Hydrostatic transmission and methods for braking with it

Abstract

A hydrostatic transmission for a drive system, comprising a drive shaft (4) of a hydrostatic primary unit (1) whose stroke volume is adjustable by control with a variable actuating pressure (p_st) and at least one hydrostatic secondary unit (2) that can be coupled to an output of the drive system, wherein the primary unit (1) and the at least one secondary unit (2) are connected to each other via two working lines (10) of a closed hydraulic circuit, and wherein the primary unit (1) is preferably an axial piston machine with an adjustable swivel angle (angle_pump) which can be controlled by an electrical control unit (8) when the hydrostatic transmission is braked, characterized in that forces act in the direction of a reduction of its swivel angle (angle_pump) during pump operation of the primary unit (1).which depend on a pressure difference (Δp) of the two working lines (10) and a rotational speed (n_pump_act) of the drive shaft (4) and the swivel angle (angle_pump), wherein a characteristic map (36) of the primary unit (1) is stored in the control unit (8), in which a respective swivel angle (angle_pump) or a respective stroke volume (Vg_pump) of the primary unit (1) is assigned to the actuating pressure (p_st), the pressure difference (Δp) and the rotational speed (n_pump_act).

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 at least one hydrostatic motor (secondary unit) are fluidically connected to each other via a closed hydraulic circuit. A primary unit can be located in a closed hydraulic circuit with several secondary units, the secondary units being arranged in parallel to each other. 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] 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, thereby protecting the internal combustion engine from overspeeding. 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 affected high-pressure line. Initially, a slightly higher engine speed than the permissible limit is accepted.

[0006] The respective primary units of the aforementioned prior art are angle-controlled. For this, they require feedback on their current angle of rotation. Furthermore, within the framework of the angle feedforward control, a generally higher angle of rotation than the ideal angle is revealed.

[0007] Furthermore, patent specification DE 103 03 206 A1 is known from the prior art.

[0008] In contrast, the invention is based on the objective of creating a hydrostatic transmission and a method in which the swivel angle and thus the stroke volume (in braking operation, this is a displacement volume) of the primary unit is optimally controlled or regulated, so that the speed of the internal combustion engine is brought up to its maximum speed as quickly as possible and maintained as close as possible to this maximum speed. The internal combustion engine should never exceed its maximum speed.

[0009] Furthermore, the technical complexity of the primary unit should be 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 11.

[0011] The claimed hydrostatic transmission is designed for a drive system comprising an internal combustion engine, e.g., a diesel engine, and an output, e.g., an axle or a wheel. The hydrostatic transmission has a drive shaft of a primary unit, which can be coupled to the internal combustion engine of the drive system and operates as a pump during traction, and at least one secondary unit, which can be coupled to an output of the drive system and operates as a motor during traction. One or more secondary units can be assigned to a primary unit; for example, four secondary units on two axles and four wheels in a field sprayer, or one secondary unit in a forklift. The primary unit and the secondary unit(s) are fluidically connected to each other via two working lines of a closed hydraulic circuit. The primary unit has an adjustable stroke volume and is preferably an axial piston motor with an adjustable swivel angle.The hydrostatic transmission further comprises an electrical control unit that allows for the control or regulation of braking, whereby a braking torque of the secondary unit(s), acting as a pump, is supported at the drive shaft of the primary unit, acting as a motor. With regard to the entire drive system, the braking torque of the output is thus supported at the internal combustion engine via the hydrostatic units. According to the invention, during pump operation, forces act on the drive mechanism of the primary unit in the direction of a reduction in its swivel angle. These forces depend on a pressure difference between the two working lines, the rotational speed of the drive shaft, and the swivel angle of the primary unit. This gives the primary unit a load-sensitive behavior.Furthermore, according to the invention, a corresponding characteristic map is stored in the control unit, in which a respective swivel angle or stroke volume is assigned to various pressure differentials and different rotational speeds, which can be controlled by the control unit during braking. This ensures that the combustion engine never over-revs. Furthermore, due to the consideration of the characteristic map, no feedback of the swivel angle of the primary unit is required. Thus, instead of position-controlled primary units, such as those found in agricultural machinery, load-sensing primary units, which are usually somewhat less expensive, are used according to the invention. These can also be used in agricultural machinery, for example.

[0012] In a load-sensitive primary unit, the adjustment device for the stroke volume can be designed such that an actuating piston of the adjustment device is subjected to a control pressure, specified, for example, by a proportionally adjustable pressure reducing valve, depending on a setpoint signal. Since the restoring force against which the actuating piston must work depends not only on the force of one or more restoring springs, which increases with increasing stroke volume, but also on the operating pressure and the rotational speed, a specific actuation of the pressure reducing valve is not assigned to a single, specific stroke volume.

[0013] With regard to the hydrostatic transmission according to the invention, which does not contain the internal combustion engine, the speed of the crankshaft of the internal combustion engine to be limited can be transmitted to the control unit via a signal input or taken directly from the drive shaft of the primary unit if the crankshaft and the drive shaft are one piece, or can also be calculated from a speed of the drive shaft of the primary unit if a mechanical transmission stage is provided between the crankshaft and the drive shaft.

[0014] Further advantageous embodiments of the invention are described in the dependent patent claims.

[0015] It is particularly preferred if the control unit can control a control pressure in the direction of increasing the swivel angle and thus the displacement volume of the primary unit via an actuating cylinder of an adjusting device. The characteristic map specifies the dependence of this control pressure, or a deviation thereof, on the pressure difference between the two working lines, the rotational speed of the drive shaft, and the swivel angle or displacement volume of the primary unit.

[0016] In a device-technically simple variant, the actuating cylinder is single-acting against a spring, so that the actuating pressure only acts in the direction of increasing the swivel angle and the stroke volume, while the spring acts in the direction of decreasing the swivel angle and the stroke volume.

[0017] In a preferred embodiment of the hydrostatic transmission according to the invention, the swivel angle and the stroke volume of the primary unit are adjustable on both sides of a neutral position. Accordingly, the pressure differentials of the two working lines and the rotational speeds of the drive shaft for the swivel angles and / or stroke volumes on both sides of the neutral position are stored in the characteristic map. This allows the affected drive system to be used in both directions of travel of the mobile machine in towing operation, with the direction of rotation of the combustion engine remaining constant, and to be braked accordingly in both directions according to the invention.

[0018] In a double-acting actuator cylinder, whose two actuating pressure chambers can be pressurized with the actuating pressure or the actuating pressure deviation via the common or a separate electrically adjustable actuating pressure valve, the affected drive system can be used in both directions of travel of the mobile working machine in towing mode while maintaining the same direction of rotation of the internal combustion engine, and can be braked accordingly in both directions according to the invention. A double-sided spring arrangement is provided, which acts in the direction of a central position in which the pivot angle and the stroke volume are zero.

[0019] 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.

[0020] 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.

[0021] Preferably, a safety correction value for the swivel angle or the displacement volume is assigned to or superimposed on the characteristic map, which reduces the swivel angle and the displacement volume, or which is subtracted from the swivel angle. In particular, the safety correction value can be included in the characteristic map. This serves to compensate for inaccuracies in the characteristic map, especially due to variations (manufacturing inaccuracies) of the primary unit, with the aim of ensuring that the internal combustion engine does not over-rev when braking begins.

[0022] In a particularly preferred embodiment, the hydrostatic drive according to the invention has a controller whose input is a speed difference between, on the one hand, an 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 limit speed, which can be the maximum permissible speed of the internal combustion engine or a desired speed lower than the maximum permissible speed, or a value derived therefrom. The output of the controller is a correction value of the swivel angle or the stroke volume, which is added to a pre-controlled swivel angle or stroke volume.Since the pre-controlled value is always below the maximum speed of the internal combustion engine due to the safety correction value, this ensures that the internal combustion engine does not over-rev at the start of braking, and that the controller then allows for a rapid and precise approximation of the actual speed to the maximum speed of the internal combustion engine. Preferably, the controller is a PID controller. Preferably, the control unit incorporates the controller.

[0023] 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, for example a brake pedal, in particular via an input for a signal line of the control element, or on the basis of automatic monitoring of the travel speed of the mobile working machine, in particular an input for a signal from a speedometer, or on the basis of automatic monitoring of the speed of the internal combustion engine or the quantity derived therefrom.

[0024] Preferably, the 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.

[0025] The adjusting device can be one in which a control pressure in an actuating cylinder is specified by an electromagnetically proportionally adjustable pressure control valve, for example an EV adjustment or an ET adjustment.

[0026] 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.

[0027] The ET control system has two pressure regulating valves. One pressure regulating valve controls the actuating pressure in one actuating chamber, and the other pressure regulating valve controls the actuating pressure in the other actuating chamber of the actuator cylinder.

[0028] One possible method for controlling or regulating braking with a previously described hydrostatic transmission involves the following steps: - Swiveling the primary unit back to a small swivel angle or a small stroke volume, e.g. to about 10% of the maximum swivel angle or maximum stroke volume, and - Pre-control of the swivel angle or the stroke volume based on the characteristic map.

[0029] The two steps can be performed in the order mentioned, thus preventing the combustion engine from over-revving with the highest possible reliability. Alternatively, the two steps can be performed simultaneously, saving time and allowing braking to be initiated quickly.

[0030] Preferably, the actuating pressure for the smaller swivel angle is determined from the characteristic map, and the swiveling back is triggered directly by applying the determined actuating pressure.

[0031] In a particularly preferred further development of the method, the swivel angle or the stroke volume is controlled.

[0032] Preferably, the control is achieved by determining a corresponding correction value as a function of a speed difference between, on the one hand, an actual speed of the internal combustion engine or a quantity derived therefrom, in particular the actual speed of the primary unit, and, on the other hand, a maximum speed of the internal combustion engine or a quantity derived therefrom, and by adding the correction value to a pre-controlled swivel angle or to a pre-controlled stroke volume.

[0033] In a preferred embodiment of the inventive method, braking is initiated in advance by a driver request or due to an exceedance of the internal combustion engine speed (actual speed exceeds a maximum permissible speed) or due to an exceedance of a driving speed (driving speed exceeds a desired or a maximum permissible driving speed).

[0034] An embodiment of the hydrostatic transmission and drive according to the invention are shown in the drawings. The invention will now be explained in more detail with reference to the figures in these drawings.

[0035] 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 characteristic map of the primary unit of the hydrostatic transmission from Fig. 1, Fig. 4 a schematic overview of the map-based feedforward control with additional control of the stroke volume of the primary unit during braking, Fig. 5 Two diagrams of map-based feedforward control without additional control during braking and Fig. 6 Two diagrams of map-based feedforward control with additional control during braking.

[0036] 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. The primary unit 1 is an axial piston machine to whose drive shaft 4 a crankshaft of a (not shown) diesel engine 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.

[0037] 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 the mobile working machine.

[0038] 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 work machine. During operation, the drive shaft 4 of primary unit 1 serves as the drive shaft and primary unit 1 functions as a pump, while secondary unit 2 operates as the engine, and the drive shaft 12 of secondary unit 2 is an output shaft.

[0039] 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, via its friction and acceleration forces of the pistons, dissipates at least part of the braking energy of the mobile working machine.

[0040] Each working line 10 is equipped with a pressure relief valve 14, through which the respective working line 10 can be relieved of pressure to a feed line 16. This feed line is filled with feed fluid from a tank T by a feed pump 18, which is rotationally fixed to the drive shaft 4 of the primary unit 1. Furthermore, the feed line 16 can be relieved of pressure to a tank T via a pressure relief valve 20.

[0041] The feed line 16 is connected to the two working lines 10 via a spring-loaded check valve 22, so that the 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 towards the respective working line 10.

[0042] The displacement volume of the primary unit 1 is adjusted via an adjusting device 24. This device has a double-acting actuating cylinder 26, whose two pressure chambers act against each other on an actuating piston that is coupled to a swashplate of the axial piston machine 1. Each of the two 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) by which the piston of the actuating cylinder 26 and the swashplate of the primary unit 1 are biased to a central position. From there, 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 allows for forward driving and forward braking, as well as reverse driving and reverse braking.

[0043] 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.

[0044] Fig. Figure 2 shows in a diagram the basic temporal profile of the stroke volumes Vg_pump and Vg_mot of the primary unit 1 and the secondary unit 2, first during a transition from driving mode to braking mode and then during a transition from braking mode back to driving mode. Such 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 speed of the diesel engine which is borderline high for the load of the diesel engine or a target speed which should not be exceeded and is lower than the limiting speed for reasons of comfort, for example, and which was determined by the speed sensor 6 and the control unit 8, c) a limiting travel speed v_veh of the mobile working machine, which was indirectly determined by a speed sensor 34 of the drive shaft 12 of the secondary unit 2 and transmitted to the control unit 8.

[0045] The braking process is controlled by the control unit 8. To achieve this, the swivel angle angle_pump, and thus the displacement volume Vg_pump of the primary unit 1, is quickly reduced to a low value. This low value can, for example, be approximately 10% of the maximum swivel angle angle_pump_max or the maximum displacement volume Vg_pump_max of the primary unit 1. This prevents the diesel engine from over-revving, as its inertia prevents it from accelerating to an excessive speed during the short time it takes for the primary unit 1 to swivel back. Consequently, the pressure in the now high-pressure working line 10 rises rapidly, and the corresponding high-pressure valve 14 opens a connection to the feed line 16.

[0046] Subsequently, the secondary unit 2 is controlled by the control unit 8 via the adjusting device 30 such that its swivel angle and thus its displacement volume Vg_mot are increased. The swivel angle angle_pump of the primary unit 1 is also slightly increased again (not shown), thereby accelerating the diesel engine and dissipating some of the braking power via this engine. To maximize this engine speed without exceeding a permissible maximum speed, the control system explained with reference to the following figures is used.

[0047] Fig. Figure 3 shows a characteristic map 36 of the primary unit 1, which according to the invention is designed as a load-sensing unit. The characteristic map 36 represents the influence of the pressure difference Δp between the two working lines 10 and the actual rotational speed n_pump_act and, furthermore, the swivel angle angle_pump of the primary unit 1 on the actuator in the actuating cylinder 24 (see Figure 3). Fig. 1) prevailing control pressure p_st. More precisely, the relationship is represented such that the pressure difference Δp of the two working lines 10 is plotted along a first axis, where, for example, the positive pressure differences Δp represent the overrun operation and the negative pressure differences Δp represent the braking according to the invention. On a further axis, the actual rotational speed of the drive shaft 4 n_pump_act of the primary unit 1 is plotted. The different areas of the shown set of areas apply by way of example to three positive and three negative swivel angles angle_pump of the swashplate of the primary unit 1. On the vertical axis, a control pressure deviation p_st_comp is plotted, which is necessary to maintain the swivel angle angle_pump at the respective operating point. In the example mentioned above, the positive pressure values ​​represent overrun operation, while the negative pressure values ​​apply to the braking according to the invention.This means that when initiating braking in relation to the . Fig. As explained in section 1, adjusting device 24 requires a change of the pressure chamber pressurized with actuating pressure medium p_st.

[0048] The characteristic map 36 is in the control unit 8 (see Fig. 1) such that, by means of a pressure sensor (not shown) provided on each of the working lines 10, the control unit 8 can control any desired swivel angle angle_pump by adjusting the corresponding control pressure deviation p_st_comp by appropriately actuating the control pressure valves 28. In contrast to prior art hydrostatic transmissions, no feedback of the swivel angle angle_pump from the primary unit 1 is required. Pressure sensors can also be omitted. The pressure that develops in the corresponding working line can be determined from the opening pressure and the characteristic curve of a pressure relief valve 14.

[0049] Due to variations in the manufacturing of primary units 1 and other factors, such as wear and viscosity of the pressure medium used, small deviations of the various specific primary units 1 from the characteristic map 36 are possible, so that the map-based control described above is used as feedforward control and additionally according to Fig. 4 is refined. In the middle area, the characteristic map 36 is from Fig. Figure 3 shows how this is used to assign a corresponding correction value p_st_comp to the control pressure p_st. In map 36, a safety correction value is also included to account for the aforementioned possible variations in the behavior of different primary units 1 and to always set a displacement volume Vg_pump_comp through the device-technically simple feedforward control using map 36, at which the diesel engine is reliably not over-revved and also dissipates a high but non-critical braking force in the event of a failure of the functions described below.

[0050] Furthermore, a control circuit is superimposed on the feedforward control, which compares the actual speed n_eng_act of the diesel engine with a limit speed n_eng_max, which is the maximum permissible speed of the diesel engine or a predefined maximum speed that is lower than the maximum permissible speed, and generates a further correction value for the control pressure p_st. Alternatively, the speed n_pump_act of primary unit 1 can also be compared with a converted maximum permissible speed n_pump_max of primary unit 1, and the further correction value for the control pressure p_st can be generated.

[0051] Fig. Figure 5 shows the pure feedforward control according to the characteristic map 36 (see Figure 5). Fig. 3 and Fig. 4) The primary unit 1 is thus already pivoted close to the ideal value. In the example shown, however, the primary unit 1 is still set to a slightly too large displacement volume Vg_pump (swallowing volume), which would lead to a slightly too high rotational speed n_eng_act of the diesel engine. In the case of a pivot angle that is too small angle_pump, the braking capacity of the diesel engine would only be partially utilized. While a reliable braking function is already fundamentally possible with the described load-sensitive primary unit 1, it can be further improved.

[0052] The final deviation from the ideal value is ultimately balanced by the superimposed control with the values ​​in Fig. The PID controller 38 shown in Figure 4 uses the "last" actual speed n_eng_act of the diesel engine from its maximum permissible speed n_eng_max as an input and outputs an increased displacement volume Vg_pump_add_incr. This determines an optimal displacement volume Vg_pump_brake_corr for braking under the conditions of the specific primary units 1.

[0053] The impact is exemplified in Fig. Figure 6 illustrates this. This results in a consistent load on the diesel engine during braking without it reaching the impermissible speed range.

[0054] A hydrostatic transmission has been revealed that enables braking in which at least one traction motor, acting as a pump and with an adjustable stroke volume, is supported via a closed hydraulic circuit by an adjustable axial piston pump, which in turn acts as a motor during braking and can be supported by an internal combustion engine. To prevent over-revving the internal combustion engine, a control unit can adjust the pump's swivel angle according to a characteristic map. This map defines the relationship between an actuating force and a pressure differential between the two working lines of the closed hydraulic circuit, a pump speed, and the swivel angle. This allows for map-based control of the pump's swivel angle, eliminating the need for feedback on the current swivel angle.The swivel angle of the pump is chosen so that the combustion engine is not over-revved or does not exceed a desired speed that is less than the maximum permissible speed of the combustion engine.

[0055] This control alone, however, is often insufficient to set the desired pump displacement with sufficient accuracy and thus fully utilize the engine's support torque for braking, as the engine force map is subject to a certain tolerance. Therefore, the map-based setting of the pump's swivel angle is advantageously used as feedforward control for subsequent control of the swivel angle as a function of the engine's rotational speed. This feedforward control is superimposed with a control circuit that compares the actual rotational speed of the combustion engine (nD_ist), which is typically a diesel engine, with its maximum permissible rotational speed (nD_max) or a desired rotational speed lower than the engine's maximum permissible speed, and generates a further correction value for the actuator pressure.Since the feedforward control sets a swivel angle at which the combustion engine will not over-rev, the swivel angle and the stroke volume are increased by the control.

[0056] The invention will now be presented again with a slightly different focus and according to a simplified method: When an electrical control unit detects that high-performance braking is necessary, the displacement of the hydrostatic primary unit (the pump) is adjusted so that only as much power is delivered to the combustion engine as it can safely handle. Since, during high-performance braking, a portion of the braking energy is dissipated via one of the two pressure-limiting valves, the high pressure is fixed within certain limits. The flatter the characteristic curve of the pressure-limiting valves, the more accurately the high pressure can be determined. Consequently, the torque delivered to the crankshaft by the pump, which is acting as a motor, depends, to a good approximation, only on the pump's oscillation angle.

[0057] The special feature of a load-sensitive pump is that its actual swivel angle represents an equilibrium state of the following main influencing factors: - Adjustment force generated by the actuating pressure p_st; - Adjustment force dependent on the speed; - Adjustment force depends on the pressure difference across the pump.

[0058] This is a significant technical difference compared to a position-controlled pump control, where there is a linear relationship between the control signal and the pump swivel angle.

[0059] In addition, other influencing factors such as manufacturing tolerances, aging, oil temperatures, viscosities, etc., complicate the behavior of a pump, especially the adjustment of a load-sensitive pump to a specific swivel angle.

[0060] Since a swivel angle sensor is to be omitted for cost reasons, the challenge now arises of dissipating some of the braking energy via the combustion engine without over-revving it. It must also often be considered that, depending on the application and operating strategy, sometimes the maximum braking power of the combustion engine should be utilized, and sometimes, for example, for noise reasons, only a portion of the combustion engine's braking power is required. In the latter case, the combustion engine's speed is then lower than the maximum permissible speed.

[0061] With a position-controlled pump, specifying different limit speeds of the combustion engine is not particularly difficult, as the following applies: torque = f(swivel angle) = f(control signal). However, with a load-sensing pump, the challenge is considerable. Here, the aforementioned influencing factors must be taken into account to set the correct control pressure p_st for the desired swivel angle.

[0062] The following describes how this can be done relatively easily with a load-sensitive hydrostatic primary unit whose stroke volume is adjustable.

[0063] The operating strategy determines how much braking power is to be dissipated from the combustion engine. For every combustion engine, there is a characteristic relationship: braking power equals 2π * rotational speed * torque. Since in most cases the rotational speed—either a maximum permissible speed or a limit speed desired by the operating strategy—is predetermined, the torque can be calculated from the above relationship. This means that for the hydrostatic primary unit, a specific swivel angle must be set during high-performance braking operation, based on the known system pressure level (the pressure difference between the two working lines, or, since the low pressure is largely predetermined, the pressure in the high-pressure working line).

[0064] In the case of the load-sensitive pump, the following procedure is followed: The dependencies mentioned above are taken into account. a) A swashplate control system is used. Based on the target engine speed for braking, an experimentally or theoretically determined characteristic curve is used to calculate a control pressure p_Ctrl_init for the pump. This pressure results in a swashplate angle at which the target engine speed is not yet reached. In the simplest case, if the target engine speed is always the same during braking, for example, always the maximum engine speed, the control pressure is not determined using a characteristic map or curve. Instead, the same control pressure is always selected, which reliably results in a swashplate angle at which the engine speed (non-critical base overrun speed) is reliably below the target speed. b) A speed control is superimposed on the feedforward control, which ensures that the support power of the combustion engine is utilized up to the defined target speed. This is achieved by calculating the control pressure p_Ctrl_speed from the difference between the target and actual speed of the combustion engine using a controller (P, PI or PID controller). c) Depending on the application, an additional dynamic component p_ctrl_dyn is added to accelerate or assist the return to or holding of the pump. This dynamic component can, for example, depend on the estimated flow rate that passes through the pressure relief valves and is converted into the control pressure component p_Ctrl_dyn using a factor K.

[0065] The sum of the individual components yields the control pressure p_Ctrl, which is directed to one or the other control chamber of the pump depending on the direction of travel. Generally, "negative" control pressures also occur. This means that when braking while traveling forward, the control chamber for reverse travel must be energized, and vice versa, in order to maintain the pump's oscillation angle at the desired operating point. Reference symbol list: 1 hydrostatic primary unit 2 hydrostatic secondary units 4 Drive shaft 6 Speed ​​sensor 8 Control unit 10 Work Management 12 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 36 characteristic map 38 regulators angle_pump Primary unit swivel angle angle_pump_max maximum swivel angle of the primary unit n_eng_act Actual speed of the internal 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_act Actual rotational speed of the primary unit n_pump_max Primary unit speed derived from the maximum speed of the internal combustion engine p_st Primary unit control pressure p_st_comp Primary unit control pressure deviation Vg_mot secondary unit displacement Vg_pump Primary unit displacement volume Vg_pump_add_incr increased stroke volume of the primary unit for braking Vg_pump_brake: Pre-controlled stroke volume of the primary unit for braking Vg_pump_brake_corr Target stroke volume of the primary unit for braking Vg_pump_comp Primary unit displacement deviation Vg_pump_max maximum stroke volume of the primary unit v_veh Driving speed Δn_eng Speed ​​difference between the target speed and the actual speed of the internal combustion engine Δp pressure difference between the two working lines T Tank

Claims

[1] Hydrostatic transmission for a drive system, comprising a drive shaft (4) of a hydrostatic primary unit (1) whose stroke volume can be coupled to an internal combustion engine of the drive system, and at least one hydrostatic secondary unit (2) that can be coupled to an output of the drive system, wherein the primary unit (1) and the at least one secondary unit (2) are connected to each other via two working lines (10) of a closed hydraulic circuit, and wherein the primary unit (1) is preferably an axial piston machine with an adjustable swivel angle (angle_pump) which can be controlled by an electrical control unit (8) when braking the hydrostatic transmission. characterized by, that in the pumping operation of the primary unit (1) forces act in the direction of a reduction of its swivel angle (angle_pump) which depend on a pressure difference (Δp) of the two working lines (10) and a rotational speed (n_pump_act) of the drive shaft (4) and the swivel angle (angle_pump), wherein in the control unit (8) a characteristic map (36) of the primary unit (1) is stored in which the actuating pressure (p_st), the pressure difference (Δp) and the rotational speed (n_pump_act) are assigned a respective swivel angle (angle_pump) or a respective stroke volume (Vg_pump) of the primary unit (1). [2] Hydrostatic transmission according to claim 1, wherein the control unit (8) can control a control pressure (p_st) via an electric control pressure valve (28) and a control cylinder (26) of an adjusting device (24) of the primary unit (1), and wherein the control pressure (p_st) acts in the direction of an increase in the swivel angle (angle_pump) of the primary unit (1), and wherein the dependence of the control pressure (p_st) or a control pressure deviation (p_st_comp) on the pressure difference (Δp) of the two working lines (10) and the rotational speed (n_pump_act) of the drive shaft (4) and the swivel angle (angle_pump) or the displacement volume (Vg_pump) is stored in the characteristic map (36). [3] Hydrostatic transmission according to one of the preceding claims, wherein the swivel angle (angle_pump) and the stroke volume (Vg_pump) of the primary unit (1) are adjustable on both sides of a zero position and wherein the control pressures (p_st), the pressure differences (Δp) between the two working lines (10) and the rotational speeds (n_pump_act) of the drive shaft (4) for the swivel angles (angle_pump) or the stroke volumes (Vg_pump) on both sides of the zero position are stored in the characteristic map (36). [4] Hydrostatic transmission according to claims 2 and 3, wherein the actuating cylinder (26) is double-acting and has two actuating pressure chambers which can be actuated with the actuating pressure (p_st) from the control unit (8) via the common or a respective electrically adjustable actuating pressure valve (28). [5] 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), and wherein the braking is a high-performance braking system in which a first part of the braking power can be reduced via one of the pressure relief valves (14), while a second part of the braking power can be reduced via the primary unit (1). [6] Hydrostatic transmission according to claim 5, wherein the pressure limiting valves (14) each have a flat characteristic curve with respect to their pressure difference as a function of their volume flow rate. [7] Hydrostatic transmission according to one of the preceding claims, wherein a safety correction value of the swivel angle (angle_pump) or the stroke volume (Vg_pump) is assigned or superimposed on the characteristic map (36), which causes a reduction of the swivel angle (angle_pump) or the stroke volume (Vg_pump). [8] Hydrostatic transmission according to one of the preceding claims with a controller (38) whose input variable is a speed difference (Δn_eng) between an actual speed (n_eng_act) of the internal combustion engine or a quantity derived therefrom (n_pump_act) and a limit speed (n_eng_max) of the internal combustion engine or a quantity derived therefrom (n_pump_max), and whose output variable is a correction value (Vg_pump_add_incr) for the swivel angle (angle_pump_brake) controlled via the map (36) or for the stroke volume (Vg_pump-brake) controlled via the map (36), which can be added to it. [9] 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 the mobile working machine or on the basis of automatic monitoring of a travel speed (v_veh) of the mobile working machine or a quantity derived therefrom (n_mot_act) or on the basis of automatic monitoring of the speed (n_eng_act) of the internal combustion engine or a quantity derived therefrom (n_pump_act). [10] 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. [11] Method for controlling or regulating a braking system with a hydrostatic transmission according to any of the preceding claims, comprising the step: - Swiveling back the primary unit (1) and thus pre-controlling the swivel angle (angle_pump) or the stroke volume (Vg_pump) based on the characteristic map (36). [12] The method of claim 11 comprising the step: - Rules of the swivel angle (angle_pump) or the stroke volume (Vg_pump). [13] Method according to claim 12, wherein the rules are: - Determination of a corresponding correction value (Vg_pump_add_incr) depending on 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 maximum speed (n_eng_max) of the internal combustion engine or a derived quantity (n_pump_max) and by - The correction value (Vg_pump_add_incr) is added to a pre-controlled swivel angle (eng_pump_brake) or to a pre-controlled stroke volume (VG_pump_brake). [14] Method according to claim 13 with the preceding step: - Initiation of braking by a driver request or due to an exceedance of the actual rotational speed (n_eng_act) of the internal combustion engine or the quantity derived therefrom (n_pump_act) or due to an exceedance of a driving speed (v_veh) or a quantity derived therefrom (n_mot_act).

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