BRAKE DEVICE, IN PARTICULAR FOR ELECTRICALLY DRIVEN MOTOR VEHICLES

Abstract

A brake apparatus, for electrically driven motor vehicles, includes a traction motor at an axle of a vehicle, which traction motor is used both as drive motor and as brake system with recuperation of brake energy, a first piston-cylinder unit, which is actuatable by means of an actuating device, in particular brake pedal, a second piston-cylinder unit, which is actuatable by means of an electromotive drive and a non-hydraulic gearing apparatus, in particular spindle drive. The piston-cylinder units are connected via hydraulic connecting lines to wheel brakes of the motor vehicle. A pressure chamber of the first piston-cylinder unit is connected to two wheel brakes of a vehicle axle, and a pressure chamber of the second piston-cylinder unit is connected to a vehicle axle for active brake force feedback control and recuperation control in interaction with the traction motor.

Description

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a Continuation of co-pending U.S. patent application Ser. No. 16/615,547, filed Nov. 21, 2019 which is a Section 371 U.S. National Stage Filing of International Application No. PCT/EP2018/063274, filed May 22, 2018, which was published in the German language on Nov. 29, 2018, under International Publication No. WO 2018/215397 A1, which claims priority to German Patent Application No. 10 2017 111 077.1, filed on May 22, 2017, the disclosures of which are incorporated herein by reference.

BACKGROUND

The invention relates to a brake apparatus or a brake system, in particular for electrically driven motor vehicles.

Such brake systems are already known, for example from EP 1 907 253 B1. In the case of said brake system, a first piston-cylinder unit is provided which is actuatable by means of an actuating device. Said actuating device has, on the one hand, a device or brake pedal that can be actuated by the driver, and a device that can be actuated by an electromotive drive. The two pressure chambers of the piston-cylinder unit are assigned to in each case one vehicle axle or one brake circuit and are connected to the wheel brakes via inlet/outlet valves assigned to the wheel brakes. Furthermore, said document has already disclosed solutions for a fall-back level and various concepts for pressure feedback control based on travel and electrical current.

DE 10 2005 055 751 has furthermore disclosed a brake system with pressure-volume control or pressure gradient control by means of the pressure-volume characteristic curve.

DE 10 2012 002 791 A1 has also disclosed a brake system, having a first piston-cylinder unit or master cylinder with an actuating device that is actuatable by the driver, and having a second piston-cylinder unit, which is driven by an electromotive actuating device. Here, by means of isolating valves provided in the hydraulic lines to the wheel brakes, the first piston-cylinder unit can be separated from the brake circuits, such that only the second piston-cylinder unit acts on the brake circuits.

One design solution of an electromotively driven piston-cylinder unit is known for example from PCT/EP2013/057609.

In said solution, the rotor of the motor is mounted unilaterally in the motor housing.

In automobile racing, aside from the Formula 1 championship, the Formula E championship, to which particular regulations and technical requirements apply, has existed since the year 2014. Said requirements place particular restrictions on developers and constructors and present them with new challenges; in particular, highly precise braking torque feedback control of traction motor and electrohydraulic brake is required.

The known brake systems have various disadvantages which make them appear unsuitable or non-optimal for use in motor racing, in particular in vehicles in the Formula E championship.

SUMMARY OF THE DISCLOSURE

It is an object of the invention to create a brake system for vehicles which satisfies at least the prerequisites and conditions of Formula E vehicles, that is to say inter alia vehicles with a high-power traction motor (100-300 kW) at one axle, and, at the same time, to permit in particular highly precise braking feedback control and advantageous recuperation and optimized deceleration of the vehicle in accordance with demand.

Said object may be achieved with a brake system having the features of the various claims.

For this purpose, in a first axle, a pressure is generated in the master brake cylinder, which generates a braking torque, exclusively by means of the actuating force of the driver. In the second axle, it is the intention for feedback control of the braking torque to be performed through a combination of braking power of the traction motor and active pressure feedback control of an electrically driven piston-cylinder unit.

With the solution according to the invention or the embodiments thereof described below, an ideal recuperation and braking feedback control system for vehicles with a high-power electric motor (100-300 kW) is created. It is thus highly advantageously possible to implement innovative braking management, wherein a distribution of the braking action between the traction motor and the electromotively driven second piston-cylinder unit (EHB axle module) is realized.

It is expedient here for a setpoint pressure (p_setpoint) and the pressure change (dp/dt) to be adjusted through travel control of the piston of the second piston-cylinder unit. In accordance with the characteristic of the brake system, the pressure leads to a braking torque and a deceleration of the hydraulic system and a deceleration of the traction motor. The overall deceleration a_total is determined as the sum of the deceleration of the EHB axle module a_EHB and of the traction motor a_TM.

It is advantageously the case that, in the presence of low pressures, through decoupling, the braking is performed only by means of the traction motor, such that maximum recuperation is attained. In the fall-back level (RFE), in the event of failure of the second piston-cylinder unit, the braking is performed by means of the traction motor and the first piston-cylinder unit, in particular by means of two circuits or by means of one circuit. In the case of a one-circuit fall-back level (FIG. 2a), the traction motor is utilized in the fall-back level for deceleration at one axle.

Highly precise pressure feedback control in accordance with demand is performed by means of the second piston-cylinder unit (EHB axle module). Here, it is expedient for the pressure-volume characteristic curve (DVK) to be used for the travel control of the piston of the second piston-cylinder unit. It is thus possible to attain ideal pressure feedback control through fast attainment of the target pressure by pilot control by means of the pressure-volume characteristic curve and feedback control of the pressure change, such that an advantageous adjustment to the traction motor braking torque feedback control is realized.

A maximization of the pressure build-up dynamics is advantageously realized by assistance of the electric motor of the second piston-cylinder unit (EHB module) in the lower pressure range. In this way, a very fast onset of braking is possible, which is important in particular in emergency braking operations. Time-to-lock (TTL) values of <100 ms are thus achieved. The typical TTL time in classic systems in the passenger motor vehicle sector with 12 V is 150 ms. The dynamics are ultimately restricted by the chassis. Since these restrictions have less of a limiting effect during racing operation, shorter times can be implemented. Also, a braking system with 48 V may be used in order to further increase the dynamics. These advantages can expediently be utilized such that braking is performed later and the vehicle can thus be operated at maximum speed for longer, which has a significant influence on lap time. Also, modern intra-vehicle distance feedback control systems and emergency braking systems can be implemented in order to react quickly to braking operations of a vehicle travelling in front and avoid the risk of an accident. This is highly important in particular during races in urban areas.

The pressure gradient feedback control may advantageously also be used in the context of an optimized deceleration which is variable over time, in order to likewise optimize the lap time or for intelligent inter-vehicle distance feedback control, with which the burden on the driver is reduced, for example through active utilization of camera sensors and controlled braking processes in particular during cornering.

Since very precise and highly dynamic pressure feedback control and/or braking deceleration feedback control is highly advantageous for racing operation, various sensors and characteristic maps are evaluated in order to react very quickly to changes. This is based on highly dynamic and precise travel control of the piston.

Here, for the travel control of the piston, the pressure-volume characteristic curve of the wheel brake and the pressure at the axle or alternatively the pressure of the piston-cylinder unit of the EHB axle module is evaluated. The pressure-volume characteristic curve changes for example as a result of air in the hydraulic system. An ongoing adaptation is necessary for this purpose. For the adaptation, a pressure transducer is used in order to adapt a corresponding assignment of piston travel to pressure.

The piston position is expediently calculated by means of an angle encoder a of the electric motor. As a further sensor for optimum highly precise feedback control, the phase current i of the electric motor and the temperature T of the electric motor or of the piston-cylinder unit are evaluated. The phase current is utilized such that, by means of the torque constant kt that replicates the relationship between phase current and torque of the electric motor, a torque can be set which, owing to a constant cross section of the piston-cylinder unit, correlates with a pressure. For this purpose, the losses in the torque transmission (for example efficiency of the gearing, mechanical losses) must be known or determined. This may be performed by means of trimming using a pressure sensor. By means of the evaluation of the electrical current, it is possible, through intelligent pilot control, to very quickly adjust to a pressure that approximately corresponds to the desired target pressure, and the delay times of the inert pressure measurement by the pressure transducer are thus compensated, that is to say the pressure is adjusted to in a substantially exact manner by means of phase current feedback control and, after a time delay, is validated by means of the pressure encoder.

By means of the temperature sensor, changes in the hydraulic system (changes in viscosity in the fluid) and a change in the torque constant kt as a result of warming of the electric motor are determined.

By means of the temperature information, the viscosity in the hydraulic system, which varies owing to the influence of temperature, can be replicated and utilized for adapted travel control of the EHB axle module. This is highly significant in particular for the pressure gradient feedback control, because a different pressure difference must be set by the EHB axle module in order, in the case of varying viscosity, to attain the same pressure gradient by means of a greater throttling action. The reason for this is that the pressure gradient is determined by the pressure difference between EHB actuator and wheel brake and the throttling action.

The temperature may also be utilized in order to detect a change in the brake system (for example fading effect). In the case of fading, the braking action varies in a manner dependent on the set pressure as a result of warming of the wheel brake, that is to say, for a desired constant braking deceleration, a higher pressure must be set in the event of fading. This information can advantageously be utilized in order to create a characteristic map in order to optimize the dependency of the braking action in dependence on the pressure in the event of a variation of the brake system, for example owing to varying temperatures. This characteristic map can likewise be used in addition to the pressure-volume characteristic curve for highly exact braking feedback control under different conditions. This is of very great significance in particular in motor racing, because the temperatures vary highly dynamically during operation.

Further advantages of the invention and the embodiments thereof will emerge from the subclaims and the description of the exemplary embodiments.

BRIEF DESCRIPTION OF THE DRAWINGS

The foregoing summary, as well as the following detailed description of the preferred invention, will be better understood when read in conjunction with the appended drawings:

FIG. 1a: schematically shows a brake system for a motor vehicle with high-power traction motor;

FIG. 1b: is an illustration of the pressure (force)-volume/travel dependency in the active mode (active) and in the fall-back level (RFE);

FIG. 1c: shows a further development of the brake system illustrated in FIG. 1a for the purposes of substantially optimizing the hydraulic cabling complexity and the system weight;

FIG. 1d: shows an embodiment with travel simulator unit;

FIG. 1e: shows a further development of the brake system illustrated in FIG. 1a, wherein the replenishment line has been omitted and a replenishment valve is arranged in place of the replenishment line;

FIG. 1f: shows a replenishment valve as per FIG. 1e;

FIG. 2a: schematically shows another embodiment of a brake system for a vehicle with high-power traction motor, wherein the first piston-cylinder unit acts exclusively only on the brakes of one brake circuit or of one vehicle axle;

FIG. 2b: is an illustration of the pressure (force)-volume/travel dependency in the active mode (active) and in the fall-back level (RFE) of the brake system as per FIG. 2a;

FIG. 2c: is an illustration of the first and second piston-cylinder units of the embodiment as per FIG. 2a with the associated hydraulic and electrical connecting lines;

FIG. 3a: shows a characteristic map of deceleration a_vehicle and braking torque M_braking=f(pressure) in the case of fading as a result of intense warming of the wheel brake;

FIG. 3b: shows a characteristic map of pressure-volume characteristic curves=f(piston travel) in the case of air in the system;

FIG. 4a: is an illustration of a first operating strategy of the brake system according to the invention, with maximized deceleration;

FIG. 4b: is an illustration of a second operating strategy of the brake system according to the invention, with controlled deceleration;

FIG. 5: is an illustration of the braking management with distribution of the braking torques by means of ECU between the traction motor and the second piston-cylinder unit or the EHB module;

FIG. 6: shows a structural design of the EHB module, with a second piston-cylinder unit, electric motor, spindle drive, valves, sensors and ECU.

DETAILED DESCRIPTION

FIG. 1a schematically shows the rear axle HA and the front axle VA of a vehicle with wheels and wheel brakes RB1, RB2 (HA) and RB3, RB4 (VA) and with an electric vehicle traction motor TM with high power (>100 kW) and high torque at one axle, preferably, as illustrated, at the rear axle.

The brake system illustrated in FIG. 1a, which is however specifically designed not exclusively for Formula E applications, has a first piston-cylinder unit 2, which in this case has the function of a (brake) master cylinder and which is actuatable by means of an actuating device 4, in particular a brake pedal. The piston-cylinder unit 2 has two pistons 6, 8 and pressure chambers 6a, 8a assigned to these.

Here, the piston 8 is, as illustrated, expediently but not imperatively designed as a stepped piston. In this way, a different brake pressure distribution at the axles can be achieved through corresponding configuration of the stepped piston. The pressure chambers 6a, 8a are connected via hydraulic connecting lines 10a and 12a to a reservoir (VB) 14, and via hydraulic connecting lines 16, 18 to wheel brakes RB1 and RB2, and RB3 and RB4, respectively. The hydraulic connecting lines 16, 18 form brake circuits BK1 and BK2. An in particular normally-open isolating valve (TV) 19 is arranged in the hydraulic line 16 that leads from the pressure chamber 6a to the wheel brakes RB1, RB2 of the driven axle (in this case HA). No valve is arranged in the hydraulic line 18 that leads from the pressure chamber 8a to the wheel brakes of the non-driven axle in this case VA. In other words, in particular, in each case one working chamber of the first piston-cylinder unit is connected to the wheel brakes of a vehicle axle.

At the first piston-cylinder unit 2 and the hydraulic connecting lines, there are provided various sensors, in particular, as illustrated in the drawing, two pressure transducers at the line 16 upstream of the isolating valve 19 and a further one downstream of the isolating valve and downstream of an isolating valve (DMV) which is described further below and which is assigned to the second piston-cylinder unit, and a pressure transducer at the line 18 or BKI. The first piston-cylinder unit 2 forms a first structural or assembly unit BE together with the actuating device 4 and various sensors, in particular pressure transducers 5, 5a, 5b and a travel sensor 7.

A second piston-cylinder unit 20 is a constituent part of an electromotively driven system or electric plunger. The second piston-cylinder unit 20 has a (plunger) piston 22 which delimits a pressure chamber 24 which is connected via a hydraulic line 26 and an in particular normally-closed valve (DMV) 28 to one of the brake circuits BK1. Here, the connecting line 26 of the second piston-cylinder unit 20 opens into the brake circuit line downstream of the isolating valve (TV) 19 as viewed in a direction from the first piston-cylinder unit 2. A further hydraulic connecting line is connected to the reservoir (VB) 14, such that, in the retracted position of the piston 22, pressure medium can pass out of the reservoir 14 into the pressure chamber 24.

The electric plunger has an electromotive drive, with a highly dynamic electric motor 30 and a gearing 32, in particular spindle gearing.

Sensors, in particular pressure transducers, angle encoders, rotational speed encoders, are, as illustrated, assigned to the electric plunger and/or integrated into the corresponding separate unit.

The electric plunger with the above-described components (second piston-cylinder unit with drive and gearing, isolating valves, sensors) forms a separate second structural or assembly unit or pressure feedback control unit (bordered by a dashed line) for the pressure feedback control of the brake system.

FIG. 1b is an illustration of the pressure (force)-volume (travel) dependency in the active mode (active) and in the fall-back level (RFE). The first curve shows the profile in the case of an intact booster (active), a shallower second curve shows the profile in the case of an active booster (EHB) with functional impairment for example owing to an air inclusion, and the third, even shallower curve (RFE) shows the profile in the so-called fall-back level in the event of failure of the booster. Upon actuation of the actuating device 4 or of the brake pedal, pressure 2 is built up in the two pressure chambers 6a, 8a of the first piston-cylinder unit, which pressure is firstly transmitted via the line 16 and the isolating valve 19 to the wheel brakes RB1, RB2 of the driven rear axle and via the line 18 (without isolating valve) to the wheel brakes RB3, RB4 of the non-driven front axle. Independently of this, by means of the second piston-cylinder unit, by actuation of the electric motor controlled by an electronic control and feedback control unit ECU, a booster pressure can be dynamically built up and dissipated or modulated. Said booster pressure is transmitted via the line 26 and the isolating valve (DMV) 28 likewise to the brake circuit BKI. This may be performed independently of the pressure build-up by means of the first piston-cylinder unit 2 or in parallel therewith. The feedback control of the pressure in the brake circuit BKI is performed by means of the travel control of the piston 22 of the second piston-cylinder unit 20 by means of the electric motor 30, utilizing the sensors (pressure transducers, angle encoders) and the pressure-volume characteristic curve.

Here, pressure feedback control is performed by means of the EHB unit and the torque feedback control of the traction motor TM of the vehicle, which together determine the deceleration of the vehicle at the axle 1. The deceleration at the axle 2 is determined exclusively by the actuation force and the pressure, wherein, in the two-circuit embodiment, the pressure of the working chamber is transmitted via a floating piston, and the cross-sectional area of the second working chamber determines the pressure in the wheel brakes of the axle 2.

Here, the braking feedback control at the axle 1 does not have an effect on the pedal feel. This is determined exclusively by the hydraulic connection of the first piston-cylinder unit 2 to the wheel brakes of the axle 2.

FIG. 1c shows a weight-optimized alternative to the schematic illustration of the brake system illustrated and described in FIG. 1a. Here, the connecting line between reservoir 14 and pressure chamber 24 is omitted. Furthermore, the isolating valve (DMV) 28 is omitted entirely. The main motivation for this is the weight saving of the line from the reservoir 14 to the piston-cylinder unit 20. Since, in this way, a volume compensation in the pressure chamber 24 in the electrically deenergized state is no longer ensured, the isolating valve (DMV) 28 must also be omitted. A volume compensation caused for example by a temperature variation or by knock-back at the wheel brakes RB1 or RB2 can thus be realized via the hydraulic line 26, the valve 19, the hydraulic line 16 and the pressure chamber 6a to the reservoir.

For the switch to the fall-back level, the omission of the isolating valve (DMV) entails the following:

• • I. Failure in the case of approximately equal pressure levels to the right and to the left of the isolating valve (TV):
    • By contrast to the system with isolating valve (DMV), the driver in this case experiences a finite pedal drop. A loss of volume in relation to the mechanical fall-back level does not occur.
  • II. Failure in the presence of high recuperation power, such that the pressure in the hydraulic line 16 is considerably higher than the pressure in the hydraulic line 26:
    • As in the system with isolating valve (DMV), the driver in this case experiences a finite pedal drop. A loss of volume in relation to the mechanical fall-back level does not occur here either.
  • III. Failure in the presence of high pressure boosting, such that the pressure in the hydraulic line 16 is considerably lower than the pressure in the hydraulic line 26:
    • By contrast to the system with isolating valve (DMV), the driver in this case, too, experiences a finite but greater pedal drop. A loss of volume in relation to the mechanical fall-back level does not occur.
    • Summary: The omission of the DMV does not result in any severe functional disadvantages.
    • The movable seals for the pressure chamber 24 can now be designed differently owing to the omission of the connecting line to the reservoir. The fade of the hitherto provided high-pressure and low-pressure seals fixed to the housing with closed snifter bore is possible as a first solution (as depicted in FIG. 1c). An alternative solution would be a single seal fixed to the housing in a wide variety of different technically known embodiments. In addition to this, there is also the possibility of a seal fixed to the piston, which has advantages in terms of the wear of the seal.

An embodiment as per FIG. 1c furthermore offers functional advantages. Firstly, owing to the modified construction without reservoir port, the seal closing travel is reduced. This increases the pressure feedback control dynamics in the low pressure range up to 10 bar. A second functional degree of freedom is described in DE 10 2008 051 316 A1. Through the use of a seal that can withstand negative pressure for the pressure chamber 24, it is now also possible to realize a so-called active retraction of the brake pads. This can, depending on the design of the wheel brakes, advantageously influence the energy consumption of the vehicle.

FIG. 1d shows a further embodiment of FIG. 1a with additional travel simulator unit 101. This may be mechanically integrated into the electrohydraulic brake EHB or installed as a separate module. The travel simulator unit 101 has the effect that the pedal travel to pedal force characteristic curve can be configured optimally during active operation. For example, it may be the case that, owing to the stiff hydraulic regions 6a and 16, the pedal travel is too short in relation to the respective pedal force. Then, the installation of a travel simulator unit 101 is necessary. Said travel simulator unit is composed of substantially 4 components, the piston travel simulator 102 for replicating the travel-pressure characteristic, the simulator shut-off valve (SiV) 103 for shutting off the piston travel simulator in the fall-back level, the throttle 104 for realizing optimum pedal damping, the check valve 105 for the purposes of bypassing the throttle and the SiV, and having no pedal damping, when the pedal is released. In motor sport, the throttle and check valve are preferably omitted for weight reasons. In certain configurations, it is also possible for the simulator shut-off valve to be omitted. This is possible only if the volume capacity of the piston travel simulator is very small in relation to the total volume capacity of the wheel brakes, and the design permits this.

FIG. 1e shows a further possible refinement of the brake system described in FIG. 1a, wherein the hydraulic pressure compensation between brake system EHB and the reservoir is such that no replenishment line is necessary. The overall weight is thus reduced significantly, which is a considerable advantage specifically in motor sport.

The brake system EHB must, in the non-actuated state, be connected directly or indirectly to the reservoir in order that no positive pressure or negative pressure can form in the pressure chamber 24, and pad wear in the wheel brakes RB1 and RB2 can be compensated.

By contrast to the construction described in FIG. 1a, it is the case here that the direct hydraulic connection between the brake system EHB and the reservoir VB is severed and replaced by an indirect connection via the replenishment valve 110, the line 16, the chamber 6a and the connecting line 10a.

The connecting valve 110 is designed so as to be open in the non-actuated state and so as to automatically close if pressure is built up in the pressure chamber 6a. Thus, in the non-actuated state, there is an indirect connection between the pressure chamber 24 and reservoir VB.

FIG. 1f illustrates the specific construction of the replenishment valve 110 and the 2 possible switching positions.

The valve plunger 111 is guided in the valve housing 112 and, in the non-actuated state, is pushed by a valve spring 113 into the position in which the two valve ports 114 and 115 are connected to one another.

The valve plunger additionally has a bore with an orifice 116. As soon as pressure is built up in the pressure chamber 6a and thus also in the connecting line 16, a back pressure forms in the orifice 116, which back pressure pushes the valve plunger 111, counter to the spring 113, into the valve seat 118. The valve ports 114 and 115 are thus hydraulically separated.

FIG. 2a schematically shows another embodiment of a brake system for a vehicle with high-power traction motor, wherein the first piston-cylinder unit acts exclusively only on the brakes of one brake circuit or of one vehicle axle.

In this embodiment, the first piston-cylinder unit 3 has a pressure chamber 5 which is connected via a hydraulic connecting line 7 (BKII) to the wheel brakes RB3, RB4 of an axle 2. Two pressure transducers are arranged at the line 18a. The movement of the single piston 5a can be sensed by means of a pedal travel encoder 7 (not illustrated in any more detail).

The second piston-cylinder unit 20 is connected via a hydraulic connecting line 26 (BKI) to the wheel brakes RB1, RB2 of an axle 1 which is driven by the traction motor TM. A pressure transducer is provided at the connecting line 26.

FIG. 2c is an illustration of the first and second piston-cylinder units of the embodiment as per FIG. 2a with the associated hydraulic and electrical connecting lines.

The brake circuits BKI for the driven axle and BKII for the non-driven axle are, in this embodiment, completely separated from one another with regard to the actuating devices, that is to say first piston-cylinder unit and second piston-cylinder unit (EHB). The pedal feel is, as in FIG. 1a, determined by the pressure-volume characteristic of the wheel brakes of the axle 2 and the cross-sectional area of the first piston-cylinder system, which is hydraulically connected to the first piston-cylinder unit.

FIG. 2b is an illustration of the pressure (force)-volume (travel) dependency in the active mode (active) and in the fall-back level (RFE). The first curve shows the profile in the case of an intact booster (active), and a shallower second curve shows the profile in the case of an active booster with functional impairment for example owing to an air inclusion.

FIG. 2c shows, in another illustration, the structural or assembly units of a brake system as per FIG. 2a. Here, a first structural or assembly unit has the first piston-cylinder unit 3 with the actuating device (not illustrated here). The reservoir 14 is fastened to this. A hydraulic connecting line leads from the first structural or assembly unit to the non-driven axle 2 of the vehicle. An electrical connection leads from the first unit to the ECU of the second unit (EHB).

The second unit comprises the second piston-cylinder unit 20 with the valves and the electronic control and feedback control unit (ECU). Hydraulic connections lead from the second structural or assembly unit to that axle 1 of the vehicle which is driven by means of traction motor TM, and to the reservoir 14.

FIG. 3a shows illustrations of the relationship of the vehicle deceleration a_vehicle or braking torque M_braking on the pressure p, which may vary in the hydraulic brake as a result of warming or wear of the brake system. For example, in the event of fading, the braking action deteriorates for example owing to a high temperature in the brake system, that is to say the pressure must be increased in order to realize a desired constant braking action.

This results in a steeper curve M_braking=f(p) without fading or a shallower curve (shown by dashed lines) with fading. For targeted feedback control of the deceleration, it is therefore important to detect and evaluate the relationship between braking torque and pressure and store said relationship in a characteristic map in the memory of the ECU.

FIG. 3b is an illustration of the pressure p as a function of the piston travel x_piston (pressure-volume characteristic curve or pressure-travel characteristic curve), wherein the relatively steep curve shows the pressure without air in the system and the relatively shallow curve shows the pressure with air in the system. For exact feedback control of the pressure as a function of the piston travel, it is therefore important to detect variations in the brake system and adjust the feedback control to these. It is therefore important to adaptively perform trimming of the pressure-volume characteristic curve and/or to utilize a characteristic map, which is evaluated. For this purpose, it may suffice to only evaluate pressures in particular positions (for example phase with constant pressure) in order, from this, to evaluate the relevant pressure-volume characteristic curve of the characteristic map.

FIG. 4a illustrates a first operating strategy of a brake system according to the invention. Here, a maximum deceleration is sought. Highly dynamic pressure generation up to the maximum deceleration a_max is performed here by means of the traction motor TM and the EHB or second piston-cylinder unit 20. Here, a_total is determined from the sum of the values a_TM and a_EHB. The value of a_max is in this case variable and can take into consideration a fading situation, for example. With regard to further details, reference is made directly to FIG. 5.

FIG. 4b shows a second operating strategy, wherein effective recuperation by means of the traction motor TM is sought. The feedback control Δp/dt and the controlled deceleration with a_setpoint are influential variables which are particularly important for motor sport. In the initial range, deceleration is performed with maximum recuperation by means of the traction motor a_TM. In the middle range, a steep increase of a_total (Δp/dt) occurs, and a controlled deceleration with a_setpoint is performed in the subsequent range.

FIG. 5 shows, in principle, the braking management of a brake system with high-power electric motor TM and electrohydraulic brake system EHB. There, the setpoint deceleration a_setpoint is divided between traction motor TM and hydraulic brake EHB in accordance with the objective of the braking process (maximum recuperation, maximum deceleration, controlled deceleration). In this context, restrictions such as maximum torque of the traction motor are taken into consideration in a manner dependent on the vehicle speed or motor rotational speed, and a vehicle model is used which replicates weight distribution, friction coefficient of the roadway and tires and which predefines further restrictions.

A setpoint deceleration a_setpoint,TM and advantageously also the profile of the deceleration Da/dt are transmitted to the ECU of the traction motor. In the ECU, taking into consideration the efficiency of the motor and gearing, a setpoint torque M_setpoint is transmitted to the motor controller.

At the same time, a setpoint torque p_setpoint,EHB, a pressure gradient Dp_EHB/dt and the temperature T are transmitted to the ECU of the EHB. From these variables, and after evaluation of the characteristic maps p=f(x_piston) and a=f(p) recorded in the ECU, setpoint torque M_setpoint,EHB, setpoint rotational speed n_setpoint,EHB and setpoint position x_setpoint,EHB are transmitted to the motor controller and partially or entirely used in the feedback control, wherein the focus is on the position feedback control of the piston, and the characteristic maps are used inter alia in order to adjust the setpoint pressure to variations of the wheel brake, for example variations in the braking action in the presence of predefined pressures in the event of fading. Pressure transducers in the system are then used in the outermost feedback control loop only for readjustment, owing to the inertia of the measurement. The aim is that of achieving the most precise possible pilot control or, with corresponding model accuracy, omitting the pressure transducer as a feedback control variable. The pressure transducer is therefore, aside from the readjustment, used primarily for the characteristic map recording or parameterization and variations of parameters in the brake system. The very small time constants of an electric motor and the high accuracy of the current measurement and angle encoders in relation to pressure transducers are used for highly dynamic feedback control. Viscosities in the hydraulic system furthermore lead to delay times in the pressure measurement, which are furthermore not constant in the event of temperature changes.

FIG. 6 illustrates a structural embodiment of the EHB module, in the case of which the components or structural units of the EHB module are constructed and combined in a particularly advantageous compact design. As already described, the EHB module forms a separate structural unit. This has substantially the electric motor 30, the spindle gearing 32, the second piston-cylinder unit 20 and the associated sensors.

Here, the electric motor has a motor housing 40 in which an outer stator 42 is mounted. A rotor 46 is arranged in the stator by means of an in particular unilateral bearing arrangement with only one bearing 44. For the unilateral bearing arrangement, a 4-point bearing is advantageously used which, in an axial direction, is seated substantially in the region of the radially extending housing wall, in particular, as illustrated, in an axial projection 47 of the motor housing. The inner ring 49 of the bearing is seated on the outer circumference of the rotor 46. Further details regarding the unilateral bearing arrangement emerge from PCT/EP2013/057609, to which reference is made here in this respect. A nut 48 is arranged in the front part of the rotor 46 in the interior of the rotor. Said nut is a constituent part of a ball screw drive, which includes a spindle 50 arranged in the nut 48, which spindle is equipped with a rotation prevention means 52 such that a rotation of the nut 48 results in an axial displacement of the spindle 50. Seated centrally in the spindle 50 is a plunger 54 which extends out of the spindle 50 in the direction of the plunger piston. By way of its front end, the plunger 54 is connected by means of a connecting device to the piston 22, such that, in the event of movement of the spindle in both directions, said piston is driven along by the spindle 50.

The piston 22 of the second piston-cylinder unit 20 is arranged in a corresponding bore of a housing 56. Said housing 56 also entirely or partially receives, in a recess 58, the isolating valves and the corresponding hydraulic connecting lines. The longitudinal axes of the isolating valves 19, 28 in this case run substantially perpendicular to the longitudinal axis of the second piston-cylinder unit 20. In the upper part of the housing 56, that is to say the part averted from the second piston-cylinder unit, said housing has a lateral extension 60 in order to create sufficient space for the arrangement of an electronic control and feedback control unit (ECU) 64. The contacting of the valve coils with the ECU 64 is realized by means of corresponding devices 68 in the region of the circuit board 66. The motor housing 40 is attached, in particular by screw connection, to the housing 56 laterally, or below the housing extension 60.

It will be appreciated by those skilled in the art that changes could be made to the embodiments described above without departing from the broad inventive concept thereof. It is understood, therefore, that this invention is not limited to the particular embodiments disclosed, but it is intended to cover modifications within the spirit and scope of the present invention as defined by the appended claims.

REFERENCE DESIGNATIONS

2 First piston-cylinder unit (FIG. 1a)

4 Actuating apparatus, in particular brake pedal

3 First piston-cylinder unit (FIG. 2a)

5 Pressure chamber

5a Piston

6 Piston

6a Pressure chamber

7 Travel sensor

8 Piston

8a Pressure chamber

10a Connecting line

12a Connecting line

14 Reservoir (VB)

16a Connecting line

18a Connecting line

19 Isolating valve (TR)

20 Second piston-cylinder unit

22 (Plunger) piston

24 Pressure chamber

26 Connecting line

28 Isolating valve (DMV)

30 Electric motor

32 Gearing

40 Motor housing

42 Stator

44 Bearing

46 Rotor

47 Axial projection

48 Nut

49 Inner ring

50 Spindle

52 Rotation prevention means

54 Plunger

56 Housing

58 Recess

60 Housing extension

64 Control and feedback control unit (ECU)

66 Circuit board

68 Contact device

70 Motor housing

72 Outer stator

74 Rotor

76 Piston

78 Piston-cylinder unit

80 Recess

82 Recess

84 Flexural rod

86 Spindle

88 Nut

92 Bore

94 Housing cover

96 Rotation prevention means

100 Electronic control and feedback control unit (ECU)

101 Travel simulator unit

102 Piston travel simulator

103 Simulator shut-off valve (SiV)

104 Throttle

105 Check valve

110 Replenishment valve

111 Valve plunger

112 Valve housing

113 Valve spring

114 Valve port

115 Valve port

116 Bore with orifice

117 Replenishment port

h Efficiency

p Pressure

x Travel

M Torque

a Deceleration

n Rotational speed

T Temperature

V_vehicle Vehicle speed

Claims

1. A brake apparatus for electrically driven motor vehicles having a first and a second axle, each having two wheel brakes, the brake apparatus including:

a. a traction motor at the first axle, wherein the traction motor is used both as a drive motor and as a brake system with recuperation of brake energy,

b. a first piston-cylinder unit, which is actuatable by means of an actuating device, the first piston-cylinder unit either having a first piston separating a first pressure chamber, or having a first piston separating a first pressure chamber and a second piston separating a second pressure chamber, wherein the first pressure chamber, or at least one of the first pressure chamber or the second pressure chamber, of the first piston-cylinder unit is connected to the two wheel brakes of one of the first axle or the second axle via a first hydraulic connecting line,

c. a second piston-cylinder unit, which is actuatable by means of an electromotive drive and a non-hydraulic gearing apparatus, wherein the second piston-cylinder unit has exactly one piston and one working chamber, and

d. at least one electronic control and feedback control unit,

wherein the working chamber of the second piston-cylinder unit is connected to the wheel brakes of the first vehicle axle, via a second hydraulic connecting line, enabling active brake force feedback control and recuperation control in interaction with the traction motor, and

wherein the at least one electronic control and feedback control unit is adapted to perform a distribution of braking torques at the wheel brakes or the corresponding axles between the traction motor and the second piston-cylinder unit, wherein a vehicle model is taken into consideration for performing the distribution, the vehicle model replicating a weight distribution and a friction coefficient between a roadway and tires of the vehicle.

2. The brake apparatus as claimed in claim 1, wherein the brake apparatus is adapted to perform pressure metering in accordance with demand by means of the second piston-cylinder unit, wherein the control is performed via travel control or combined travel and pressure control of the piston of the second piston-cylinder unit through utilization of a pressure-volume characteristic curve.

3. The brake apparatus as claimed in claim 2, wherein the pressure-volume characteristic curve is configured to be adaptively adjusted after every braking operation.

4. The brake apparatus as claimed in claim 2, wherein the brake apparatus is adapted to utilize a characteristic map for refined brake force feedback control, the characteristic map providing a relationship between brake pressure and deceleration depending on temperatures of the wheel brakes.

5. The brake apparatus as claimed in claim 4, wherein the brake apparatus is adapted to perform pilot control using an isolating valve and feedback control of the pressure change, to enable attainment of a target pressure or to enable adjustment to the traction motor braking feedback control.

6. The brake apparatus as claimed in claim 1, wherein the brake apparatus is adapted to maximize the recuperation of the vehicle by means of intermittent braking only by means of the traction motor.

7. The brake apparatus as claimed in claim 2, wherein the brake apparatus is adapted to replicate a pedal feel by means of the pressure-volume characteristic of a wheel brake or multiple wheel brakes of one of the first or second axles.

8. A method for braking a motor vehicle having an electric traction motor and having an electrohydraulic brake system with a piston-cylinder unit, the method including:

controlling, in a first operating strategy, an electric motor of the piston-cylinder unit and the electric traction motor to maximize a total deceleration of the motor vehicle, wherein the total deceleration is a sum of a deceleration established by the traction motor and a deceleration established using the electrohydraulic brake system, and

controlling, in the first operating strategy, the electric motor of the piston-cylinder unit and the traction motor to minimize a time-to-lock,

wherein a vehicle model is taken into consideration to distribute the total deceleration between the deceleration established by the traction motor and the deceleration established using the electrohydraulic brake system, the vehicle model replicating a weight distribution and a friction coefficient between a roadway and tires of the vehicle.

9. The method as claimed in claim 8, further including controlling, in the first operating strategy, the traction motor such that up to at least the total deceleration of 1 g the deceleration established by the traction motor is maximized.

10. The method as claimed in claim 8, further controlling, in the first operating strategy, the traction motor in a first phase such that the deceleration established by the traction motor is maximized to approach a maximum total deceleration.

11. The method as claimed in claim 10, further include controlling, in the first operating strategy, the total deceleration in a second phase after the first phase such that the total deceleration is increased.

12. The method as claimed in claim 8, further including controlling, in the first operating strategy, the traction motor such that the deceleration established by the traction motor is increased in at least a first interval of 50 ms and maintained as the total deceleration approaches a maximum total deceleration and/or at least until at least a total deceleration of 1 g is established.

13. The method as claimed in claim 8, further including decelerating the vehicle, in a second operating strategy, with the total deceleration lower than a setpoint deceleration, whereby the deceleration established by the traction motor is kept constant.

14. The method as claimed in claim 8, wherein for controlling the deceleration established by the traction motor a maximum torque of the traction motor is taking into consideration, wherein the maximum torque depends on vehicle speed or rotational speed of the traction motor.

15. The method as claimed in claim 8, wherein the deceleration established using the electrohydraulic brake system is controlled using a set pressure, wherein for determining the set pressure a temperature is taken into consideration to compensate a fading effect.

16. A method for managing a brake apparatus of a motor vehicle having an electric traction motor and having an electrohydraulic brake system with a piston-cylinder unit, the method comprising:

determining an objective of a braking operation,

dividing a setpoint deceleration between the traction motor and the electrohydraulic brake system in accordance with the determined objective of the braking operation, and transmitting setpoint values to a first control and feedback control unit, which performs control and feedback control of the traction motor, and to a second control and feedback control unit, which performs control and feedback control of the electrohydraulic brake system, wherein at least one of a traction motor deceleration setpoint value and/or a traction motor torque setpoint value is determined for the traction motor, wherein at least one of an electrohydraulic brake system deceleration setpoint value, an electrohydraulic brake system pressure setpoint value, or an electrohydraulic brake system torque setpoint value is transmitted to the electrohydraulic brake system simultaneously with the setpoint value(s) transmitted to the traction motor, and

braking using the traction motor and the electrohydraulic brake system in accordance with the divided setpoint deceleration.

17. The method as claimed in claim 16, further comprising:

determining the setpoint deceleration at the traction motor and the electrohydraulic brake system in accordance with an objective of maximum recuperation, maximum deceleration or controlled deceleration, and

maximizing recuperation of the motor vehicle by means of intermittent braking only by means of the traction motor.

18. The method as claimed in claim 16, further comprising:

maximizing pressure build-up dynamics in a pressure range up to 10 bar by assistance of a further electric motor of the piston-cylinder unit.

19. The method as claimed in claim 16, further comprising:

evaluating a pressure-volume characteristic curve of a wheel brake and a pressure at an associated axle or a pressure of the electrohydraulic brake system for travel control of a piston of the electrohydraulic brake system.

20. The method as claimed in claim 16, further comprising:

taking a temperature into consideration for detecting a change in the electrohydraulic brake system.