EP2614274A1

EP2614274A1 - Hydrostatic drive for a motor vehicle

Info

Publication number: EP2614274A1
Application number: EP11749441.9A
Authority: EP
Inventors: Siegfried A. Eisenmann
Original assignee: Individual
Current assignee: Individual
Priority date: 2010-09-06
Filing date: 2011-09-01
Publication date: 2013-07-17
Also published as: WO2012031970A1

Abstract

The invention relates to a hydrostatic drive for a motor vehicle, with a low-pressure hydraulic accumulator (111), a high-pressure hydraulic accumulator (104) and a pressure accumulator (105) which is connected downstream of the high-pressure hydraulic accumulator (104) for pretensioning with equalizing pressure. A fluid hydraulic pump (103) for filling the high-pressure hydraulic accumulator (104) with working fluid can be driven by means of an internal combustion engine (101). At least one hydraulic drive motor (110) which is coupled to a drive wheel (113) of the motor vehicle has a high-pressure connection (1), which is connected in a forwards operating mode (D) to the high-pressure hydraulic accumulator (104), and a low-pressure connection (70), which is connected in the forwards operating mode (D) to the low-pressure hydraulic accumulator (111). The high-pressure hydraulic accumulator (104) is divided into a plurality of smaller high-pressure hydraulic accumulators (104a, 104b, 104c) which are jointly coupled to the pressure accumulator (105) connected downstream and which can be connected substantially independently of one another to the at least one hydraulic drive motor (110) and can be emptied in an alternating manner. The at least one hydraulic drive motor (110) has an absorption volume, which is variable within an adjustment range, per shaft revolution, wherein the specific absorption quantity per shaft revolution is changeable within a range sufficient for the operation of a vehicle.

Description

HYDROSTATIC DRIVE FOR E IN MOTOR VEHICLE

The invention relates to a hydrostatic drive for a motor vehicle, in particular a passenger car, according to the preamble of claim 1.

In today's traffic are due to the sometimes high traffic volume and speed limits as well as highly dominant inner-city driving cycles the

Performance claims to those installed in the vehicle

Motorization so small that the engine in large

Time shares in the extreme partial load range works. This has the consequence that the operating point in the engine map in this part-load operation far away from his

Consumption optimum is located.

In many cases, the engine works at only about one

Sixth of its possible full load torque. Of the

However, the operating point of the lowest specific fuel consumption, measured in grams per KW per hour, is in the immediate vicinity of the highest possible torque. Modern low-turbocharged gasoline engines currently have the lowest specific fuel consumption

Range of be = 230 grams per KW and hour. In the above

However, in the partial load area mentioned above, consumption is within

Range of double, namely at about 450 grams per KW and hour.

Not so dramatically the circumstances are

Diesel engine. With turbocharger turbocharger you can reach best values of currently be = 190 grams per KW and hour and below. Although the above-mentioned partial load consumption is much better here, it is still around 270 grams per KW per hour in the optimum speed range. In a newly developed 3- Except for the highest torque at an optimum speed in the range of n = 2000 rpm, the cylinder diesel engine from Volkswagen has the optimum. To maintain the lowest fuel consumption in the engine map, it is sufficient not only to maintain the highest possible torque during operation of the engine, but also the optimum engine speed.

In order to meet these demands reasonably well, the number of gears has been dramatically increased in the last ten years, with up to 8 gears in automatic transmissions and up to 6 gears in the case of the car gears

Manual transmissions. Accordingly, the so-called translation spread has become larger. Usually the two highest gears are translated as fast

executed. 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. Of the

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

Torque behavior "lame" and unsportsmanlike.

In this conventional drive system is still a mechanically rigid connection between the engine and the drive system, so that the partial load operation of the engine can not be largely avoided. In particular, in the vehicles with manual transmissions, the profitability of the vehicle depends on the behavior of the driver. The driver can operate the engine in the optimal characteristic field in the least of cases for lack of exact knowledge. Therefore, the drive system must be designed so that this is achieved by an automatism. To solve this problem, a complete decoupling of the drive system from the engine is advantageous, the engine is always automatically operated in the consumption optimum, ie at full torque and in the associated optimal

Rotation range. The engine works at a given speed

High power RPM at point or range of

lowest specific fuel consumption. Since today many vehicles with a start-stop automatic, also called stop-and-go automatic, are equipped for automatically deactivating and reactivating the engine when the vehicle is at a standstill, it is no problem to connect the engine only temporarily

operate, always at full load. In the intervening time, the engine is shut down or kept at low idle. To further increase the efficiency of a so-called integrated starter generator, also called ISG, are used, at the same time as a generator for the

On-board electrical system and as an alternative to starting the engine is used.

It is also advantageous to switch to a higher

On-board voltage, for example 42 volts. This will be the

Starting power increased while reducing the current.

For decoupling one only in interval mode

running internal combustion engine from the drive system with

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.

An electrical coupling of the internal combustion engine with the drive train has the particular disadvantage that the known from the prior art accumulators or

Capacitors not in the shortest possible time with the

required energy can be charged and discharged. The weight and space requirements of electric motors, the Generator, the accumulators or capacitors, the

Electronics and wiring make a mechanical

Decoupling a powerful combustion engine from the drive train by an exclusively electrical coupling in practice currently unattractive.

An exclusively hydraulic coupling of the

Interval operation of running internal combustion engine with the

Powertrain with pressure storage in turn has the

Disadvantage that with stationary internal combustion engine and

Operation of the hydraulic motors with the pressure in the accumulator

stored energy of the operating pressure is not stable at a consistently high level and drops until the complete emptying of the accumulator in particular continuously. Due to this pressure drop, the torque drops to the

Hydraulic motors. Thus, the performance of the vehicle also decreases. One solution would be a general pressure limit, which in turn would be the maximum available

Would reduce torque from the outset. Thus, the mechanical decoupling means only hydraulic coupling proves to be limited in practice

can be used.

A similar objective is pursued by a recently published proposal for a hydraulic drive for passenger cars in the journal ATZ - Automobiltechnische Zeitschrift

(Issue 05 / May 2009, pages 378 ff.), In which a so

said pressure transformer is used, the

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

Increase in pressure with constant-speed engines only about the torque by a factor of 2.5, are such high pressures, especially in a hose connection of the driven steered wheel in the four-wheel drive, extremely problematic.

The object of the invention is therefore one of a

Intermittently operable internal combustion engine driven by this decoupled and with a

Energy buffer equipped hydrostatic

To provide drive for a motor vehicle, which is characterized by a high torque spread on the drive wheel de vehicle and during unloading of the

Intermediate energy storage essentially constant, high maximum torque.

This object is achieved by the realization of the features of the independent claim. Features which further develop the invention in an alternative or advantageous manner can be found in the dependent claims.

The inventive hydrostatic drive for 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. 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. In other words, the pressure accumulator is used when filling the high-pressure hydraulic accumulator with

Working fluid in the high-pressure hydraulic accumulator displaced gas is compressed in the accumulator. At the

Filling the high-pressure hydraulic accumulator with

Working fluid is so selbige against the gas of the

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

Combustion engine coupled so that they of the

Internal combustion engine is drivable and selbiger thus as

Pump drive is used. 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

Accumulator arranged. The pressure connection can via a line system, in particular an arrangement of

Hydraulic hoses and valves to be created. In other words, 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

Internal combustion engine is applied. The pressure in the high-pressure hydraulic accumulator during normal operation is always higher than the pressure in the low-pressure hydraulic accumulator. 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. Preferably, 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.

Alternatively, however, there is the possibility that between the hydraulic drive motor and the drive wheel, a transmission, in particular Reduzier planetary gear, and / or a

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. During engine operation of the hydraulic drive motor, the high pressure port provides the pressure input and the

Low pressure port is the pressure output. In 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

hydraulic drive motor and thus also the drive wheel at flows of the working fluid from the high pressure in the

Low-pressure hydraulic accumulator powered. In other words, 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

In the form of a guide in the pressure connection between the

High pressure hydraulic accumulator and hydraulic

Drive motor a working fluid flow regulator,

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. Preferably, the working fluid flow regulator in the extreme case, a sufficiently large flow rate or has a parallel bypass valve.

Conventional flow control valves are lossy, because the flow of electricity by means of a pressure balance and a

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.

It is better, for example, 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

Turning speed of the steering wheel, internally with 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

Lossless emptying speed control from the

Hydrospeicher and thus a speed control of the vehicle possible without the amount of swallowing the

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

takes up. The invention therefore provides in an embodiment that the working fluid flow regulator is used as a metering pump

is trained.

Alternatively, however, theoretically it is also possible to dispense with the use of a working fluid flow regulator or a metering pump between the high-pressure hydraulic accumulator and the hydraulic drive motor, so that there is a direct pressure connection between the high-pressure hydraulic accumulator and the hydraulic drive motor , wherein the vehicle drive, for example by

Adjustment of the drive motor and / or can be controlled by means of a brake.

According to the invention, the high-pressure hydraulic accumulator is divided into a plurality of smaller high-pressure hydraulic accumulators, which are coupled together with the downstream accumulator. In other words, the downstream one spans

Accumulator before the several smaller high-pressure hydraulic accumulator together before. These are connected to the at least one hydraulic drive motor such that the plurality of smaller high-pressure hydraulic accumulators in the

Essentially independent of each other with the at least one hydraulic drive motor can be connected and alternately emptied. In other words, 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

preferably always operated only with the working fluid from one of the several smaller high-pressure hydraulic accumulator. The other smaller high-pressure hydraulic accumulators are now either filled with pressurized fluid by the hydraulic fluid pump or remain closed. 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

Drive motor per shaft revolution in one for the

Vehicle operation sufficient range stepless or

graded is changeable. In a further development of

Invention is the hydraulic drive motor as

hydrostatic high torque rotary piston engine after the

Orbit formed, which has a variable in an adjustment range continuously variable displacement per shaft revolution. Such a hydrostatic Hochmoment- rotary piston engine is known from European patent application EP 10166995.0 entitled "Infinitely variable volume hydrostatic rotary piston engine", registered on

23.06.2010, which is hereby incorporated by reference, the same applicant known. The hydrostatic described there

Rotary piston engine will be described in more detail below.

Instead of at least one infinitely adjustable

Hydraulic drive motor with continuous displacement per shaft revolution can according to the invention also at least one gradual adjustable hydraulic drive motor for

Use come. Such, graduated or adjustable in discrete stages hydraulic motors whose displacement in several stages, especially narrowly graduated in a variety of levels are adjustable, are in different

Designs known from the prior art. 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 zusammenschliessbar, so that in particular the number of switching states corresponding number of intake volumes and thus gear ratios results.

In the following, for the sake of simplicity, 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

adjustable motor is understood.

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.

In one embodiment of the invention is a

Valve arrangement, in particular a plurality of directional control valves

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. By means of the valve arrangement can be between one of the number of several smaller

High-pressure hydraulic accumulator corresponding number

Switching states are switched,

wherein in each switching state, one of the plurality of smaller high-pressure hydraulic accumulator exclusively with the

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.

Thus, 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

Driving the drive motor are emptied.

A development of the invention provides that the

hydrostatic drive includes a mode switching unit, for example, as a selector slide,

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. By means of the mode switching unit are the connections between the above forward mode and a

Recuperation mode switchable. In the recuperation mode, the high pressure port with the low pressure hydraulic accumulator and the low pressure port with the

High-pressure hydraulic accumulator connected. In other words, the connections to the hydraulic drive motor in the recuperation mode are directly or indirectly reversed. Thus, 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.

In order to enable a Rekuperationsbetrieb, by means of which in the overrun mode of the vehicle and in particular during braking, a part of the kinetic energy of the vehicle is converted into a potential energy which is stored in the form of working pressure in the high-pressure hydraulic accumulator, the invention also provides a especially

electrical, hydraulic, mechanical or pneumatic control device which is in control connection with the mode changeover unit. These

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

Delaying the motor vehicle in the forward overrun mode, in particular during braking, 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

is adjusted. The braking power is in particular by means of a brake pedal sensor and a flow control valve

dosed. In the recuperation mode in forward thrust mode, 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

preferably supplied exclusively with the working fluid from the high-pressure hydraulic accumulator. In other words, the hydraulic, in their swallowing amount per

Engine revolution stepless or graduated changeable

Wheel motors fed exclusively from the high pressure hydraulic accumulator with working fluid. Thus, the liquid hydraulic pump and the at least one hydraulic

Drive motor via the acting as a buffer memory

High-pressure hydraulic accumulator decoupled.

Preferably, the at least one hydraulic

Drive motor to a continuous drive shaft, wherein on the continuous drive shaft of the at least one hydraulic drive motor, 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. By means of this

In the recuperation mode in the recuperation mode, in which the hydraulic drive motor acts as a pump, the filling pump is prevented from being located in the intake line between the low-pressure hydraulic accumulator and the hydraulic pump

Drive motor too much negative pressure at the high-pressure connection is formed. Such a negative pressure increases the

Risk of cavitation in the drive motor considerably. By means of the filling pump, it is thus possible, the braking power of

hydraulic drive motor in recuperation mode

significantly increase. The specific amount of swallowing

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

Shaft rotation. Preferably, the specific one is

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

Amount 1.1 to 1.25 on.

In one embodiment of the invention is in one

Pressure connection between the hydraulic fluid pump and the high pressure hydraulic accumulator a power slide for reversing the hydraulic fluid pump from a

depressurized circulation in a high performance mode

Working pressure switched. A particular advantage of the invention is recuperative

Braking energy usage, which is the overall efficiency of the

can improve hydrostatic vehicle drive according to the invention. In this case, for example, operates the at least one hydrostatic wheel motors as a pump. He must therefore be able to suck, especially at high speed, cavitation free the full volumetric oil quantity. Both

Preferably used for coming controllable orbit engines, a commutator is necessary, the system has only limited passage cross-sections. With still so big

dimensioned masses, it may be necessary that 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

Pumps working wheel motors from the slightly biasing pressure low pressure hydraulic accumulator can suck u. U. be dispensed with such a filling pump.

For the operation of hydraulic drive motors as

Wheel motors, it is desirable that the operating pressure is kept substantially constant to pulsating

To avoid fluctuations in the drive torque. in the

Storage operation but would be an absolute constant storage oil pressure with downstream gas storage only possible if the gas storage is a very large volume

having. The usable oil-high pressure volume is calculated according to the formula dV = Vo [Po / Pl - Po / P2] [1], where Vo is the gas volume and dV is the useful volume. The quotient Po / Pl can be assumed to be about 1.0 at the pressures used here, then Pl = Po. As can be seen from this formula [1], when filling the oil reservoir with almost constant pressure, ie P2 = PI, the gas volume Vo would have to be extremely large if one wanted to achieve a usable useful volume dV for the hydraulic drive. So you have to accept a pressure difference dP = P2 - PI. When changing the formula [1] becomes

Vo = dV / (Po / Pl - Po / P2) = dV / (1 - P1 / P2) [2], if Po = Pl is set. From this one can now the necessary gas volume Vo at a given dV and the selected

Calculate pressure ratio P1 / P2. However, in the vehicle there is only a limited space available for all the required storage, namely high pressure hydraulic feeders, low pressure hydraulic feeders and accumulators, e.g. 100 liters. This creates a complex problem:

1. Does the combustion engine drive a hydraulic pump?

constant volume, then fluctuates

Torque around the pressure fluctuation of the system. This should only be so large that the operation is largely in the

Optimum range of specific fuel consumption remains.

2. The pressure fluctuation determines the torque fluctuation of the wheel motors in the drive wheels.

3. 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.

Example: a) Assuming a permissible pressure fluctuation of 16%, P1 / P2 = 1 / 1.16 = 0.86. If a usable oil storage volume of dV = 27 liters is required, this would be without splitting the high-pressure hydraulic accumulator

Vo = 27 / (1 - 0.86) = 193 liters.

The gas storage then needed 193 - 27 = 166 liters. b) However, if the high pressure hydraulic accumulator is disconnected e.g. in two smaller memory, each with 13.5 liters of content, which are connected in particular via rotary valve serially, then

Vo = 13, 5 / (1 - 0, 86) = 96, 4 liters.

The gas storage then requires 96.4 - 13.5 = 83 liters.

If you split the high-pressure hydraulic accumulator into three smaller accumulators, which are connected serially, in particular via a rotary valve, then it becomes

Vo = 9 / (1 - 0, 86) = 64 liters The gas storage then only needs a volume of 64 - 2x9 = 46 liters.

In this case, 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

Driving cycle. Of course, then the liquid hydraulic pump and thus the internal combustion engine must be switched on and off 3 times per driving cycle. Since the memory content is very small each time, it is affected by the liquid

Hydraulic pump charged at 2000 U / min in just a few seconds with high pressure. This shows the considerable advantage of the high-pressure hydraulics due to their high

Power density compared to electrical engineering.

Working pressures between 200 and 300 bar can be expected since hydraulic pumps and hydraulic motors available on the market can operate reliably in this area. The

Refill time of the memory becomes extremely short.

The efficiency of the system is calculated from the

Loss energies of those involved in the process

Individual components. Example at V = 85 km / h: 1. Filling the memory:

Filling time 6 sec;

Pump power 35 KW;

n = 2000 rpm;

Pm = 240 bar;

Pump efficiency dead = 95% (axial piston pump with q = const);

Loss energy 10500 Ws;

Line loss up to the tank: dP = 1 bar;

Oil flow = 90 1 / min; Loss energy 882 Ws;

Loss when filling the accumulator dP = 1.2 bar;

Oil flow = 90 1 / min;

Loss energy 1059 Ws.

2. Emptying the memory:

Emptying time 45 sec;

Loss when emptying the accumulator dP = 1.2 bar;

Oil flow = 12 1 / min;

Loss energy 1059 Ws;

Line loss up to the hydraulic motor dP = 1 bar;

Oil flow = 12 1 / min;

Loss energy 882 Ws;

Losses of the hydraulic motor at efficiency tot = 90%; Engine power 5 KW;

Loss energy 22500 Ws;

Total Energy Loss 36823 Ws.

Total efficiency (35000x6 - 36823) / 35000x6 = 0.8244

As can be seen from this calculation, it is important that the oil viscosity must be as low as possible (HP oil, warm oil) and that the line cross sections and the valves be made as large as possible. Only then are these line losses realistic.

For the thermostability of the drive system, the

Loss of heat are cooled away by the Oldachsatz. This is relatively large and flows through a plethora of pipes, hoses, governors and the oil cooler with fairly large surfaces. The oil cooler can be flowed through by the cooling water of the engine. An example calculation has the following result: Vehicle weight 700 kg (city car);

Driving speed 125 km / h;

Pro ecting area 1.85 square meters;

Drag coefficient Cw = 0.24;

Drive through two hydrostatic wheel motors;

Suction volume 2x14 = 28 cc / rev;

Wheel speed 1400 rpm;

average oil pressure 250 bar;

Oil throughput 2352 liters per hour;

Overall efficiency of the hydrostatic drive 0.824;

Oil heating 3.7 degrees Celsius, is cooled by

Convection in the wind and by the engine cooling.

Fuel consumption: at V = 125 km / h at Pan = 15 KW, be = 195 g / KWh;

Be = 3.298 liters / 100 km diesel oil; at V = 85 km / h at Pan = 5 KW, be = 195g / KWh;

Be = 1.618 liters / 100 km diesel oil

With brake energy recovery, these numbers are still below.

As already mentioned, the high-pressure hydraulic accumulator is divided into three smaller units and designed in particular as a bladder accumulator. They become high pressure side

in particular connected serially by a pump rotary valve or preferably by electrical directional control valves with the hydraulic pump. Drive side, the same memory

in particular via a rotary valve motor or preferably via electrical valves of the same type connected in series with the hydraulic drive motors. Both rotary valves or directional control valves are preferably mechanically coupled. But it is also conceivable that the quasi-hydraulic

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

relevant hydraulic market available.

Because of the unavoidable pressure fluctuation of the high-pressure hydraulic accumulator it makes sense that one between

Motor slide valve or the Motorwegventile and the flow control valve, a pressure reducing valve with set

Working pressure for the hydraulic motors switched. When the set pressure is exceeded, part of the volume flow flows into the low-pressure hydraulic accumulator. But it is also conceivable that the excess flow rate flows through check valves in those memory whose pressure is currently on

lowest is. The loss then becomes smaller. Of the

Working pressure on the hydraulic motors then remains a long way

Stretching constantly. A "bucking" of the vehicle is then avoided. Of course, there are other means for damping pressure fluctuations, such as Damper diaphragm accumulator or spring damper cylinder.

The torque of a hydraulic motor is calculated according to the formula:

T = qxP / 20xPi [3], if q is used in ccm / shaft revolution and P in bar. The result is then a size in Nm. To the torque T to keep it constant for smooth driving, the product of the amount of swallowing times the working pressure must therefore remain the same. Since the high-pressure hydraulic accumulator does not empty suddenly, but during a certain

Vehicle speed-dependent period of time, it is no problem, according to the invention by means of a two-edge electronic control continuously variable in their swirl q per shaft revolution wheel motors constantly adapt to the changing working pressure such that the

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

specific amount of swallowing according to the pressure drop for

Constant holding of the engine torque increased accordingly. Of course, in order to maintain the vehicle speed, the oil flow rate will have to increase, due to the

High-pressure accumulator is automatically ensured by a slightly faster emptying. In other words that is

Torque of the at least one hydraulic drive motor at waste of the working pressure in the high-pressure hydraulic accumulator by the variability of the swallowing volume substantially constant durable.

In this way, no additional arises

hydrostatic loss. As can be seen from this fact, the realization of the inventive decoupling of the drive system of a vehicle from

Internal combustion engine by means of high-pressure hydraulic accumulator only economically possible with infinitely variable hydraulic motors. At the same time, the hydraulic motors must be so-called high-torque motors, so that the dimensions and weight are small are because they add to the unsprung mass of the wheel

have to be beaten.

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.

Of course, there must be the possibility, in extreme driving conditions, the maximum performance of the

To be able to call up the internal combustion engine. In this case, the operating point of the engine is usually removed from the point of lowest specific fuel consumption. in the

hydrostatic drive system, this means that the

Emptying of the hydraulic accumulator must be done much faster while increasing the amount of swirl of the variable displacement or the hydraulic motors on e.g. Maximum value for the torque increase at the drive wheels. The oil throughput through the hydraulic motors then reaches a maximum

correspondingly high hydrostatic power. Of the

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

Internal combustion engine can be serious in the vehicle

Comfort deficits lead, so that such vehicles could be unpopular. This problem can be solved by using so-called rotary piston engines (WANKEL engines) instead of reciprocating piston engines. These engines have excellent NVH (NOISE-VIBRATION-HARSHNESS) properties. They are currently being used as RANGEEXTENDER generator sets for cyclically charging the batteries of electric cars

favored. 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.

Of course, there is also the possibility to switch the engine from the low idle or by start-stop automatic in the power mode, which would be a little more economical. With the decoupled from the engine, 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

To make passenger car drive again economically usable, which draws its energy in a simple way from the gasoline or hydrogen tank. When operating the WANKEL engine with hydrogen, this proves to be particularly advantageous over the reciprocating engine, because more time is available for the mixing of the hydrogen with the atmospheric oxygen in the 4-stroke process. In particular, the hydraulic drive motor, so the hydraulic motor, designed as a hydrostatic high-torque rotary piston engine according to the orbit, the one in a

Adjustment has continuously variable displacement per shaft revolution. Preferably, the comes

High-torque hydrostatic rotary piston engine disclosed in European patent application EP 10166995.0 entitled "Continuously variable volume hydrostatic

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

Suction volume per shaft revolution has to be used.

In the following, these are preferably used

hydrostatic high-torque rotary piston engine more accurate

described.

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

Performs rotary motion about its own axis and this axis makes a circular motion in the opposite direction of rotation than the rotor. This means that in this overall movement, the axial distance between the rotor and stator center points performs a rotational movement in the opposite sense of rotation as the rotor about its own axis. Such a rotary engine according to the orbit principle is thus with respect to

Principle of movement fundamentally different than a so-called ring gear pump, as you for example from the

WO 2009/092719 is known. There is already the 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.

The 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

Characteristic features are explained.

The hydrostatic rotary piston engine, also called

In particular, 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

is arranged. 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

Number of teeth d. 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

Number of teeth a, b, c, d in such a ratio

to each other, that form with working fluid ver and disposable tooth chambers between the stator inner teeth of the stator and the rotor outer teeth of the rotary piston.

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.

By means of a disc-shaped rotary valve, which is mounted centrically running to the shaft and the stator, the

Control of the supply and disposal of the dental chambers with the working fluid. In other words, the rotary valve is rotatably mounted about the central, longitudinally extending shaft of the geometric shaft axis. By means of the rotary valve, the supply and disposal of the tooth chambers with the working fluid by rotation of the rotary valve in such a way over substantially radially arranged in the rotary valve pressure window

Kommutierend controllbar that the working fluid of the

High pressure connection in a first part of the tooth chambers with the working pressure into and from a second part of the

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.

In addition, the rotary engine comprises a gear transmission in the form of a wobble gear

Exzentergetriebes, that between a second

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. Preferably, the wobble gear

exclusively in the leakage oil area of the hydrostatic

Rotary piston engine, so in non-working pressure leading region of the rotary engine, which also with the

Low pressure port is connected, arranged.

The shaft is mounted on both sides of the power unit immediately adjacent radial bearings, in particular bearings.

Due to the mentioned construction of the hydrostatic

Rotary piston engine, a continuous shaft with large shaft diameters and high torsional strength can be used. Thus, it is possible to expose both shaft ends to a high torque flux and, for example, both

To use 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

Drive unit can be made compact.

Since the rotary valve itself does not orbit movement, but rotates about the shaft axis, due to the

Placement of the eccentric gear in the oil leakage space of the engine and by the use of low-cost extruded or Sintered parts as gear members thus creates a compact and inexpensive construction.

Serving as a synchronous drive for the rotary valve

Taumelgetriebe has an eccentric, which has an eccentric internal teeth with a number of teeth x and an eccentric Außenverzahlung with a number of teeth y. The eccentric internal teeth partially mesh with a second

Shaft external toothing of the shaft with a number of teeth w. The eccentric Außenverzahlung combs partly with a

fixed inner ring gear with a number of teeth z. The ratio of the numbers of teeth w, x, y, z is dimensioned such that the eccentric performs eccentric to the shaft axis between the shaft and the internal ring gear during rotation of the shaft an orbital movement. 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.

Are each similar teeth engaged, 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. In particular, the following, known from the prior art

Equation represents the speed ratio shaft to the rotary piston or shaft to the rotary valve: X

^■ z - y

a w

As can be seen from this equation, the numbers of teeth of the wobble mechanism can be performed quite differently.

A first option is, for example, the design exactly as in the power section, in particular with w = 12, x = 14, y = ll and z = 12. It should only be noted that the eccentricity of the two internal gears are exactly the same. The equation expression is a positive integer, preferably equal to 3. It must also be sought that in this

Range of diameter of the shaft is sufficiently large, so that their torsional strength for the maximum torque is still sufficient. Here, however, the eccentricity of

Taumelgetriebes relatively large, so that the

Tumble angle is correspondingly large. However, then the rotational speed of the eccentricity is relatively small.

The relationship between the rotational speed Ne of the eccentricity of the wobble mechanism and the rotational speed Nw of the shaft is given by the equation

Ne wy

Nw x - z - w ^■ y

this ratio is preferably between -3 and -9.

A second option is the preferred interpretation of

Number of teeth after a = 12, b = 14, c = ll, d = 12, w = 12, x = 13, y = 23 and z = 24 or after a = 12, b = 14, c = ll, d = 12 , w = 9, x = 10, y = 17 and z = 18, each with a very small eccentricity. As can be seen from the above equation Ne / Nw, the increases

Speed of eccentricity in this case. In the design of the wobble mechanism with the numbers of teeth a = 12, b = 14, c = ll, d = 12, w = 12, x = 13, y = 23 and z = 24, there are known advantages: As in the assembly of the engine the rotational position of the rotary valve must always exactly match the rotational position of the motor in the power unit in the phase position, it makes sense that the number of teeth w and their rotational position on the shaft is exactly the same as the number of teeth a

External toothing on the shaft on the power unit and its rotational position. Thus, the shaft can always be mounted without having to pay attention to what rotational position it is, whereby the assembly is considerably simplified.

The proposed numbers of teeth a = 12, b = 14, c = ll, d = 12, w = 9, x = 10, y = 17 and z = 18 have respect to the teeth for the

Taumelgetriebe the advantage that the toothing module is greater, the stability of the shaft grows in this area and in particular the negative-running speed of the eccentric axis of the eccentric disc drops sharply, resulting in a smoother running of the transmission. It is thereby accepted that the wobble angle is slightly larger, and dispenses with the above-described advantage during assembly.

In an advantageous embodiment, 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

mechanical-hydraulic efficiency can be significantly increased.

To seal the disk-shaped rotary valve axially against leakage of the working fluid and with its pressure windows with sufficient pressure against the control plate with their To push windows, which lead through channels to the tooth chambers, 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. In particular, 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.

When using such a hydrostatic

High-torque motor as a wheel motor, the output-side radial bearing needs a higher radial load rating for additional recording of the wheel load. Preferably, 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

optionally outside the leakage space of the rotary piston engine with a permanent rolling bearing grease is arranged directly in the housing part of the rotary piston engine. The described

Rotary engine is suitable due to the advantageous

Bearing arrangement and the powerful, optionally continuous shaft, among other things outstanding as a wheel motor for direct driving a wheel. In this case, 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

Shaft axis axially displaceable. For this purpose, 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. In order to move the rotary piston axially, 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

in particular hydraulic, mechanical, pneumatic or electric drives, inside or outside of the

Housing the rotary piston engine can be arranged. In particular, the adjusting members between the rotary piston and the rolling bearings are left and right axially displaceable.

By means of the axial displacement of the rotary piston, the axial longitudinal extent of the tooth chambers and thus the volume of the tooth chambers is adjusted continuously. For this purpose, a radially ersteckende first wall of the tooth chambers is axial

fixed and axially coupled to the housing. By contrast, 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. Thus, 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

Mobility in the radial direction and in the circumferential direction. The stator plate, however, 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.

Preferably, 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 Verzahnungsgeometrien 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.

In one embodiment, 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

Direction facing wall of the rotor plate, which the

Orbit movement, slides.

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. For this purpose, the stator plate external teeth and the stator internal teeth have the same number of teeth d and essentially the same

Tooth geometry. Thus, the stator plate is axially displaceable relative to the axially fixed stator together with the rotary piston. Due to the fit of the

Stator plate external teeth with the stator internal teeth are both in axially sealing, but axially movable contact.

In a further development is a balance piston with the

Stator plate coupled axially. It is possible the

Stator plate and the balance piston integrally form. 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

Starting position with a reduced volume of

Dental chambers acts. Thus, on the one side of the axially displaceable stator plate, a first force resulting from the working pressure in the first part of the tooth chambers and on the opposite other side of the stator plate a second force resulting from the working pressure in the

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.

In order to support this compensation of forces and / or to force a starting position of the rotary piston in the pressureless state of the rotary piston engine, provides a development between the balance piston and the balance cylinder acting control springs, the axial piston against the rotary piston and thus the rotary piston in the direction of a starting position with a Press the reduced volume of the tooth chambers. The control springs thus act in the same direction as the working pressure in the

Compensating cylinder.

The balance piston and the balance cylinder have

for example, a ring shape and in particular extend around the shaft. Preferably, they are between the

Power part and the wobble gear arranged.

Balance piston and balance cylinder can be reversed. Under a balance piston is thus also a

To understand compensation cylinder, and vice versa. In a further development it is provided that the

Adjusting members are designed as a control piston which is axially engaged with the rotary piston. In other words, the control piston with the rotary piston is so axial

coupled, that the rotary piston by means of the control piston is axially displaceable. The 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. For this purpose, the control cylinder can be acted upon by a control pressure. Preferably, the

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

Volume of the tooth chambers with increasing control pressure

increased. In other words, an increasing causes

Control pressure an increase in the volume of the dental chambers.

The 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

Section of the shaft on which the wobble gear is arranged opposite. In order to transmit the axial force from the control piston to the circular cylinder, for example, 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

Direction facing end face on a facing in the opposite axial direction surface of the washer.

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. In other words, 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. For this purpose, there is in particular such a direct or indirect connection between the high pressure port and the control cylinder, that the control piston can be acted upon by the working pressure of the working fluid. In this case, 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

Torque on the driven shaft and thus with increasing working pressure the rotary piston in the direction of

increased volume of the tooth chambers presses. Conversely, a decrease in the volume of the tooth chambers is caused by a decrease in torque and drop in working pressure. Thus, the intake volume increases with increasing torque and decreases with decreasing torque. This makes it possible, high

Speeds at a relatively low

To achieve output torque, low displacement and relatively low working fluid throughput, for example, when driving faster with a vehicle, whereas at low speed high torque at large

Sucking volume is available, for example

Accelerate or slow uphill. Also included is a kinematic reversal.

Moreover, it is possible to use the rotary piston engine with a Provide control pressure connection, which is connected to the control cylinder such that a control pressure from the

Control pressure connection acts on the control piston. Thus, the volume of the rotary engine by changing the

Control pressure can be varied. In 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. In other words, 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. In this case, 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

Connection while driving again be used. Because of the risk of cavitation negative pressures of about 0.2 to 0.4 bar

Not to fall below atmospheric pressure, it is possible by means of the rotary piston engine to adjust the negative pressure by varying the absorption volume of the tooth chambers as described above such that a limit negative pressure is not

is fallen short of. The swallow volume of

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

in particular the danger of cavitation in suction operation

to reduce . A major advantage of the 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

opposite side of the hydraulic motor can be grown by the continuous shaft driven filling pump, for example, as a gerotor pump, which in their

Suction behavior is not subject to any narrow limits. If the displacement of this filling pump per shaft revolution is designed to be greater than the smallest displacement of the hydraulic motor, it is able to supply it with a higher pressure than the

To fill atmospheric pressure. This filling pressure can be far above the atmospheric pressure, so that in a large

Speed range of the vehicle in braking mode strong negative pressure and thus cavitation in the hydraulic motor can be avoided. One then speaks in this case in the

Hydraulic language of a "closed circuit".

Also to reduce the above cavitation, especially in pumping operation and to enable an even higher speed of the rotary piston engine, it is advantageous to design the rotary valve as large as possible. Preferably, the rotary valve with its pressure brakes in relation to the rotary piston on an enlarged rotary valve outer diameter. In particular, 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

economical direct drive, in particular wheel hub drive with recuperative brake energy recovery in a vehicle be used. The variable displacement allows economical operation of the rotary engine, since the displacement of the current power and

Adjust torque demand and thus hydraulic losses can be reduced.

The hydrostatic drive according to the invention will now be described in more detail purely by way of example with reference to concrete exemplary embodiments shown schematically in the drawings.

In detail show:

Fig. La is a schematic representation of the hydraulic

Drive for a motor vehicle in the operation of the hydraulic drive motor from the third high-pressure hydraulic accumulator and by means of the liquid hydraulic pump filled first high-pressure hydraulic feeders;

Fig. Lb is a schematic representation of the hydraulic

Drive for a motor vehicle during operation of the hydraulic drive motor from the first high-pressure hydraulic accumulator and filled by means of the liquid hydraulic pump second

High pressure hydraulic Speieher;

Fig. Lc is a schematic representation of the hydraulic

Drive for a motor vehicle during operation of the hydraulic drive motor from the second

High-pressure hydraulic accumulator and third filled by means of the hydraulic fluid pump

High pressure hydraulic Speieher; 2 shows a schematic representation of a hydraulic drive motor with a filling pump;

Fig. 3 is a schematic plan view of a with the

hydraulic drive equipped vehicle;

Fig. 4 is a longitudinal section through a wheel motor

trained hydrostatic rotary piston engine in the position of the working and adjusting organs at maximum absorption per shaft revolution;

Fig. 5 is a same longitudinal section in the position of

Working and adjusting with minimum

Amount of swallow per shaft revolution;

Fig. 6 shows a cross section along the section line A-A of

Figure 1, in which the work spaces as tooth chambers between the orbiting rotary piston and the

Stator are visible;

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

Figure 4, the non-axially movable, but together with the rotor an orbit movement

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;

10 shows a longitudinal section through the wheel motor in the region of the required pressure lines for acting on the regulating members designed as control pistons which are to be under the control pressure; and

Fig. 11 is a longitudinal section through an alternative

From a guide shape of a wheel motor according to the invention in the position of the working and adjusting members with maximum absorption per shaft revolution, wherein a spread of the amount of absorption in the ratio of about 6: 1 is possible.

In the following, possible embodiments with reference to several figures, which partially show a single embodiment in different views with sporadically different levels of detail, shown, the figures are partly explained together. Reference is made to features and reference signs already mentioned in the preceding figures.

In the figures la to lc is in a schematic

Illustration of a possible embodiment of the hydraulic drive for a motor vehicle abstractly shown. 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.

By means of an internal combustion engine 101, known as gasoline,

Diesel, gas, Stirling or other engine may be formed, is a liquid hydraulic pump 103rd

drivable. 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

The structure for simplifying the illustration with reference to a single hydraulic drive motor 110 will be explained below. 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

Vehicle operation sufficient range is continuously variable. 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. One

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. 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. In a recuperation mode B, 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. In the

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

Check valve 122, the 4/3-way valve 108, so that the working fluid from the low-pressure hydraulic accumulator 111 can be sucked to the high-pressure port 1. Alternatively, it is possible, instead of the recuperation mode B the

Reverse mode R to use for recuperation.

In the pressure connection between the liquid hydraulic pump 103 and the high-pressure hydraulic accumulator 104, 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. In the pressure connection between the high-pressure hydraulic accumulator 104 and the selector slide 108 is a

Current regulator 109 in the form of a flow valve, by means of which of the high-pressure hydraulic accumulator 104 through the

hydraulic drive motor 110 in the low-pressure hydraulic accumulator 111 flowing mass flow of the working fluid to

Regulation of the output of the drive motor 110 and thus the drive wheel 113 is adjustable. The flow control valve 109 can serve as a throttle valve or preferably as a flow control valve

be formed, by means of which the volume flow is constantly adjustable even at a pressure gradient. In order to be able to retrieve the maximum performance of the internal combustion engine under extreme driving conditions, 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.

Furthermore, 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 and is designed such that when

Driving the motor vehicle in forward traction, ie the torque output of the drive motor 109, the forward mode D and the deceleration of the motor vehicle in

Forward-overrun, especially when braking, the

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. Alternatively, a mechanical, hydraulic or pneumatic

Control and / or control device possible. In the

Recuperation mode B acts in forward-coasting operation of the hydraulic drive motor 110 as one of the

at least one drive wheel 113 driven hydraulic pump. 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. The kinetic energy of

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. Thus, a recuperative Bremsenergie- use is possible.

The hydraulic drive motor 110 has a continuous drive shaft 11, as shown schematically in FIG shown. On the continuous drive shaft 11 of the hydraulic drive motor 110, a filling pump 114 is arranged, which in the direction of at least one

hydraulic drive motor 110 pumps and is driven directly by this. The filling pump 114 is located in

Saugstrang between the low-pressure hydraulic accumulator 111 and the high-pressure port 1 of the hydraulic drive motor 110. In the recuperation mode B by means of the filling pump 114 is an excessively strong vacuum on

High pressure port 1 and thus the risk of cavitation in the drive motor 110 avoided. The specific amount of suction of the filling pump 114 per shaft revolution is significantly greater than the smallest possible amount of hydraulic fluid

Drive motor 110 per shaft revolution. Preferably, 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.

In the pressure connection between the several smaller ones

High-pressure hydraulic accumulators 104a, 104b, 104c, the

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. By means of this valve arrangement 115, 116 and 117 is between one of

Number of several smaller high-pressure hydraulic accumulators 104a, 104b, 104c corresponding number of switching states switchable. As in the embodiment, three smaller

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. Thus, 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.

In FIG. 1 a, 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

However, communication between the first smaller high pressure hydraulic accumulator 104a and the working fluid flow regulator 109 and thus the hydraulic drive motor 110 is

blocked. By contrast, 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.

In FIG. 1b, however, 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

illustrated. The connection between the second small high pressure hydraulic accumulator 104b and the

Working fluid flow regulator 109 and thus 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

smaller high pressure hydraulic accumulator 104c of the

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.

In the following, a preferred, as operating on the orbit principle, variable volume hydrostatic high-torque rotary piston engine will be described with reference to Figures 4 to 11. These figures show in part a single embodiment in different views with sporadically different levels of detail, which is why these figures are partially explained together. Reference is made to features and reference signs already mentioned in the preceding figures.

The working pressure of the volume flow

Working fluid in the form of a working fluid is

For example, when driving forward of the wheel motor shown in Figure 4 a high-pressure port 1 of an input and

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. By the

Commutation of the rotary valve 3 flows the working fluid through window 7 in a power section 51st

As shown in FIG. 6, the power unit 51 has a center fixed stator 52 having one

cylindrical rollers 13 formed stator inner teeth with the number of teeth d = 12, designed as a rotary piston rotor 8 with a partially engaging in the stator inner teeth 13 rotor outer teeth 12 with a number of teeth c = ll, a rotor internal teeth 9 with a number of teeth b = 14 and a rotor outer diameter 68. In addition, the power unit 51 comprises the shaft 11 rotatably mounted centrally about the shaft axis 50 with a partially engaging in the rotor internal teeth 9 first shaft outer teeth 10 with a number of teeth a = 12. The rotary piston 8 is thus for performing an orbital movement

arranged eccentrically to the shaft axis 50 and the numbers of teeth a, b, c, d are, as shown above, dimensioned such that with working fluid ver and disposable tooth chambers 6 form radially between the stator inner teeth 13 and the rotor outer teeth 12.

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

Input acting high-pressure port 1 in a first part 6a of the tooth chambers 6 with the working pressure into and out to a low-pressure connection acting as an output from a second part 6b of the tooth chambers 6 out. To the largest possible amount of working fluid m

To be able to guide tooth chambers 6 into and out of the tooth chambers, 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.

Since the rotary valve 3, together with windows 7, ensures that e.g. In FIG. 6, only the first part 6a of the tooth chambers 6 to the left of the axial distance line D-D is acted upon by working pressure, the rotary piston 8 in FIG. 6 rotates in FIG

Clockwise while outputting its torque via the internal rotor teeth 9 of the rotary piston 8 and a first shaft outer teeth 10 of a shaft 10, wherein the lower

lower working fluid in the second part 6b of the tooth chambers 6 is displaced to the right of the axial distance line D-D and discharged via the low-pressure connection 70. The torque can be tapped via a driven out of the housing 19 output-side portion 65 of the shaft 10. The rotary piston 8 and the shaft 10 rotate in the same direction, in the embodiment shown in the

Clockwise, as in Figure 6 by the two arrows

illustrated.

The rotary valve 3 is in a known manner by means of a

Taumelgetriebes 56, as a synchronous drive for the

Rotary valve 3 is used, driven. An eccentric 57 of the

Taumelgetriebes 56 has an eccentric internal teeth 58 with a number of teeth x and an eccentric Außenverzahlung 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 Außenverzahlung 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

cup-shaped connecting part 62 with the centrally rotatable rotary valve 3 1: 1 rotatably connected via a tumbling connection. A rotational movement of the shaft 11 is thus transmitted via the eccentric 57 to the rotary valve 3.

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.

In order to fix the rotary valve 3 axially and axially seal against leakage of the working fluid, a rotationally fixed axial second compensating piston 55, which acts axially on the rotary valve 3, arranged around the wobble gear 56. Of the

fixed internal ring gear 61 is formed within this axial second balance piston 55. This mode of operation of a hydrostatic rotary piston engine is essentially known from WO 2006/010471.

The amount of swallowing and thus the torque that the

Rotary piston 8 and the shaft 11 passes, hang

decisively on the size of the effective width of the tooth engagement in the tooth chambers 6 between a rotor external toothing 12 and the stator internal toothing 13. In the figure 6, the 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

is displaceable. 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. In the following we explain this general arrangement in more detail.

The axially movable stator plate 14, which in Figure 7 in

Cross-section is shown, engages with their Statorplatten- external teeth 64 in the stator inner teeth 13 with the number of teeth d = 12 in a sealing manner in the axial direction fitting that the stator plate 14 relative to the axially fixed stator 52 together with the rotary piston 8 axially displaceable is. With the stator 14 is a - in particular annular - compensating piston 15a axially coupled. The compensating piston 15 a is formed by a piston extension of the stator plate 14. In other words, the balance piston 15a and the

Stator plate 14 integrally formed. This

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

Working pressure in the tooth chambers 6 counteracts and thus a force on the rotary piston 8 in the direction of

Starting position with a reduced volume of

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.

As can be seen from FIG. 4, the width of the

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. In this case, the stator plate 14 with its piston neck, the

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 highest possible controlled working pressure in the tooth chambers 6 is achieved, provided that this is not yet increased over the block size of the control springs 16, which would be quite possible. In FIG. 5, the stator plate 14 m is at its left limit position. The control springs 16 are largely relaxed, corresponding to a low pressure in the tooth chambers 6. The effective width of the tooth engagement between the rotor outer teeth 12 and the stator inner teeth 13 is greatly reduced. This results in the desired strong reduction of the specific amount of swallowing in the ratio of the reduced effective tooth width to the maximum effective tooth width in Figure 4. The transition from the maximum

effective tooth width to the minimum effective tooth width can be done continuously, which is why a stepless

Control of the swallowing amount is possible.

In the drive motor 110, 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.

The closing of the tooth chambers 6 to the left in the

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

is displaceable. This rotor plate 17 thus has the

Orbit movement of the rotor 8 with as little as possible

Join in friction loss. It is not axially movable and is guided radially movable with a minimal axial clearance between the housing parts 19a and 19b. 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

is pierced, so that by the orbit movement of the

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

Output section 65 of the shaft 11 arranged 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. In the housing 19, 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

Moving the volume of the tooth chambers 6 can be actively changed in particular. 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 fact that it is even possible to design a rotary piston machine according to the orbit principle so that a stepless change in the absorption volume is possible, is due to the fact that according to the execution of such

Machine according to WO 2006/010471, 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

is axially movable and there against tilting a good

Leadership relationship gets. The axial force vectors in this rotary piston machine are in no way offset without offset, but run past each other at a considerable distance. While the force of the balance piston 15a on the stator plate 14 and the control piston 22a exactly offset in the center of the machine and compensate in part, the resulting hydraulic axial force of the tooth chambers 6 is significantly offset from the piston forces. This results in a considerable tilting moment on the one hand on the stator plate 14 and the

Rotor plate 17 and thus also on the rotary piston 8. Um

To avoid tilting and resulting large frictional forces, you need for both the rotary piston 8 and in particular for the stator plate 14 with her

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

perform.

Extensive tests have shown that even under high working pressure the required axial displacement in both directions is possible with almost no hysteresis. It is on it To ensure that the corresponding, under load sliding surfaces are processed as smooth as possible.

In FIG. 9 channels 71 are shown, in which the

Working pressure from the high pressure port 1 to the balance piston 15a, the balancing cylinder 15b and a

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.

As shown in FIGS. 4 and 5, in a first embodiment of the invention, there is such a direct or indirect connection between the high-pressure port 1 and the control cylinder 22b that the control piston 22a communicates with the control piston 22a

Working pressure of the working fluid is applied, so that the control piston 22a with increasing working pressure the rotary piston 8 and the stator plate to the right in the in Figure 4

shown position with an increased volume of the tooth chambers 6 suppressed, so that the torque increases with increasing working pressure. The working pressure in the control piston 22a thus counteracts the control springs 16 as well as the

Working pressure in the tooth chambers 6. In another embodiment, shown in Figure 10, 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. In FIG. 10, therefore, a separate pressure line with a control pressure connection 67 is provided for the supply of the regulating piston 22a. Thus, it is possible to reduce the amount of

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.

There are two adjustment conceivable:

When the control pressure port 67 is connected to the high pressure port 1, the engine automatically operates as

Torque converter. As in the application of the wheel motor as a vehicle drive with increasing driving resistance of

Working pressure at high pressure port 1 and at

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

Meshing width and thus in the direction of a larger amount of swallowing. Since the torque generated on the shaft 11 is calculated from the product pressure times the amount of swirl, the torque of the wheel motor increases disproportionately

Pressure increase, so that the traction of the drive

increases dramatically accordingly. This automatism can be achieved by the design of the core lines of the control springs 16

be modified. At a constant flow of the supply pump, which the working fluid to the

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. The

Vehicle speed therefore drops. This results in a similar behavior as in a hydrodynamic

Föttinger converter, as it is used in almost all automatic transmissions. The efficiency of the hydrostatic

However, converter is better than the hydrodynamic converter, since the oil flow losses are smaller. In the current effort to use as much braking energy recuperatively, however, is another kind of

Regulation makes sense. As explained above, such orbit motors can only be used to a limited extent as hydraulic pumps. When braking from high vehicle speed, the hydraulic motor can be operated as a pump only with a small amount of swirl due to the risk of cavitation. Does that fall?

Vehicle speed in the course of the braking process, however, from, the amount of swirl of the variable wheel motor in

increased inversely proportional to speed decrease. In this regulation, care must be taken that the total amount of oil produced by the hydraulic motor has a certain

Limit does not exceed. As already mentioned, the hydraulic motor during braking as a pump promotes the fluid against a largely constant accumulator pressure. The braking torque can thus only with decreasing vehicle speed

Increasing the absorption volume can be increased in compliance with the mentioned limit of the oil flow rate. In this case, a pressure independent and thus

Demand-dependent regulation required. This is done via a control pressure at the control pressure port 67.

For the control of the continuously variable amount of swirl per shaft revolution of the drive motor 110, 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.

A big advantage in the inventive

volume-variable hydrostatic rotary piston engine, as well as in the hydraulic motor according to WO 2006/010471, the continuous wave. 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

Stator plate 14 and rotor plate 17 and thus also a larger adjustment of the adjusting 22 shown. As a result, a spread of the swallow volume with the value 6: 1 is achieved. This corresponds to

Gear automatic transmissions a 6-speed gearbox with the appropriate design effort. Opposite the

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

Embodiment to which reference is hereby made.

Of course, the invention is not limited to the exemplary embodiments shown schematically and illustrated here in the drawings.

Claims

Hydrostatic drive for a motor vehicle with

A low-pressure hydraulic accumulator (111) for receiving a working fluid,

A high-pressure hydraulic accumulator (104) for receiving the working fluid,

• an accumulator (105), in particular a

Gas storage such as a nitrogen bottle, the the

High-pressure hydraulic accumulator (104) is connected downstream for biasing with compensating pressure,

An internal combustion engine (101),

A liquid hydraulic pump (103),

- Which is coupled to the internal combustion engine (101) such that it is driven by the internal combustion engine (101), and

- In a pressure connection between the

Low-pressure hydraulic accumulator (111) and the

High-pressure hydraulic accumulator (104) for filling the high-pressure hydraulic accumulator (104) with the

Working fluid under working pressure against the

Balancing pressure of the pressure accumulator (105) is arranged,

• at least one hydraulic drive motor (110), the

- With at least one drive wheel (113) of

Motor vehicle is coupled directly or indirectly,

a high pressure port (1) connected in a forward mode (D) to the high pressure hydraulic accumulator (104) and a low pressure port (70) operating in the forward mode (D) with the low pressure hydraulic Memory (111) is connected, and - With the working fluid from the high-pressure hydraulic accumulator (104), in particular with a means of a working fluid flow regulator (109) metered amount of working fluid for driving the at least one drive wheel (113)

Emptying the high pressure hydraulic accumulator (104) is operable,

characterized in that

• the high-pressure hydraulic accumulator (104) in several

smaller high-pressure hydraulic accumulator (104a, 104b, 104c) is divided,

- Which are coupled together with the downstream pressure accumulator (105) and

- The so with the at least one hydraulic

Drive motor (110) are interconnected, that the several smaller high-pressure hydraulic accumulator

(104a, 104b, 104c) are substantially independently connectable to the at least one hydraulic drive motor (110) and are alternately emptied,

and

• The at least one hydraulic drive motor (110) has a variable displacement volume in an adjustment range, in particular stepless or graduated

Has shaft rotation, the specific

Swallowing amount of the hydraulic drive motor (110) per shaft revolution in an area sufficient for vehicle operation - in particular stepless or stepped - is variable.

Hydrostatic drive for a motor vehicle after

Claim 1,

characterized in that • the plurality of smaller high-pressure hydraulic accumulators (104a, 104b, 104c) are connected to the liquid hydraulic pump (103) in such a way that the

a plurality of smaller high-pressure hydraulic accumulator (104a, 104b, 104c) substantially independent of each other with the liquid hydraulic pump (103) can be connected and alternately filled.

Hydrostatic drive for a motor vehicle after

Claim 1 or 2,

marked by

a valve arrangement (115, 116, 117), which in particular comprises a plurality of directional control valves,

• in a pressure connection between the several

smaller high-pressure hydraulic accumulators (104a, 104b, 104c), the at least one hydraulic drive motor (110) and the liquid hydraulic pump (103) is arranged and

• By means of which between one of the number of several smaller high-pressure hydraulic accumulators (104a, 104b, 104c) corresponding number of switching states

is switchable,

wherein in each switching state

One of the plurality of smaller high pressure hydraulic accumulators (104a, 104b, 104c) exclusively for filling with the liquid hydraulic pump (103)

is connected and

Another one of the plurality of smaller high-pressure hydraulic accumulators (104a, 104b, 104c) is connected exclusively to the at least one hydraulic drive motor (110) for emptying,

so that the several smaller high pressure hydraulic accumulators (104a, 104b, 104c) are alternately fillable and be emptied.

4. Hydrostatic drive for a motor vehicle according to one of claims 1 to 3,

marked by

an operating mode switching unit (108), in particular a selector slide as a 4/3 way valve, in a

Pressure connection between the at least one

hydraulic drive motor (110) and the high-pressure hydraulic accumulator (104) and between the at least one hydraulic drive motor (110) and the

Low-pressure hydraulic accumulator (111) is arranged and by means of which between

• the forward mode (D) and

A recuperation mode (B), in which the

High-pressure port (1) with the low-pressure hydraulic accumulator (111) and the low-pressure port (70) is connected to the high-pressure hydraulic accumulator (104) is switchable.

5. Hydrostatic drive for a motor vehicle after

Claim 4,

marked by

a control device (121), which is in such a control connection with the mode changeover unit (108) and is designed such that

When driving the motor vehicle in the forward traction mode, the forward mode (D) and

• When decelerating the motor vehicle in the forward thrust operation, in particular when braking, the

Recuperation mode (B) is granted, wherein in the recuperation mode (B) in the forward overrun mode, the at least one hydraulic drive motor (110) as one of the at least one drive wheel (113) directly or indirectly driven hydraulic pump, which pumps the working fluid from the low pressure hydraulic accumulator (111) into the high pressure hydraulic accumulator (104), acts for recuperative braking energy usage.

6. Hydrostatic drive for a motor vehicle according to one of claims 1 to 5,

characterized in that

the at least one hydraulic drive motor (110) exclusively with the working fluid from the

High pressure hydraulic accumulator (104) is deliverable.

7. Hydrostatic drive for a motor vehicle according to one of claims 1 to 6,

characterized in that

the at least one hydraulic drive motor (110) has a continuous drive shaft (11).

8. Hydrostatic drive for a motor vehicle after

Claim 7,

characterized in that

on the continuous drive shaft (11) of the at least one hydraulic drive motor (110), a filling pump (114) which pumps in the direction of the at least one hydraulic drive motor (110) in the suction line between the

Low-pressure hydraulic accumulator (111) and the at least one hydraulic drive motor (110) - in particular to avoid strong negative pressure on

High-pressure connection (1) at least in the recuperation mode (B) - is arranged.

9. Hydrostatic drive for a motor vehicle after

Claim 8,

characterized in that the specific amount of swirl of the filling pump (114) per shaft revolution is substantially greater than the smallest possible amount of swirling of the at least one hydraulic drive motor (110) per shaft revolution.

10. Hydrostatic drive for a motor vehicle after

Claim 8 or 9,

characterized in that

the specific amount of swallowing of the filling pump (114) is approximately equal to half the maximum specific amount of swallowing of the at least one hydraulic drive motor (110) per shaft revolution.

11. Hydrostatic drive for a motor vehicle according to one of claims 1 to 10,

characterized in that

the ratio of the gas pressure of the pressure accumulator (105) between the maximum value and the lowest value has the value 1.1 to 1.25.

12. Hydrostatic drive for a motor vehicle according to one of claims 1 to 11,

characterized in that

for controlling the - in particular stepless or graduated - variable displacement per

Shaft rotation of at least one hydraulic

Drive motor (110) is provided an electronically actuated magnetic two-edge control valve. 13. Hydrostatic drive for a motor vehicle according to one of claims 1 to 12,

characterized in that

in a pressure connection between the liquid hydraulic pump (103) and the high-pressure hydraulic accumulator (104) a power shifter (118) for switching the liquid hydraulic pump (103) is switched from a non-pressurized circulation in a high working pressure power mode.

14. Hydrostatic drive for a motor vehicle according to one of claims 1 to 13,

characterized in that

the working fluid flow regulator (109) for the

In extreme cases, a sufficiently large flow rate

or in parallel has a bypass valve.

Hydrostatic drive for a motor vehicle according to one of claims 1 to 14,

characterized in that

the internal combustion engine (101) is designed as a WANKEL rotary engine.

16. Hydrostatic drive for a motor vehicle according to one of claims 1 to 15,

characterized in that

in the at least one hydraulic drive motor (110) is a swallow volume - in particular continuously - changing organ, in particular a balance piston (15a) with a control spring (16), arranged such that at fluctuating working pressure to the drive wheel (113) output torque automatically is kept substantially constant. 17. Hydrostatic drive for a motor vehicle according to one of claims 1 to 16,

characterized in that

in a pressure connection between the high pressure Hydraulic accumulator (104) and the hydraulic

Drive motor (110)

• a working fluid flow regulator (109), the

is designed in particular as a flow control valve, or

• a particular electric motor driven

metering

is arranged to control the drive.

18. Hydrostatic drive for a motor vehicle according to one of claims 1 to 17,

characterized in that

the at least one hydraulic drive motor (110)

• Is designed as a hydrostatic high-torque rotary piston engine according to the orbit principle, and

• Has a continuously variable displacement volume per shaft revolution in an adjustment range, the specific displacement of the hydraulic fluid

Drive motor (110) per shaft revolution in a sufficient for vehicle operation range is continuously variable, that the torque of the at least one hydraulic drive motor (110) be drop in the working pressure in the high-pressure hydraulic accumulator (104) is substantially constant durable.

19. Hydrostatic drive for a motor vehicle after

Claim 18,

characterized in that

the hydrostatic rotary piston engine after the

orbit principle

A power unit (51) which acts as an output and is arranged in a housing (19) - A centric fixed stator (52) with a stator inner toothing (13) with the

Number of teeth d,

- A designed as a rotary piston (8) rotor with a partially in the stator inner teeth (13) engaging the rotor outer teeth (12) having a number of teeth c, a rotor internal toothing (9) with a number of teeth b and a rotor outer diameter (68 ), and

- A centrally about a shaft axis (50) rotatable

mounted shaft (11) with a partially in the

Rotor internal teeth (9) engaging first

Shaft external toothing (10) having a number of teeth a includes, wherein the rotary piston (8) for performing an orbital movement is so eccentric to the shaft axis (50) and the numbers of teeth a, b, c, d are dimensioned such that ver with working fluid and disposable tooth chambers (6) form radially between the stator inner toothing (13) and the outer rotor toothing (12),

Having an inlet and outlet part (53) for the supply and disposal of the power part (51) with the working fluid, with the high-pressure port (1) as input for the working fluid under working pressure,

Having a disc-shaped rotary valve (3) which is about the shaft axis (50) centrally rotatable and to the shaft (11) and the stator (52) running centrally, for commuting such as the supply and disposal of the tooth chambers (6) with the Working fluid that the working fluid into a first part (6a) of the dental chambers

(6) with the working pressure into and out of a second part (6b) of the tooth chambers (6) out to generate the drive or driven output, wherein the

Rotary valve has a rotary valve outer diameter (69), • a wobble gear (56) as a synchronous drive for the rotary valve (3), with

- An eccentric (57) having an eccentric internal teeth (58) with a number of teeth x and an eccentric Außenverzahlung (59) having a number of teeth y, wherein the eccentric internal teeth (58) with a second shaft outer toothing (60) of the shaft (11) with a number of teeth w and the eccentric outer payment (59) with one fixed

Internal toothed ring (61) meshes with a number of teeth z, wherein the number of teeth w, x, y, z are dimensioned such that the eccentric (57) eccentric to

Shaft axis (50) between the shaft (11) and the inner ring gear (61) upon rotation of the shaft (11) performs an orbital movement, and

- A pot-shaped connecting part (62) which rotatably with the eccentrically rotatable eccentric (57)

coupled and with the centrically rotatable rotary valve (3) 1: 1 rotatably on a tumbling

Connection is connected such that a

Rotary movement of the shaft (11) via the eccentric (57) on the rotary valve (3) is transferable,

and

Two radial bearings (63) arranged on both sides of the power section (51) and arranged on the shaft (11) for radially supporting the shaft (11),

in which

• the rotary piston (8) with its internal rotor toothing

(9) on the first shaft outer toothing (10) relative to the shaft (11) and with its rotor outer toothing

(12) is axially displaceable on the stator inner toothing (13) relative to the stator (52) parallel to the shaft axis (50), • Adjustment (22) with the rotary piston (8) are engaged, by means of which the rotary piston (8) is axially displaceable and

• the tooth chambers (6) in the axial direction between

- An axially fixed and radially with the

Rotary piston (8) coupled, the orbital movement of the rotary piston (8) exporting rotor plate (17) and

- An axially movable, axially with the rotary piston (8) coupled and radially from the rotary piston (8) decoupled stator plate (14)

are arranged and axially sealed by the rotor plate (17) and the stator plate (14), wherein the volume of the tooth chambers (6) by the axial

Moving the rotary piston (8) and the stator plate (14) by means of the adjusting members (22) continuously

is changeable.