering cylinder on the
  • Dosing pump is coupled, is varied.
  • Such metering pumps are, for example, so-called gerotor or gerotor pumps, whose leakproofness makes no special demands, since the pressure difference is very small or virtually zero. In this way, one is as far as possible
  • Hydromotors must be changed.
  • the electric motor can be dimensioned very small, since the metering pump practically no torque because of the dropping pressure difference
  • the invention therefore provides in an embodiment that the working fluid flow regulator is used as a metering pump
  • the high-pressure hydraulic accumulator is divided into a plurality of smaller high-pressure hydraulic accumulators, which are coupled together with the downstream accumulator.
  • the downstream one spans
  • each of the plurality of smaller high pressure hydraulic accumulators can be independently connected to the at least one hydraulic drive motor, the at least one hydraulic drive motor
  • the at least one hydraulic drive motor has a displacement volume which can be varied either steplessly or stepwise in an adjustment range per shaft revolution, the specific absorption quantity of the hydraulic fluid
  • Invention is the hydraulic drive motor as
  • Hydraulic drive motor with continuous displacement per shaft revolution can according to the invention also at least one gradual adjustable hydraulic drive motor for
  • Such an engine is formed for example by an axial piston motor with a plurality of pistons, in particular 24 pistons, wherein the pistons individually and / or groups decoupled, short-circuited and / or sixteenschliessbar, so that in particular the number of switching states corresponding number of intake volumes and thus gear ratios results.
  • part of it is exclusively a steplessly adjustable one
  • hydraulic drive motor reference including here also a - in particular in a plurality of stages and thus substantially continuously adjustable - graded
  • the plurality of smaller high pressure hydraulic accumulators may be connected to the fluid hydraulic pump such that the plurality of smaller high pressure hydraulic accumulators are substantially independently connectable and alternately inflatable with the fluid hydraulic pump.
  • Valve arrangement in particular a plurality of directional control valves
  • valve arrangement comprises, in pressure communication between the plurality of smaller high pressure hydraulic accumulators, the at least one hydraulic drive motor and the liquid hydraulic pump is arranged.
  • Liquid hydraulic pump is connected for filling and another of the several smaller high-pressure hydraulic accumulator connected exclusively to the at least one
  • hydraulic drive motor is connected for emptying.
  • the plurality of smaller high-pressure hydraulic accumulators can be alternately operated either by the liquid hydraulic pump or in the recuperation operation of the Drive motor filled with working fluid or to
  • a development of the invention provides that the
  • hydrostatic drive includes a mode switching unit, for example, as a selector slide,
  • the mode switching unit is designed in particular as a 4/3 way valve.
  • the mode switching unit is arranged in the pressure connection between the at least one hydraulic drive motor and the high-pressure hydraulic accumulator and in the pressure connection between the at least one hydraulic drive motor and the low-pressure hydraulic accumulator.
  • recuperation mode switchable.
  • High-pressure hydraulic accumulator connected.
  • the connections to the hydraulic drive motor in the recuperation mode are directly or indirectly reversed.
  • the engine can be switched to a pump mode and thus to a recuperation mode in overrun mode, as will be explained in more detail below.
  • the invention also provides a especially
  • Control connection is such and the control device is formed such that when driving the motor vehicle in the forward-train operation, the forward mode and at
  • the recuperation mode is granted.
  • the switching can take place by means of a switch arranged on the gas and / or brake pedal of the motor vehicle, wherein the operating mode switching unit has an electrical actuator between the two operating modes
  • the braking power is in particular by means of a brake pedal sensor and a flow control valve
  • the at least one hydraulic acts
  • a drive motor as a hydraulic pump driven by the at least one drive wheel which pumps the working fluid from the low-pressure hydraulic accumulator into the high-pressure hydraulic accumulator for regenerative braking energy use. After switching to the forward mode, this energy thus obtained can be used again to accelerate the motor vehicle.
  • the at least one hydraulic drive motor is
  • the at least one hydraulic valve Preferably, the at least one hydraulic valve
  • a filling pump is arranged, which is in the pressure connection between the low-pressure hydraulic accumulator and the at least one hydraulic drive motor and pumps towards the at least one hydraulic drive motor.
  • the filling pump is prevented from being located in the intake line between the low-pressure hydraulic accumulator and the hydraulic pump
  • Filling pump per shaft revolution should be much greater than the smallest possible amount of swirl of the at least one variable volume hydraulic drive motor per
  • Suction amount of the filling pump approximately equal to half the maximum specific amount of swallowing of the at least one hydraulic drive motor per shaft revolution.
  • the ratio of the gas pressure of the pressure accumulator between the maximum value and the lowest value preferably has the
  • hydrostatic vehicle drive can improve hydrostatic vehicle drive according to the invention.
  • a well-suction pump driven by the motor shaft preferably designed as a gerotor pump, causes a charging effect, as described above. This can be done in a closed circuit, but preferably in a hermetically separate to the outside world total hydraulic system. Since the as
  • the pressure fluctuation determines the torque fluctuation of the wheel motors in the drive wheels.
  • the pressure variation ie the ratio P1 / P2 determines the size of the gas storage together with the gas volume proportion of the high-pressure oil reservoir. From these complex contexts results according to the invention the requirement that the high-pressure hydraulic accumulator is divided into several smaller high-pressure hydraulic accumulator, the serially connected provide a sufficient working volume dV. As a result, the volume of the downstream pressure accumulator in the form of the gas storage is greatly reduced.
  • the three high-pressure hydraulic accumulators each with 9 liters of maximum oil content, are connected in series, resulting in a total content of 27 liters per liter
  • the efficiency of the system is calculated from the
  • n 2000 rpm
  • Oil flow 90 1 / min; Loss energy 882 Ws;
  • the high-pressure hydraulic accumulator is divided into three smaller units and designed in particular as a bladder accumulator. They become high pressure side
  • Algorithm of memory filling and emptying is not performed by means of electrically operated, but by means of hydraulically operated directional control valves.
  • the impulses come then directly from the fluctuations of the oil pressures of the hydraulic accumulator.
  • the cyclic control of the system is then completely independent of the on-board electrical system of the vehicle.
  • Such hydraulically operated directional control valves are available from the
  • the torque of a hydraulic motor is calculated according to the formula:
  • Torque remains sufficiently constant at the drive wheel of the vehicle. In other words: falls z. B. in the emptying of the high-pressure hydraulic accumulator, the working pressure on the hydraulic motor due to the system, is due to the automatic displacement control on the controllable hydraulic motor the
  • High-pressure accumulator is automatically ensured by a slightly faster emptying.
  • the electronic two-edge control can be dispensed with, if the continuously variable wheel motors are designed in such a way that the control spring and control piston automatically increase their displacement when the supplied working pressure P decreases, as explained above, as shown below. Even so, the torque of the drive wheels can be kept constant despite fluctuating working pressure. In the calculation example mentioned above, only a cyclical correction of the adjustable engine intake of approximately 16% would be required.
  • the operating point of the engine is usually removed from the point of lowest specific fuel consumption.
  • Working fluid flow controller in particular the adjustable flow rate valve, thus must be designed large enough or by a parallel-connected bypass valve, which is also referred to as a bypass valve, are added. In this driving condition then the speed of the
  • Combustion engine increased to almost maximum value with simultaneous maximum torque at maximum power. Since the hydraulic pump to the motor preferably coupled torsionally rigid is, the filling time for the hydraulic accumulator is extremely short. Accordingly, the cycle frequency of the serially connected memories increases. The constant, cyclic starting and stopping of the
  • the very noiseless and vibration-free running of the WANKEL engine offers z. B. the possibility that the engine passes through at a constant speed and that the power slide switches from the high idle at the same speed for power output to the hydraulic pump.
  • hydraulic drive according to the invention thus has the possibility of the WANKEL engine with its known disadvantage of much higher fuel consumption and its higher CO 2 emission (up to 300 g / km) as efficient and, above all, very noise
  • Adjustment has continuously variable displacement per shaft revolution. Preferably, the comes
  • Rotary piston machine filed Jun. 23, 2010, assigned to the same Applicant, is used as a hydraulic drive motor, but it is also possible to use a different one
  • volume variable hydraulic motor in particular hydrostatic high-torque rotary piston engine according to the orbit principle, which is infinitely variable in an adjustment range
  • the hydrostatic high-torque rotary piston engine is a rotary engine after the
  • Orbit principle with an internally toothed stator is fixed in space, designed as a rotary piston rotor
  • WO 2009/092719 is known.
  • Adjustment of the amount of swallowing of such a gerotor pump by varying the effective meshing width between a rotatable inner rotor and a rotatable outer rotor, as generally known in the art. Since in such a hydraulic pump, the axial distance of the two rotors is fixed in space, these approaches are not suitable for use in high-torque rotary piston engine according to the orbit principle. Thus, the constructive task is a new and much more complicated type than a gerotor pump.
  • hydrostatic high-torque rotary piston engine represents a further development of the hydrostatic rotary piston engine known from WO 2006/010471, for which reason the general basic structure of the hydrostatic rotary piston engine, which is already essentially known from the state of the art, will be discussed in the following
  • hydrostatic rotary piston engine also called
  • low-speed high-torque rotary piston engine may be referred to, includes a power output acting as a power output, in the housing of the rotary piston engine
  • the power section consists essentially of a fixed, centric stator, a
  • movable rotary piston as a rotor and a centrally rotatably mounted, serving as an output shaft together.
  • the stator has a stator internal toothing with the
  • the rotary piston has a rotor outer toothing partially engaging in the stator inner toothing of the stator and having a number of teeth c and a rotor internal toothing having a number of teeth b.
  • the rotor has a rotor outer diameter.
  • the wave meshes with her first Shaft external teeth with a number of teeth a partially the rotor internal teeth of the rotary piston, wherein the rotary piston for performing an orbital movement is arranged and dimensioned so eccentric to the shaft axis and the
  • An inlet and outlet part is used for supply and disposal of the power unit with the working fluid, wherein the working fluid at a working pressure via a high pressure port, which serves as an input, the rotary piston engine can be supplied.
  • the rotary valve is rotatably mounted about the central, longitudinally extending shaft of the geometric shaft axis.
  • Tooth chambers is directed out to a low pressure port, so that the working pressure in the first part of
  • Toothed chambers leads to an orbit movement of the rotor and the working fluid displaced from the second part of the tooth chambers, whereby the shaft is set in rotation, and vice versa.
  • the rotary valve has a specific rotary valve outside diameter, either by the actual Outer diameter of the rotary valve or by the outer or average diameter of the disc-shaped
  • Pressure valve radially arranged pressure resistant is determined.
  • the rotary engine comprises a gear transmission in the form of a wobble gear
  • Shaft external teeth of the shaft with a number of teeth w and an internal toothing of a fixed internal gear with a number of teeth z is arranged as a synchronous drive for the rotary valve.
  • Low pressure port is connected, arranged.
  • the shaft is mounted on both sides of the power unit immediately adjacent radial bearings, in particular bearings.
  • Rotary piston engine a continuous shaft with large shaft diameters and high torsional strength can be used.
  • both shaft ends can be exposed to a high torque flux and, for example, both
  • shaft ends as output, or one shaft end as the output and the other shaft end for connecting a brake or a second drive, whereby the entire
  • Taumelgetriebe has an eccentric, which has an eccentric internal teeth with a number of teeth x and an eccentricthinkvermay with a number of teeth y.
  • the eccentric internal teeth partially mesh with a second
  • a cup-shaped connecting part is rotatably coupled to the eccentrically rotatable eccentric and with the centrically rotatable rotary valve in
  • Transmission ratio 1 1 rotatably connected via a wobbling connection such that a rotational movement of the shaft via the eccentric to the rotary valve is transferable.
  • the number of teeth a is smaller than the number of teeth b, the number of teeth c smaller than the number of teeth d, the number of teeth w smaller than the number of teeth x and the number of teeth y smaller than the number of teeth z, the respective tooth difference preferably one or two teeth.
  • Equation represents the speed ratio shaft to the rotary piston or shaft to the rotary valve:
  • the equation expression is a positive integer, preferably equal to 3. It must also be sought that in this
  • Tumble angle is correspondingly large. However, then the rotational speed of the eccentricity is relatively small.
  • this ratio is preferably between -3 and -9.
  • the stator internal toothing is formed by circular-cylindrical rollers, which leads to further increased pressure and excellent service life. Measurements have shown that by the transition to roles in the stator of the starting efficiency and also the stator internal toothing.
  • a rotationally fixed axial second balance piston is preferably provided, which acts axially on the rotary valve.
  • the axial second compensating piston is arranged, for example, around the wobble gear.
  • the fixed internal ring gear is in the axial second
  • Balancing piston formed.
  • the force on the rotary valve is applied for example by means of a plate spring.
  • the output-side radial bearing needs a higher radial load rating for additional recording of the wheel load.
  • it should be arranged as close to the center of the wheel. It is advantageous if this bearing is as close as possible to the wheel flange and
  • the shaft is preferably formed integrally with a wheel flange on which a wheel for direct drive is directly mountable.
  • the rotary piston is parallel to the fixed stator and relative to the axially fixed shaft
  • the first shaft outer toothing of the shaft and the stator inner toothing of the stator extend axially such that the inner rotor toothing can be displaced axially on the first outer shaft toothing and the outer rotor toothing on the inner stator toothing.
  • the rotary piston engine adjusting, which with the
  • Rotary piston axially to its axial displacement are engaged.
  • Under adjustment means are generally means for axially displacing the rotary piston to understand
  • Housing the rotary piston engine can be arranged.
  • the adjusting members between the rotary piston and the rolling bearings are left and right axially displaceable.
  • a radially extending second wall of the tooth chambers is axially displaceable together with the rotary piston and axially coupled to the rotary piston.
  • the second wall is the first wall spaced parallel opposite, wherein the
  • Toothed chambers between the first and second wall are arranged and extend there axially.
  • the first wall is formed by a so-called rotor plate and the second wall by a so-called stator plate.
  • the rotor plate and the stator plate seal the tooth chambers axially. This means that there is a tight contact both between the stator plate and the stator internal teeth and the external rotor teeth, as well as between the rotor plate and the stator inner teeth and the rotor outer teeth.
  • the tooth chambers of the stator plate, the rotor plate, the stator internal teeth and the rotor outer teeth are limited.
  • the rotor plate is axially fixed and radially coupled to the rotary piston so that it also performs the orbital motion of the rotary piston. The radially with the
  • Rotor piston coupled rotor plate has in the housing one of the orbital movement of the rotor plate corresponding free
  • the stator plate Mobility in the radial direction and in the circumferential direction.
  • the stator plate is axially movable and axially coupled to the rotary piston, but decoupled radially from the rotary piston, so that they so no orbital motion and no rotational movement performs, but is rotationally fixed.
  • the volume of the tooth chambers is infinitely variable by the axial displacement of the rotary piston and thus also the stator plate by means of the adjustment by the distance between the axially displaceable stator plate and the axially fixed rotor plate can be reduced and enlarged.
  • the axially fixed rotor plate engages with its internal rotor-plate toothing in the external rotor toothing, both of which have the number of teeth c and have similar Vernierungsgeometrien so in a sealing manner in the axial direction a perfect fit, that the rotary piston is axially displaceable relative to the axially fixed rotor plate. Due to the fit of the
  • Rotor plate internal teeth with the rotor external teeth are both in axial sealing, but axially movable contact.
  • the rotor plate in the housing is guided radially sealingly in such a way that the rotor plate engages its orbit movement seals the tooth chambers outwards. This is done, for example, in that the rotor plate in a disc-like recess in the housing, in particular between a first housing part and a second
  • Housing part of the housing mounted radially movable and radially sealing.
  • the radial sealing takes place in particular by means of an adjacent to the stator, radially outwardly extending, pointing in the axial direction wall, on which also a radially extending, in axial
  • the axially movable stator plate preferably engages with its stator outer teeth in the stator inner toothing in the axial direction sealing manner a precise fit.
  • the stator plate external teeth and the stator internal teeth have the same number of teeth d and essentially the same
  • stator plate is axially displaceable relative to the axially fixed stator together with the rotary piston. Due to the fit of the stator plate
  • Stator plate external teeth with the stator internal teeth are both in axially sealing, but axially movable contact.
  • the balance piston engages in an axial
  • Adjusting cylinder on.
  • the balance piston is arranged in the fixed housing and acted upon by the working pressure of the working fluid of the tooth chambers, by a direct or indirect connection between the
  • Compensating cylinder and the high-pressure connection consists.
  • the Pressurization takes place in such a way that the working pressure in the compensating cylinder over the stator plate at least partially counteracts the working pressure in the tooth chambers and thus a force on the rotary piston in the direction of a
  • Dental chambers acts.
  • Balancing cylinder results, exercised. These two forces counteract each other and partially compensate, so that the axial holding and adjusting force of the rotary piston is reduced.
  • the balance piston and the balance cylinder have
  • a ring shape for example, a ring shape and in particular extend around the shaft. Preferably, they are between the
  • Balance piston and balance cylinder can be reversed. Under a balance piston is thus also a balance piston
  • Adjusting members are designed as a control piston which is axially engaged with the rotary piston.
  • the control piston with the rotary piston is so axial
  • control piston engages in an axial control cylinder arranged in the fixed housing such that a force can be exerted on the rotary piston for axial displacement of the rotary piston and for changing the volume of the tooth chambers.
  • control cylinder can be acted upon by a control pressure.
  • Control cylinder and the control piston arranged such that the rotary piston is displaced by the pressurization in the control cylinder in a direction such that the
  • Control pressure an increase in the volume of the dental chambers.
  • control piston and the control cylinder may be arranged annularly in the housing around an output section of the shaft in a ring-shaped manner, wherein the output section corresponds to that
  • a radially extending intermediate disc is arranged between the control piston and the rotary piston. The rotary piston slides in the execution of its orbital motion with its in axial
  • the control piston is arranged for example in such a way that it counteracts the rotary piston, in particular via this intermediate disk presses the opposite balance piston and in particular against the control springs.
  • the working pressure in the compensating piston counteracts the control pressure in the control piston kinematically.
  • the control pressure in the control cylinder can be generated in one possible variant by the working pressure of the working fluid.
  • the control piston can be acted upon by the working pressure of the working fluid.
  • the control pressure is equal to the working pressure.
  • the effective surfaces of the control cylinder, the control piston, the stator plate, the rotor plate and optionally the compensating cylinder and the compensating piston are in this case preferably dimensioned such that the control cylinder upon increase in the
  • Sucking volume is available, for example
  • Control pressure connection acts on the control piston.
  • Control pressure can be varied.
  • one of the control piston is selectively acted upon, in particular by means of a valve either with the working pressure of the working fluid or with a control pressure from the control pressure port.
  • the hydrostatic rotary piston engine is not only used as a motor to generate an output, but is also excellent as a pump.
  • the hydrostatic rotary piston engine has the great advantage that a recuperative braking energy recovery is possible by the particular designed as a wheel motor hydraulic motor is switched in the case of braking to a pumping operation and receives torque.
  • this hydrostatic rotary piston engine sucks the working fluid from the tank and conveys it under pressure into one or more
  • Hydraulic accumulator The stored energy can be stored in the
  • the rotary piston machine can therefore be changed, in particular reduced, when changing from traction mode to overrun mode, which is also called suction or pumping mode
  • variable volume hydrostatic rotary piston engine as well as the
  • Hydraulic motor according to WO 2006/010471, the continuous wave. If, for example, at maximum vehicle speed, the oil pressure flow is too great, despite the reduced amount of swirl, in order to avoid cavitation during pump suction operation, the following option is available: At the output of the shaft
  • the continuous shaft driven filling pump for example, as a gerotor pump, which in their
  • This filling pressure can be far above the atmospheric pressure, so that in a large
  • the rotary valve As large as possible.
  • the rotary valve outer diameter is larger by a factor of 1.3 to 1.5 than the rotor outer diameter.
  • the hydrostatic rotary piston engine can thus as
  • variable displacement allows economical operation of the rotary engine, since the displacement of the current power and
  • Fig. La is a schematic representation of the hydraulic
  • Fig. Lb is a schematic representation of the hydraulic
  • Fig. Lc is a schematic representation of the hydraulic
  • High pressure hydraulic Speieher shows a schematic representation of a hydraulic drive motor with a filling pump
  • Fig. 3 is a schematic plan view of a with the
  • Fig. 4 is a longitudinal section through a wheel motor
  • Fig. 5 is a same longitudinal section in the position of
  • Fig. 6 shows a cross section along the section line A-A of
  • Fig. 7 shows a cross section along the section line B-B of
  • Figure 4 which illustrates the axially displaceable, but not rotatable, the axial wall of the tooth chambers representing stator plate, which is hatched from bottom left to top right, in cross section;
  • Fig. 8 shows a cross section along the section line C-C of
  • exporting rotor plate which is hatched from bottom right to top left, represents in cross section; 9 shows a longitudinal section through the wheel motor in the region of the required pressure lines for acting on axially acting compensation surfaces which are to be under the working pressure;
  • Fig. 11 is a longitudinal section through an alternative
  • FIG. 3 shows the positioning of the components of a corresponding drive in a motor vehicle.
  • the figures la, lb, lc and 3 are therefore described together below.
  • the drive comprises a low-pressure hydraulic accumulator 111 in the form of a pressure vessel for receiving a Working fluid in the form of a hydraulic oil and a high-pressure hydraulic accumulator 104 also for receiving the working fluid.
  • the high-pressure hydraulic accumulator 104 is divided into a plurality of smaller high-pressure hydraulic accumulators 104a, 104b and 104c. These are connected together downstream of a pressure accumulator 105 in the form of a nitrogen bottle. This accumulator 105 is used for biasing the high-pressure hydraulic accumulator 104 with compensating pressure.
  • the high-pressure hydraulic accumulators 104a, 104b and 104c are preferably designed as bladder accumulators or membrane accumulators, wherein the
  • Gas when filling the high-pressure hydraulic accumulators 104a, 104b and 104c with the working fluid in the pressurized pressure accumulator 105 can escape.
  • the ratio of the gas pressure of the pressure accumulator 105 between the maximum value and the lowest value preferably has the value 1.1 to 1.25.
  • Diesel, gas, Stirling or other engine may be formed, is a liquid hydraulic pump 103rd
  • This fluid hydraulic pump 103 is in pressure communication between the low-pressure hydraulic accumulator 111 and the high-pressure hydraulic accumulator 104 for filling the smaller high-pressure hydraulic accumulators 104a, 104b and 104c with the working fluid under working pressure against the compensating pressure of the accumulator 105 arranged.
  • At least one hydraulic drive motor 110 in FIG. 3 four hydraulic drive motors 110, is coupled in a torsionally rigid manner to a drive wheel 113 of the motor vehicle. in the
  • This hydraulic drive motor 110 for example, as a hydrostatic high-torque rotary piston engine after the Orbit formed, has a variable in an adjustment area variable displacement per shaft revolution, the specific absorption amount of the hydraulic drive motor 110 per shaft revolution in one for the
  • the drive motor 110 has a high-pressure port 1, which in a forward mode D, as shown in Figures la to lc, with the high-pressure hydraulic accumulator 104, namely with at least one of the smaller high-pressure hydraulic accumulators 104a, 104b and 104c , connected is.
  • Low pressure port 70 of the hydraulic drive motor 110 is connected to the low pressure hydraulic accumulator 111 in the forward mode D.
  • the high-pressure port 1 and the low-pressure port 70 are connected to a mode switching unit 108, which as a
  • Selector slide 108 is formed in the form of a 4/3 way valve connected. By means of this selector slide 108, the controls of the high pressure port 1 and the
  • Low pressure port 70 are interchanged.
  • the selector spool 108 is disposed in pressure communication between the hydraulic drive motor 110 and the high-pressure hydraulic accumulator 104 and between the hydraulic drive motor 110 and the low-pressure hydraulic accumulator 111.
  • the high-pressure port 1 In the forward mode D shown in the figures, the high-pressure port 1 is connected to the high-pressure hydraulic accumulator 104, whereas in a reverse R mode, the high-pressure port 1 communicates with the low-pressure hydraulic accumulator 111 and the low-pressure port 70 with the high pressure port
  • High-pressure hydraulic accumulator 104 is connected. Thus, a reversal of direction is possible.
  • the low-pressure port 70 is also connected to the high-pressure hydraulic accumulator 104, wherein a check valve 122, a suction of the working fluid of the Low-pressure hydraulic accumulator 111 allows the high-pressure port 1, as shown in Figure la recognizable.
  • Rekuperations mode B blocks the 4/3-way valve 108, the local connection between the high-pressure port 1 and the high-pressure hydraulic accumulator 104 or the low-pressure hydraulic accumulator 111. Instead, bridges the
  • a power slide 118 is arranged, by means of which the liquid hydraulic pump 103 can be switched from a non-pressurized circulation to a high-working-pressure performance mode.
  • a power slide 118 is arranged, by means of which the liquid hydraulic pump 103 can be switched from a non-pressurized circulation to a high-working-pressure performance mode.
  • the flow control valve 109 can serve as a throttle valve or preferably as a flow control valve
  • the working fluid flow regulator 109 should be made large enough or by a bypass valve connected in parallel, which is also referred to as a bypass valve (not shown in the figures) to be supplemented.
  • the drive comprises a control device 121 in the form of an electronic control unit. This is with the mode switching unit 108 in such
  • Control connection is designed such that when
  • Recuperation mode B is granted.
  • the detection of the train operation and the overrun operation can be done for example by means of sensors on the brake and accelerator pedal.
  • the activation of the operating mode switching unit 108, that is to say the selector slide, can be done electrically by means of electrical actuators.
  • Recuperation mode B acts in forward-coasting operation of the hydraulic drive motor 110 as one of the
  • the working fluid is pumped from the low pressure hydraulic accumulator 111 via the check valve 122 and the high pressure port 1 to the low pressure port 70 into the high pressure hydraulic accumulator 104.
  • Motor vehicle is thus converted into potential energy in the form of a compression of the working fluid in the high-pressure hydraulic accumulator 104 and can be used at a later time for driving the vehicle in the forward mode D.
  • a recuperative Bremsenergy- use is possible.
  • the hydraulic drive motor 110 has a continuous drive shaft 11, as shown schematically in FIG shown.
  • a filling pump 114 is arranged, which in the direction of at least one
  • the hydraulic drive motor 110 pumps and is driven directly by this.
  • the filling pump 114 is located in
  • the specific swallowing amount of the filling pump 114 is approximately equal to half the maximum specific amount of swallowing of the hydraulic drive motor 110 per shaft revolution.
  • Working fluid flow regulator 109 and the liquid hydraulic pump 103 is a valve assembly 115, 116 and 117 arranged in the form of three 5/4 way valves.
  • This valve arrangement 115, 116 and 117 is between one of
  • High-pressure hydraulic accumulators 104a, 104b, 104c are provided, there are three switching states, which are shown in Figures la, lb and lc. In each of the three switching states, one of the three smaller high-pressure hydraulic accumulators 104a, 104b or 104c is connected exclusively to the liquid hydraulic pump 103 for filling the respective high-pressure hydraulic accumulator, while another one of the three smaller high-pressure hydraulic accumulator 104a, 104b or 104c is connected exclusively to the hydraulic drive motor 110 for emptying the respective high-pressure hydraulic accumulator and thus to the energy output to the drive motor 110.
  • the plurality of smaller high pressure hydraulic accumulators 104a, 104b, and 104c are alternately inflatable and deflatable, each being either decompressed or filled, but not filled and deflated at the same time.
  • the first directional control valve 115 is connected such that the first smaller high-pressure hydraulic accumulator 104 a can be filled by means of the liquid hydraulic pump 103, as illustrated by the rhombic pattern.
  • the third smaller high-pressure hydraulic accumulator 104c is reversed. This is connected to the hydraulic drive motor 110 and drives it, as shown by the horizontal stroke pattern, without being connected to the liquid hydraulic pump 103.
  • the second smaller high pressure hydraulic accumulator 104b is blocked.
  • the second directional control valve 116 is connected in such a way that the second smaller high-pressure hydraulic accumulator 104b can be filled by means of the liquid hydraulic pump 103, as shown by the diamond pattern
  • the hydraulic drive motor 110 is blocked.
  • the first smaller high pressure hydraulic accumulator 104a is connected to the hydraulic drive motor 110 connected and drives this, as shown by the horizontal line pattern, without being connected to the liquid hydraulic pump 103.
  • the third smaller high pressure hydraulic accumulator 104c is neither filled nor emptied.
  • Figure lc shows the opposite case of Figure la.
  • the third directional valve 117 is switched such that the third
  • Liquid hydraulic pump 103 is filled, as also shown by the diamond pattern. However, communication between the third smaller high pressure hydraulic accumulator 104c and the working fluid flow regulator 109 is interrupted.
  • the second smaller high-pressure hydraulic accumulator 104b is connected to and drives the hydraulic drive motor 110, as also shown by the horizontal stroke pattern, without being connected to the liquid hydraulic pump 103.
  • the first smaller high pressure hydraulic accumulator 104a is disconnected.
  • Working fluid in the form of a working fluid is
  • Outlet part 53 is supplied in the housing 19, so that one with the High pressure port 1 connected outer annulus 2 is pressurized by a disk-shaped rotary valve 3, as shown in Figure 4.
  • the disc-shaped rotary valve 3 is centered about a shaft axis 50 of a shaft 11 rotatably mounted and the shaft 11 centrally running.
  • Commutation of the rotary valve 3 flows the working fluid through window 7 in a power section 51st
  • the power unit 51 has a center fixed stator 52 having one
  • the rotary valve 3 ver and disposed of the tooth chambers 6 via axial channels 4 and radial channels 5 in the stator 52 with the working fluid such that the working fluid of the as
  • the rotary valve 3 has as large a size as possible
  • Rotary valve outer diameter 69 as shown in Figures 4, 5 and 10.
  • the rotary valve outer diameter 69 is preferably larger by a factor of 1.3 to 1.5 than the rotor outer diameter 68.
  • the rotary valve 3 is in a known manner by means of a
  • Rotary valve 3 is used, driven.
  • Taumelgetriebes 56 has an eccentric internal teeth 58 with a number of teeth x and an eccentric technicallyverprin 59 with a number of teeth y.
  • the eccentric internal toothing 58 meshes with a second shaft outer toothing 60 of the shaft 11 with a number of teeth w.
  • the eccentric technicallyverprin 59 combs with a fixed internal gear 61 with a number of teeth z.
  • the numbers of teeth w, x, y, z are dimensioned such that the eccentric 57 performs eccentric to the shaft axis 50 between the shaft 11 and the inner ring gear 61 during rotation of the shaft 11 an orbital movement. Possible number of teeth combinations can be found in the above explanations.
  • a pot-shaped connecting part 62 is rotatably coupled to the eccentrically rotatable eccentric 57 at one end. Otherwise it is
  • the shaft 11 is radially mounted on both sides of the power section 51 adjacent radial bearings 63, which are designed as rolling bearings in O arrangement.
  • a rotationally fixed axial second compensating piston 55 which acts axially on the rotary valve 3, arranged around the wobble gear 56.
  • stator internal teeth in the form of Rolls 13 shown, which represents a particularly advantageous embodiment of this tooth shape with serious benefits.
  • the rotary piston 8 is with its internal rotor teeth 9 on the first shaft external teeth 10 relative to the shaft 11 and with its external rotor toothing 12 on the stator inner teeth 13 relative to the stator 52 parallel to
  • Shaft axis 50 axially displaceable.
  • the displacement is made possible by adjusting members 22, which are in engagement with the rotary piston 8 and by means of which selbiger axially
  • the tooth chambers 6 are arranged in the axial direction between an axially fixed rotor plate 17 and an axially movable stator plate 14. They are axially limited and sealed by the rotor plate 17 and the stator plate 14.
  • the rotor plate 17 is axially fixed and radially coupled to the rotary piston 8. Thus, it performs the orbital motion of the rotary piston 8 together with the rotary piston 8.
  • the stator plate 14 is axially movable, axially coupled to the rotary piston 8 and decoupled radially from the rotary piston 8.
  • the volume of the tooth chambers 6 is thus continuously variable by the axial displacement of the rotary piston 8 and the stator plate 14 by means of the adjusting members 22.
  • Stator plate 14 integrally formed. This is
  • Balancing piston 15a is arranged in the stationary housing 19 and can be acted upon by the working pressure of the working fluid of the toothed chambers 6.
  • the compensating piston 15a engages in an axial - in particular annular - compensating cylinder 15b in such a way that the working pressure in the compensating cylinder 15b via the stator plate 14 to the
  • Dental chambers 6 acts. Between the compensating piston 15a and the compensating cylinder 15b act control springs 16 which press the compensating piston 15a axially against the rotary piston 8 and thus the rotary piston 8 in the direction of a starting position with a reduced volume of the tooth chambers 6.
  • Tooth engagement between the rotor outer teeth 12 and the stator inner teeth 13 is particularly bulky, so that accordingly a large displacement per rotor revolution and thus per shaft revolution is generated.
  • Balancing piston 15a forms, moved together with the rotary piston 8 to the right.
  • the control springs 16 are completely compressed and thereby generate the
  • the compensating piston 15a which modifies the absorption volume, is thus arranged with control spring 16 in such a way that, when the working pressure fluctuates, the torque delivered to the drive wheel 113 is automatically kept constant.
  • Figures 4 and 5 is effected by the rotor plate 17.
  • This has a rotor plate internal toothing 18, which corresponds exactly to the rotor outer teeth 12 of the rotary piston 8, wherein a movement of a few hundredths of a millimeter is provided, so that a high degree of tightness between the tooth chambers. 6 and the non-pressurized interior of the machine.
  • This running clearance need not be greater than to ensure that the rotor is in it axially
  • This rotor plate 17 thus has the
  • the rotor plate 17 is thus in a disc-like recess in the housing 19 between a first housing part 19a and a second
  • Housing part 19b mounted radially movable and radially sealing. In the housing 19 is thus one of the orbital movement of
  • Rotor plate 17 corresponding free mobility provided in the radial direction and in the circumferential direction.
  • FIG. 8 shows a cross section through the axial center of this rotor plate 17, into which a groove 20 extends radially from the outside
  • Rotor plate 17 between a housing ring 21 displaced viscous working fluid does not generate excessive crushing losses.
  • the adjusting members 22 are, as shown in Figures 4 and 5, as a circular centric to the
  • Control piston 22 a formed, which is axially in engagement with the rotary piston 8 via a radially extending washer 66, as shown in Figures 4 and 5.
  • the control piston 22a is arranged such that it presses the rotary piston 8 against the compensating piston 15a and against the control springs 16 via the intermediate disc 66.
  • an axial control cylinder 22 b is arranged, which is acted upon by a control pressure and in which the control piston 22 a engages such that by means of the control pressure in the control cylinder 22 b, a force on the rotary piston 8 for axial displacement of the rotary piston 8 can be exercised. Because of this
  • Both the control piston 22a and the control cylinder 22b annularly surround the output section 65 of the shaft 11 in an annular manner.
  • the torque-generating rotary piston 8 transmits its torque via its internal rotor toothing 9 to the first external shaft toothing 10 of the shaft 11, on which the rotary piston 8
  • Balance piston 15 a good leadership relationship. Therefore, it is preferable that the balance piston 15a is fixedly connected to the stator plate 14 to improve the guide ratio. Furthermore, it is desirable to constructively, and the meshing between the rotor inner teeth 9 and the first shaft external teeth 10 as long as possible
  • channels 71 are shown, in which the
  • Balancing pressure field 23 is passed to the rotor plate 17.
  • Valves 24, in particular check valves, ensure that the working pressure is always directed to the right place in different directions of rotation.
  • the rotary piston engine has a control pressure port 67 which is connected to the control cylinder 22b via lines such that the control piston 22a can be acted upon with a control pressure from the control pressure port 67.
  • a separate pressure line with a control pressure connection 67 is provided for the supply of the regulating piston 22a.
  • Rotary piston engines in particular depending on the design of the control springs 16 even substantially independent of Working pressure and torque of the active by varying the control pressure to adjust.
  • Control pressure port 67 increases, increases the force of the control piston 22a and moves the rotary piston 8 against the control springs 16 in the direction of a larger
  • High pressure port 1 supplies, which is driven for example by an internal combustion engine, thereby automatically reduces the output speed of the rotary piston engine and thus the drive speed of the wheels.
  • an electronically actuated magnetic two-edge control valve may be provided for this purpose, by means of which, for example, the
  • Control pressure at the control pressure port 67 is controllable.
  • FIG. 11 shows an alternative embodiment of a volume-adjustable hydrostatic rotary piston machine according to the invention with a circular piston 8, elongated along the shaft axis 50, an extended shaft 11, a longer control piston 22a and a longer piston
  • Control cylinder 22b and a greater distance from
  • Embodiment of the wheel motor according to Figure 4 and Figure 5 this is only about 16% longer.
  • the remaining features correspond to the features of the previous one

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Transportation (AREA)
  • Arrangement Or Mounting Of Propulsion Units For Vehicles (AREA)

Abstract

L'invention concerne un entraînement hydrostatique pour un véhicule automobile, comprenant un accumulateur hydraulique à basse pression (111), un accumulateur hydraulique à haute pression (104) et un accumulateur à pression (105) qui monté en aval de l'accumulateur hydraulique à haute pression (104) pour précontraindre par pression de compensation. Une pompe hydraulique à liquide (103) servant à remplir l'accumulateur hydraulique à haute pression (104) de liquide de travail peut être entraînée au moyen d'un moteur à combustion interne (101). Au moins un moteur d'entraînement hydraulique (110) qui est accouplé avec une roue d'entraînement (113) du véhicule automobile présente un raccord haute pression (1) qui est relié dans un type de fonctionnement en marche avant (D) à l'accumulateur hydraulique à haute pression (104) et un raccord basse pression (70) qui est relié dans le type de fonctionnement en marche avant (D) à l'accumulateur hydraulique à basse pression (111). L'accumulateur hydraulique à haute pression (104) est divisé en plusieurs plus petits accumulateurs hydrauliques à haute pression (104a, 104b, 104c) qui sont couplés ensemble à l'accumulateur à pression (105) en aval et qui peuvent être reliés pratiquement indépendamment les uns des autres audit moteur d'entraînement hydraulique (110) et peuvent être vidés alternativement. Ledit moteur d'entraînement hydraulique (110) présente une cylindrée modifiable dans une région de réglage par rotation d'arbre, la quantité balayée spécifique par rotation d'arbre pouvant être modifiée dans une région suffisante pour le fonctionnement du véhicule.
PCT/EP2011/065091 2010-09-06 2011-09-01 Entraînement hydrostatique pour un véhicule automobile WO2012031970A1 (fr)

Priority Applications (1)

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EP11749441.9A EP2614274A1 (fr) 2010-09-06 2011-09-01 Entraînement hydrostatique pour un véhicule automobile

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EP10175421 2010-09-06
EP10175421.6 2010-09-06

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Cited By (9)

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CN102642459A (zh) * 2012-05-17 2012-08-22 张志新 汽车液电气混合动力系统
WO2013167610A1 (fr) * 2012-05-10 2013-11-14 Dieffenbacher GmbH Maschinen- und Anlagenbau Procede et dispositif de commande adaptative d'une presse hydraulique
WO2014166957A1 (fr) * 2013-04-09 2014-10-16 Cnh Industrial Belgium Nv Système d'entraînement hybride pour une moissonneuse et procédé de fonctionnement de ce dernier
JP2016094149A (ja) * 2014-11-17 2016-05-26 マツダ株式会社 車両の回生制御方法及び回生制御システム
CN107187323A (zh) * 2017-07-07 2017-09-22 沙尔夫矿山机械（徐州）有限公司 一种超级电容齿轨卡轨景观小火车主机
US10377220B2 (en) 2014-02-28 2019-08-13 Dana Italia Spa Dual mode hybrid hydrostatic driveline
FR3079012A1 (fr) * 2018-03-13 2019-09-20 Ali Bal Systeme de motorisation hydraulique par pression sans moteur a combustion pour vehicule automobile
DE102021209462A1 (de) 2021-08-30 2023-03-02 Robert Bosch Gesellschaft mit beschränkter Haftung Sekundärgeregeltes hydraulisches Druckmittelversorgungssystem mit einem offenen hydraulischen Kreis
DE102022124423A1 (de) 2022-09-22 2024-03-28 Wacker Neuson Linz Gmbh Vorrichtung und Verfahren zum Verzögern eines Fahrzeugs

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Cited By (11)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2013167610A1 (fr) * 2012-05-10 2013-11-14 Dieffenbacher GmbH Maschinen- und Anlagenbau Procede et dispositif de commande adaptative d'une presse hydraulique
CN102642459A (zh) * 2012-05-17 2012-08-22 张志新 汽车液电气混合动力系统
WO2014166957A1 (fr) * 2013-04-09 2014-10-16 Cnh Industrial Belgium Nv Système d'entraînement hybride pour une moissonneuse et procédé de fonctionnement de ce dernier
BE1021140B1 (nl) * 2013-04-09 2016-01-08 Cnh Industrial Belgium Nv Een hybride aandrijfsysteem voor een oogstmachine
US10039234B2 (en) 2013-04-09 2018-08-07 Cnh Industrial America Llc Hybrid drive system for a harvester
US10377220B2 (en) 2014-02-28 2019-08-13 Dana Italia Spa Dual mode hybrid hydrostatic driveline
JP2016094149A (ja) * 2014-11-17 2016-05-26 マツダ株式会社 車両の回生制御方法及び回生制御システム
CN107187323A (zh) * 2017-07-07 2017-09-22 沙尔夫矿山机械（徐州）有限公司 一种超级电容齿轨卡轨景观小火车主机
FR3079012A1 (fr) * 2018-03-13 2019-09-20 Ali Bal Systeme de motorisation hydraulique par pression sans moteur a combustion pour vehicule automobile
DE102021209462A1 (de) 2021-08-30 2023-03-02 Robert Bosch Gesellschaft mit beschränkter Haftung Sekundärgeregeltes hydraulisches Druckmittelversorgungssystem mit einem offenen hydraulischen Kreis
DE102022124423A1 (de) 2022-09-22 2024-03-28 Wacker Neuson Linz Gmbh Vorrichtung und Verfahren zum Verzögern eines Fahrzeugs

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