Hydrostatic transmission and methods for braking with it

Abstract

Hydrostatic transmission for a drive system, comprising a primary unit (1) that can be coupled to an internal combustion engine of the drive system via a rotor (4) and a secondary unit (2) that can be coupled to an output of the drive system via a rotor (12), wherein the two units (1, 2) are connected to each other via two working lines (10) of a closed circuit, and wherein the primary unit (1) is an axial piston machine with an adjustable swivel angle (angle_pump) that can be controlled by an electrical control unit (8) when the hydrostatic transmission is braked, characterized in that during braking in engine operation, forces act on the primary unit (1) that depend on a pressure difference (Δp) of the two working lines (10) and optionally on the rotational speed (n_pump_act) of the rotor (4) and lead to an increase in the swivel angle (angle_pump) of the primary unit, wherein the hydrostatic transmission has a speed governor (38),the output variable of which is a correction control pressure (p_ctrl_speed) acting in the direction of a reduction of the swivel angle (angle_pump), which can be added by the control unit (8) to a pilot control pressure of a pilot control, wherein the pilot control pressure is a sum of a basic control pressure (p_ctrl_basic) and a proportional control pressure (p_ctrl_power), wherein during braking the proportional control pressure (p_ctrl_power) is proportional to a delivery volume flow rate (Q_mot) of the secondary unit (2).

Description

[0001] The invention relates to a hydrostatic transmission with which hydrostatic braking is possible, according to the preamble of claim 1, and a method for braking with such a hydrostatic transmission.

[0002] Hydrostatic transmissions for mobile machinery are known from the prior art, in which a hydrostatic pump (primary unit) and a hydrostatic motor (secondary unit) are fluidically connected to each other via a closed circuit. An internal combustion engine, e.g., a diesel engine of the mobile machinery, is coupled to the primary unit, and an output, e.g., an axle or a wheel of the mobile machinery, is coupled to the secondary unit in a rotationally fixed manner. Thus, the mobile machinery has a drive system that incorporates a hydrostatic transmission.

[0003] Such a hydrostatic transmission, which can also be used for braking, is disclosed in EP 1 960 699 B1. In this system, the power flows in the opposite direction to traction operation, from the output shaft via the secondary unit (acting as a pump) and the primary unit (acting as a motor) to the internal combustion engine, which is then driven in a passive towing mode. The high-pressure working line of the closed circuit is protected by a pressure relief valve, which also allows a portion of the braking power to be dissipated during braking.

[0004] The problem with such braking is that the internal combustion engine must be driven at a maximum speed to avoid being damaged. Therefore, the primary unit, designed as an adjustable axial piston machine and functioning as the engine, is set to a pivot angle during braking that generates a torque the internal combustion engine can still handle without over-revving. This is done taking into account the pressure in the working line, determined by the setting of the pressure relief valve.

[0005] A disadvantage of prior art hydrostatic transmissions is that the combustion engine is not brought up to its maximum speed during braking. Therefore, a significant portion of the braking power must be dissipated via the pressure relief valve of the affected high-pressure line, which is detrimental.

[0006] Documents DE 10 2014 211 393 A1 and DE 10 2014 211 394 A1 also disclose a hydrostatic transmission that can also be used for braking, with the aim of protecting the internal combustion engine from overspeeding. In these documents as well, a first portion of the braking power is transferred to the internal combustion engine via the primary unit, while a second portion is converted into heat via the pressure relief valve of the relevant high-pressure line. Initially, a slightly higher engine speed than the permissible limit is accepted. Within the framework of the pre-control of the swivel angle, a generally higher swivel angle than ideal is disclosed. Thus, the maximum possible braking power is not supported by the internal combustion engine.

[0007] Document DE 10 2014 224 337 A1 discloses a method for controlling a hydrostatic drive.

[0008] Furthermore, the respective primary units of the aforementioned state of the art are mechanically controlled in terms of their swivel angle. For this, they require mechanical feedback of their current swivel angle.

[0009] In contrast, the invention is based on the objective of creating a hydrostatic transmission and a method in which the equipment-related effort of the primary unit is reduced.

[0010] These problems are solved with respect to the hydrostatic transmission by the combination of features of claim 1 and with respect to the method by the combination of features of claim 10.

[0011] The claimed hydrostatic transmission is intended for a drive system comprising an internal combustion engine, e.g., a diesel engine, and an output, e.g., one or more wheels or one or more axles. The hydrostatic transmission has, on the one hand, a rotor—in particular, a drive shaft—of a primary unit that can be coupled to the internal combustion engine of the drive system and operates as a pump during traction, and on the other hand, a rotor—in particular, a drive shaft—of a secondary unit, or several such hydraulically parallel secondary units that can be coupled to the output of the drive system and operate as motor(s) during traction. Both units are fluidically connected to each other via two working lines of a closed circuit. The primary unit is an axial piston machine with an adjustable swivel angle.The hydrostatic transmission further comprises an electrical control unit via which braking can be controlled or regulated, whereby a braking torque of the secondary unit, acting as a pump, is supported on the rotor of the primary unit, acting as a motor. According to the invention, forces act on the primary unit, whose stroke volume is adjustable, which depend primarily on a pressure difference between the two working lines, but also on the rotational speed of the rotor. When these forces increase, with unchanged control of the primary unit, they lead to a reduction in the swivel angle during pump operation and an increase in the swivel angle of the primary unit during motor operation.

[0012] This means the primary unit exhibits load-sensitive behavior. Specifically, under certain pump control conditions, its oscillation angle depends on the differential pressure between the two working lines of the closed hydraulic circuit and the pump speed. During pump operation, the drive forces, together with a spring arrangement on the actuating piston or swashplate, counteract any adjustment towards larger stroke volumes. Examples of such adjustments include EV adjustment and ET adjustment.

[0013] In the EV adjustment, a control pressure is regulated by means of a pressure control valve adjustable by an electroproportional solenoid and supplied via a directional control valve to one or the other control chamber of the actuator cylinder of the primary unit. The stroke volume of the primary unit at a specific control pressure is influenced by both the rotational speed of the primary unit and the pressure differential between the two working lines.

[0014] The ET control system has two pressure regulating valves. One pressure regulating valve regulates the actuating pressure in one actuating chamber, and the other pressure regulating valve regulates the actuating pressure in the other actuating chamber of the actuator cylinder.

[0015] Furthermore, according to the invention, the hydrostatic transmission has a speed controller whose output is a correction control pressure, which can be added by the control unit to a pilot-operated control pressure of a brake-performance-dependent pilot control. The sum of the control pressures thus formed acts during braking in the direction of reducing the swivel angle of the primary unit. The pilot control largely achieves the target swivel angle of the primary unit. The speed controller takes into account and compensates for manufacturing inaccuracies and wear of the primary unit, as well as varying temperatures of the hydraulic fluid. Since no feedback of the current swivel angle and no position control of the primary unit are required, the hydrostatic and brakeable transmission is simplified in terms of its design.

[0016] With regard to the entire drive system, the braking torque of the output is thus supported via the two units on the combustion engine.

[0017] With regard to the hydrostatic transmission according to the invention, which does not include the internal combustion engine, the actual speed of the internal combustion engine to be limited can be detected by a speed sensor arranged on its crankshaft and thus also not belonging to the hydrostatic transmission according to the invention, and transmitted to the control unit via a signal input. Alternatively, the actual speed to be limited can be detected directly at the rotor of the primary unit by a speed sensor belonging to the transmission according to the invention if the crankshaft and the rotor are integral. With this arrangement of the speed sensor, the actual speed to be limited can also be calculated from the rotational speed of the rotor of the primary unit determined by the speed sensor, if a mechanical gear stage is provided between the crankshaft and the rotor. Preferably, the at least one secondary unit also has a speed sensor.

[0018] Further advantageous embodiments of the invention are described in the dependent patent claims.

[0019] Preferably, the speed controller has as its input a speed difference between, on the one hand, the actual speed of the internal combustion engine or a value derived therefrom, in particular the actual speed of the primary unit, and, on the other hand, a target speed of the internal combustion engine or a correspondingly derived value, in particular a target speed of the primary unit. By using the derived values, the hydrostatic transmission according to the invention can be manufactured, tested, and operated without an electrical or mechanical connection to the internal combustion engine.

[0020] Preferably, the speed controller is a P, PI, or PID controller. This allows for a fast and accurate approximation of the actual speed to the target speed of the internal combustion engine. Preferably, the control unit includes the speed controller.

[0021] In the case of particularly effective braking of the hydrostatic transmission according to the invention, the target speed is a maximum speed that can be supported by the internal combustion engine without it being over-revved.

[0022] In a particularly preferred further development, the pre-controlled actuating pressure is the sum of a constant basic actuating pressure and a proportional actuating pressure.

[0023] The proportional actuating pressure is preferably the product of a power factor and the delivery volume flow rate of the secondary unit. The power factor takes into account that the pivoting forces on the pivoting cradle increase with increasing braking power. The dependence on the differential pressure of the two working lines and larger pressure pulsations can also be considered here. The delivery volume flow rate of the secondary unit is almost linearly proportional to the braking power and is therefore suitable for taking this into account. Consequently, a pressure sensor can be omitted, which increases the robustness of the hydrostatic transmission according to the invention.

[0024] The swivel angle of the primary unit is brought to the optimal value particularly quickly and safely, without over-revving the combustion engine, if, during braking, a final actuating pressure acting in the direction of reducing the swivel angle is a sum of the pre-controlled actuating pressure and the correction actuating pressure.

[0025] All the aforementioned actuating pressures, when braking, result in a reduction of the swivel angle of the primary unit operated as a motor.

[0026] The final actuating pressure is preferably applied via an electric actuating pressure valve and an actuating cylinder of an adjusting device.

[0027] The forces, which depend on a pressure difference between the two working lines and the rotational speed of the rotor and the swivel angle, can act on a swashplate or on an actuating piston of the actuating cylinder.

[0028] In a preferred embodiment of the hydrostatic transmission according to the invention, the swivel angle and thus the stroke volume of the primary unit are adjustable on both sides of a zero position. This allows the affected drive system to be used in both directions of travel of the vehicle during towing operation, with the combustion engine maintaining the same direction of rotation, and to be braked accordingly in both directions according to the invention.

[0029] To enable the actuating pressure to reduce the swivel angle during braking or overrun, and to increase the swivel angle during traction, two opposing actuating pressure chambers are preferably located on the adjustment device of the primary unit. These chambers can be pressurized with their respective actuating pressures via a common or individually electrically adjustable actuating pressure valve. The two actuating pressure chambers can be provided in a double-acting actuating cylinder or in two separate actuating cylinders acting against each other on the swashplate. This allows the affected drive system to be used in both directions of travel during traction, with the combustion engine maintaining the same direction of rotation, and can be braked accordingly in both directions according to the invention.

[0030] Preferably, a spring arrangement is provided which acts in the direction of a central position of the swashplate, in which the swivel angle and the stroke volume of the primary unit are approximately zero.

[0031] To enable high-performance braking, it is preferred to have a pressure relief valve on each of the two working lines. A first portion of the braking force can be dissipated via the respective pressure relief valve, while a second portion can be dissipated via the primary unit and subsequently via the internal combustion engine. The achievable braking force is particularly high when the first portion is greater than the second.

[0032] During high-performance braking, if the flow rate through the primary unit increases, the flow rate through the affected pressure relief valve decreases. This can cause the pressure in the high-pressure working line to drop. To minimize this pressure reduction or to keep the pressure nearly constant, pressure relief valves with a flat characteristic curve regarding their pressure differential as a function of the flow rate are preferred.

[0033] If the hydrostatic transmission according to the invention has pressure relief valves whose characteristic curves are increasing, the proportional control pressure already mentioned is particularly preferred, which is the product of the power factor and the delivery volume flow rate of the secondary unit and forms part of the pilot-controlled pressure. The power factor then takes into account that the forces increase in the direction of increasing swivel angle with increasing braking power.

[0034] To determine the braking power, the delivery volume flow rate of the secondary unit can be used, as this is essentially proportional to the braking power. Instead of the delivery volume flow rate of the secondary unit, the braking torque or the volume flow fraction discharged via the pressure relief valve can also be used, since all these values ​​can be converted into one another using the target swivel angles of the units and their rotational speeds.

[0035] In order to be able to use the hydrostatic transmission according to the invention in various cases in which the internal combustion engine could over-rev according to the prior art, the control unit can be further developed in such a way that braking is initiated either via a control element, e.g. brake pedal, in particular via an input for a signal line of the control element, or on the basis of automatic monitoring of the vehicle's driving speed, in particular via an input for a signal from a speedometer, or on the basis of automatic monitoring of the actual speed of the internal combustion engine or the quantity derived therefrom.

[0036] Preferably, at least one secondary unit is also designed with an adjustable swivel angle and thus stroke volume. This can be adjusted, for example, depending on the control element. This allows the braking torque to be controlled during braking. The control can be electro-proportional.

[0037] The adjustment device can be an electrically controlled adjustment device (EV or ET) in which a specific actuating pressure is regulated according to the force of a proportional magnet.

[0038] The inventive method for braking a hydrostatic transmission according to one of the preceding claims comprises the following steps: - Pre-control of the pre-controlled actuating pressure of the primary unit, - Calculating the correction control pressure as a function of a speed difference between the actual speed of the internal combustion engine or the quantity derived therefrom and the target speed of the internal combustion engine or the quantity derived therefrom, and - Adding the pre-controlled actuating pressure with the correction actuating pressure to obtain the final actuating pressure.

[0039] Therefore, feedback on the current swivel angle is not necessary.

[0040] In a preferred embodiment of the inventive method, braking is initiated in advance by the driver's request or due to an exceedance of the internal combustion engine's speed (actual speed exceeds a desired or a maximum permissible speed) or due to an exceedance of a driving speed (driving speed exceeds a desired or a maximum permissible driving speed).

[0041] Preferably, the braking torque during braking is controlled by pivoting the secondary unit. This can be done depending on the position of the control element – ​​in particular the brake pedal.

[0042] The following advantages of the high-performance braking system according to the invention can also be utilized when a less expensive primary unit without position control is used: - maximum utilization of the braking capacity of the closed circuit, - Protection of the diesel engine from overspeed, - adjustable braking effect, - virtually wear-free (purely hydrostatic) braking, and - possibly reducing the size of the mechanical braking system by using a combined braking system.

[0043] While high-performance braking has so far only been used for hydrostatic drives with position-controlled pumps, which are mainly found in agricultural machinery, the use of the load-sensing primary unit according to the invention allows the function to also be used in the construction machinery sector.

[0044] An embodiment of the hydrostatic transmission according to the invention is shown in the drawings. The invention will now be explained in more detail with reference to the figures in these drawings.

[0045] They show Fig. 1 a circuit diagram of the hydrostatic transmission according to the invention in the exemplary embodiment, Fig. 2 a diagram of the stroke volumes of the primary unit and the secondary unit of the hydrostatic transmission Fig. 1 when braking, Fig. 3 a schematic overview of the feedforward control with additional control of the actuating pressure of the primary unit during braking, Fig. 4 Two diagrams of the feedforward control without speed control according to the invention during braking and Fig. 5 Two diagrams of the feedforward control with additional speed control according to the invention during braking.

[0046] Fig. Figure 1 shows a circuit diagram of the hydrostatic transmission according to the invention. It comprises a primary unit 1 and a secondary unit 2, both of which are designed with adjustable displacement Vg_pump and Vg_mot. The primary unit 1 is an axial piston machine to whose drive shaft 4 a crankshaft of a diesel engine (not shown) is rotationally fixed. The rotational speed of the drive shaft 4 is monitored via a speed sensor 6 and an electrical control unit 8. This indirectly monitors the rotational speed of the crankshaft of the diesel engine as well.

[0047] The primary unit 1 is fluidically connected to the secondary unit 2 via a closed hydraulic circuit comprising two working lines 10. A (not shown) output shaft is rotationally fixed to a drive shaft 12 of the secondary unit 2. The output shaft is, for example, a differential gear of a driven axle of a mobile construction machine operating in loading mode.

[0048] The hydrostatic transmission according to the invention Fig. Thus, unit 1, together with the diesel engine and the output shaft, forms a drive system for a mobile construction machine. In towing operation, the drive shaft 4 of primary unit 1 serves as the drive shaft, and primary unit 1 operates as a pump, while secondary unit 2 operates as a motor, and its drive shaft 12 is an output shaft.

[0049] In braking mode of the hydrostatic transmission according to the invention, the output is supported via the drive shaft 12 and via the secondary unit 2, which acts as a pump, and via one of the two working lines 10, and via the primary unit 1, which acts as a motor, and via the drive shaft 4 of the primary unit 1 on the diesel engine, which is then dragged along and reduces at least part of the braking energy of the mobile construction machine via its friction and acceleration forces, e.g., of its pistons.

[0050] Each working line 10 is equipped with a pressure relief valve 14, through which the respective working line 10 can be depressurized towards a supply line 16. This enables high-performance braking, in which a significant portion of the braking force is dissipated via the pressure relief valve 14 from the high-pressure working line 10.

[0051] The feed line 16 is filled from a tank T with feed pressure medium by a feed pump 18, which is rotationally fixed to the drive shaft 4 of the primary unit 1.

[0052] Furthermore, the feed line 16 can be relieved via a pressure relief valve 20 to a tank T. The feed line 16 is connected to the two working lines 10 via a spring-loaded check valve 22, so that the respective low-pressure working line 10 can be supplied with supplementary pressure medium from the feed line 10 if necessary. For this purpose, the opening directions of the two check valves 22 are directed from the feed line 16 to the respective working line 10.

[0053] The displacement volume Vg_pump of the primary unit 1 is adjusted via an adjusting device 24. This device has a double-acting actuating cylinder 26, whose two actuating pressure chambers act against each other on an actuating piston. The actuating piston is coupled to a swashplate of the primary unit 1, which is designed as an axial piston machine. In this case, the actuating cylinder is a differential cylinder, but it could also be a synchronous cylinder. Each of the two actuating pressure chambers can be filled with hydraulic fluid from the feed line 16 via a separate actuating pressure valve 28. Both actuating pressure valves 28 are electrically actuated by the control unit 8. Furthermore, the adjusting device 24 has a spring assembly (not shown) that biases the piston of the actuating cylinder 26 and the swashplate of the primary unit 1 to a central position. From this position, the displacement volume of the primary unit 1 can be adjusted in both directions.Since the primary unit 1 enables four-quadrant operation, the hydrostatic transmission according to the invention can be used for forward driving and forward braking, as well as for reverse driving and reverse braking.

[0054] In the illustrated embodiment, as already mentioned, the secondary unit 2 is also adjustable. This is achieved by an adjusting device 30, which includes a valve with an electric actuator that is also controlled by the control unit 8. The adjusting device 30 has a feedback spring 32, enabling control of the swivel angle of the secondary unit 2.

[0055] High-performance braking via the hydrostatic transmission according to the invention can be initiated by: a) a signal transmission from a control element operated by a driver, for example a brake pedal, to the control unit 8, b) a borderline high speed of the diesel engine, which was determined by the control unit 8 via the speed sensor 6, c) a borderline high driving speed v_veh of the mobile construction machine, which was indirectly detected by a speed sensor 34 of the drive shaft 12 of the secondary unit 2 and transmitted to the control unit 8.

[0056] Fig. Figure 2 shows in a diagram the time course of the swivel angle angle_pump of the primary unit 1 and the course of the swivel angle angle_mot of the secondary unit 2, first during a transition from train operation to high-performance braking operation and then during a transition from high-performance braking operation back to train operation.

[0057] The high-performance braking system is controlled by the control unit 8. Initially, the swivel angle (angle_pump) and thus the displacement volume (Vg_pump) of the primary unit 1 are significantly reduced. Simultaneously, the adjusting device 30 of the secondary unit 2 is actuated such that its swivel angle (angle_mot) and thus its displacement volume (Vg_mot) are also reduced. This increases the pressure in the now high-pressure working line 10, and the corresponding pressure relief valve 14 opens a connection to the feed line 16. The swivel angle (angle_mot) of the secondary unit 2 is then significantly increased.

[0058] To maximize the actual speed of the diesel engine while not exceeding a permissible maximum speed, the following is done in relation to the Fig. 3 to 5 explained control and regulation used.

[0059] According to Fig. 3. To generate a final control pressure p_ctrl, a feedforward control system is superimposed. The feedforward control system consists of two components of the control pressure: a constant basic control pressure p_ctrl_basic and a proportional control pressure p_ctrl_power. The latter component is the product of the flow rate Q_mot of the secondary unit 2 and a power factor K. This factor accounts for the fact that the forces acting on the pivoting cradle increase with increasing braking power. It also allows for consideration of the dependence on the differential pressure Δp of the two working lines 10 and larger pressure pulsations.

[0060] A speed controller 38 compares the actual speed n_eng_act of the diesel engine with its target speed n_eng_des and generates a correction control pressure p_ctrl_speed for the final control pressure p_ctrl. Alternatively, the speed n_pump_act of the primary unit 1 can be compared with a converted target speed n_pump_des of the primary unit 1, and the correction control pressure p_ctrl_speed can be generated.

[0061] During high-performance braking, "negative" control pressures occur compared to train operation in the same direction of travel. This means that in the case of high-performance braking while traveling forward, one of the actuating pressure chambers must be pressurized with actuating pressure, which is actually required for reverse travel, in order to hold the pivoting cradle at the corresponding pivot angle and to transition from train operation or deceleration operation without high-performance braking to high-performance braking operation within the required time. This is explained in the following Fig. 4 and Fig. 5 (each above) can be recognized by the removal of the current I_pump_f of the actuating pressure valve 28 during forward travel and the activation of the current I_pump_r of the actuating pressure valve 28 during reverse travel.

[0062] Fig. Figure 4 shows the pure feedforward control. The primary unit 1 is already pivoted close to the ideal swivel angle angle_pump. In the example shown, however, the primary unit 1 is still set to an overall excessive displacement volume Vg_pump, which leads to an unacceptably high rotational speed n_eng of the diesel engine. In the case of an excessively small swivel angle angle_pump, the braking capacity of the diesel engine would only be partially utilized. Thus, high-performance braking with the load-sensing primary unit 1 is fundamentally possible, but only with limited performance.

[0063] The solution according to the invention represents an improvement in the high-performance braking of the load-sensitive primary unit 1, such that comparable behavior to that achieved with a position-controlled primary unit 1 is obtained: The superimposed speed controller 38 finally compensates for the last deviations from the ideal value, as shown in Fig. 5 can be seen. This results in a consistent load on the diesel engine during high-performance braking, without it reaching the impermissible range of its speed n_eng.

[0064] In Fig. Figure 4 illustrates the behavior of the hydrostatic transmission according to the invention under pure feedforward control. The curve of the control current I_pump_r is shown, increasing from the start of high-performance braking at t=17.5s until t=19.5s, as required by the feedforward control. However, the value determined for the feedforward control leads to an excessively high actual speed of the diesel engine n_eng_act of 3500 rpm at this operating point.

[0065] In Fig. Figure 5 above shows the course of the control current I_pmp_r when using the superimposed speed controller 38, indicated by a solid line. It can be seen that at the start of high-performance braking at t = 17.5 s, the control current I_pmp_r is only slightly higher than in Fig. 4, however, with the effect that the actual speed n_eng of the diesel engine is only increased to 2500 rpm and then drops again in a controlled manner, although the control current I_pmp_r is higher compared to Fig. 4 is then significantly reduced.

[0066] A hydrostatic transmission has been revealed that enables high-performance braking. In this system, at least one variable-speed drive motor, acting as a pump, is supported via a closed circuit by a variable-speed axial piston machine, which in turn can be supported by an internal combustion engine. To prevent the engine from over-revving, a control unit can adjust the axial piston machine's oscillation angle during braking. The axial piston machine is load-sensitive because there is a relationship between its actuated pressure, its rotational speed, and its oscillation angle, as well as a pressure differential between the two working lines of the closed circuit. A feedforward control system sets the axial piston machine's oscillation angle during braking, and a speed controller corrects this angle to achieve a target speed at the internal combustion engine. This target speed is, in particular, the maximum speed at which the internal combustion engine will not over-rev. Reference symbol list 1 primary unit 2 Secondary unit 4 Rotor / Drive shaft 6 Speed ​​sensor 8 Control unit 10 Work Management 12 Rotor / Drive shaft 14 Pressure relief valve 16 Feed line 18 Feed pump 20 Pressure relief valve 22 Check valve 24 Adjustment device 26 actuator cylinders 28 Control pressure valve 30 Adjustment device 32 Feedback spring 34 Speed ​​sensor 38 speed controllers angle_mot Swivel angle of the secondary unit angle_pump Primary unit swivel angle I_pump_f Control current of the actuating pressure valve for forward travel I_pump_r Control current of the actuating pressure valve for reverse driving or for braking n_eng_act Actual speed of the internal combustion engine n_eng_des target speed of the combustion engine n_eng_max maximum permissible speed of the internal combustion engine n_mot_act Actual rotational speed of the secondary unit n_pump_des Target speed of the primary unit derived from the target speed of the internal combustion engine n_pump_max maximum speed of the primary unit derived from the maximum speed of the internal combustion engine K power factor p_ctrl final control pressure of the primary unit p_ctrl_basic Basic actuation pressure of the primary unit p_ctrl_power proportional control pressure of the primary unit p_ctrl_speed Correction pressure of the primary unit Vg_mot Displacement Secondary Unit Vg_pump displacement volume primary unit v_veh Driving speed Q_mot - Flow rate of the secondary unit Q_pump - Suction volume flow rate of the primary unit T Tank Δn_eng Speed ​​difference between the target speed and the actual speed of the internal combustion engine Δn_pump Speed ​​difference between the target speed and the actual speed of the primary unit, derived from the speed difference between the target speed and the actual speed of the internal combustion engine Δp pressure difference between the two working lines

Claims

[1] Hydrostatisches Getriebe für einen Fahrantrieb, das eine mit einem Verbrennungsmotor des Fahrantriebs über einen Rotor (4) koppelbare Primäreinheit (1) und eine mit einem Abtrieb des Fahrantriebs über einen Rotor (12) koppelbare Sekundäreinheit (2) aufweist, wobei die beiden Einheiten (1, 2) über zwei Arbeitsleitungen (10) eines geschlossen Kreises miteinander verbunden sind, und wobei die Primäreinheit (1) eine Axialkolbenmaschine mit einem verstellbarem Schwenkwinkel (angle_pump) ist, der von einer elektrischen Steuereinheit (8) bei einer Bremsung des hydrostatischen Getriebes steuerbar ist, characterized by, dass bei der Bremsung im Motorbetrieb der Primäreinheit (1) Kräfte wirken, die von einer Druckdifferenz (Δp) der beiden Arbeitsleitungen (10) und gegebenenfalls von der Drehzahl (n_pump_act) des Rotors (4) abhängen und zu einer Vergrößerung des Schwenkwinkels (angle_pump) der Primäreinheit führen, wobei das hydrostatische Getriebe einen Drehzahlregler (38) aufweist, dessen Ausgangsgröße ein in Richtung einer Verkleinerung des Schwenkwinkels (angle_pump) wirkender Korrektur-Stelldruck (p_ctrl_speed) ist, der von der Steuereinheit (8) zu einem vorgesteuerten Stelldruck einer Vorsteuerung addierbar ist, wobei der vorgesteuerte Stelldruck eine Summe eines Basis-Stelldrucks (p_ctrl_basic) und eines proportionalen Stelldrucks (p_ctrl_power) ist, wobei beim Bremsen der proportionale Stelldruck (p_ctrl_power) proportional zu einem Förder-Volumenstrom (Q_mot) der Sekundäreinheit (2) ist. [2] Hydrostatic transmission according to claim 1, wherein an input variable of the speed controller (38) is a speed difference (Δn_eng, Δn_pump) between an actual speed (n_eng_act) of the internal combustion engine or a quantity derived therefrom (n_pump_act) and a target speed (n_eng_des) of the internal combustion engine or a quantity derived therefrom (n_pump_des). [3] Hydrostatic transmission according to one of the preceding claims, wherein the sum of the pre-controlled actuating pressure and the correction actuating pressure (p_ctrl_speed) results in a final actuating pressure (p_ctrl). [4] Hydrostatic transmission according to claim 3, wherein the final actuating pressure (p_ctrl) acts via an electric actuating pressure valve (28) and an actuating cylinder (26) of an adjusting device (24) in the direction of reducing the swivel angle (angle_pump) of the primary unit (1). [5] Hydrostatic transmission according to one of the preceding claims, wherein the swivel angle (angle_pump) of the primary unit (1) is adjustable on both sides of a zero position. [6] Hydrostatic transmission according to claims 4 and 5 with two opposing actuating pressure chambers which, during braking, can be actuated with the final actuating pressure (p_ctrl) by the control unit (8) via a common or a respective electrically adjustable actuating pressure valve (28). [7] Hydrostatic transmission according to one of the preceding claims, wherein a pressure relief valve (14) is arranged on each of the two working lines (10), via which a first part of the braking power can be reduced during braking, while a second part of the braking power can be reduced via the primary unit (1), and wherein the pressure relief valves (14) each have a flat characteristic curve with respect to their pressure difference as a function of their volume flow rate. [8] Hydrostatic transmission according to one of the preceding claims, wherein the control unit (8) is designed such that braking can be initiated via a control element of a vehicle or on the basis of automatic monitoring of a driving speed (v_veh) of the vehicle or a quantity derived therefrom (n_mot_act) or on the basis of automatic monitoring of the actual speed (n_eng_act) of the internal combustion engine or a quantity derived therefrom (n_pump_act). [9] Hydrostatic transmission according to one of the preceding claims, wherein the adjusting device (24) is an electrically controlled adjusting device in which a specific actuating pressure is regulated according to the force of a proportional magnet. [10] Method for braking a hydrostatic transmission according to one of the preceding claims comprising the steps: - Pre-control of a control pressure of the primary unit (1), - Calculating a correction control pressure (p_ctrl_speed) as a function of a speed difference (Δn_eng) between an actual speed (n_eng_act) of the internal combustion engine or a derived quantity (n_pump_act) and a target speed (n_eng_des) of the internal combustion engine or a derived quantity (n_pump_des), and - Adding the correction control pressure (p_ctrl_speed) to the pre-controlled control pressure, where the sum is a final control pressure (p_ctrl). [11] Method according to claim 10 with the preceding step: - Initiation of braking by a driver request or due to exceeding an actual rotational speed (n_eng_act) of the internal combustion engine or a derived quantity (n_pump_act) or due to exceeding a driving speed (v_veh) or a derived quantity (n_mot_act). [12] Method according to one of claims 10 or 11 comprising the step: - Controlling the braking torque by adjusting a swivel angle (angle_mot) of the secondary unit (2).

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