The present invention relates to a method for controlling a brake system and a correspondingly designed control system.
State of the art:
Modern brake systems contain brake boosters, i.e. the pedal force is converted into a correspondingly amplified braking torque at the wheel brakes and braking force control is achieved via open or closed control and regulating circuits. Apart from a few exceptions in passenger cars, the hydraulic line is used as the transmission medium for generating the brake pressure from the pedal force.
A division into units between brake force amplification (BKV) or brake force control and brake force regulation in a hydraulic unit (HE) is widely used. This configuration is mainly used for systems such as Antilock Braking System (ABS), Anti-Slip System (ASR), Electronic Stability Program (ESP) or even Electrohydraulic Braking (EHB).
The hydraulic unit (HE) usually consists of solenoid valves, multi-piston pumps for 2-circuit braking systems (front and rear axle), electric motor to drive the pump, hydraulic accumulator and several pressure sensors. Pressure control is achieved by draining pressure fluid from the wheel brakes into an accumulator via solenoid valves to reduce braking torque. The pressure fluid is then pumped back into the master cylinder by the pump, causing a pedal movement. A corresponding system is known from US 4057301 A.
Both pressure rise and fall is controlled by solenoid valves, with some pressure transmitters being used for solenoid valve control.
Except for the EHB, the brake force is amplified with the vacuum BKV, which partly contains switching means and sensors for the so-called brake assistant function and also for the recognition of the so-called control point. In gasoline engines, the combustion engine is used as the energy source for the vacuum, but as a direct injector, especially at higher altitudes, it only provides a weak vacuum. For diesel engines, a mechanically or electrically driven vacuum pump is used. The latest ESP systems are able to achieve additional brake force boosting by switching the solenoid valves and pump or, in the event of failure of the BKV, brake force boosting with a larger time constant. The description of these systems and functions is described in detail in the Vieweg Verlag brake manual, 2003 edition.
In the mid-1980s, Teves used the so-called Mark II and Bosch the ABS3, which as integrated units contained all components for brake booster and control with hydraulic BKV, see Bosch Motor Vehicle Handbook 1986, 20th edition. For cost reasons, these systems did not become generally accepted, except for the application in special protection vehicles. The same applies to fully electric braking systems, so-called EMB, with electric motors on the wheel brakes, which were intensively developed in conjunction with the 42 V vehicle electrical system. In addition to the additional costs, a new redundant on-board power supply system is necessary here to ensure the braking capability of a brake circuit in the event of a fault.
EMB systems also include the wedge brake with electric motor drive. This also requires a redundant on-board power supply system despite the lower energy requirement. The constructive realization of the wedge brake, which requires additional rollers for hysteresis reasons, which require integration into the brake caliper, is not solved at the moment. The wedge brake with its electromotive drives with sensors must withstand the harsh environmental conditions (dust, water, high temperatures).
The systems for BKV and HE are very advanced, especially the control and regulation functions for ABS to ESP. For example, the pressure-guided control of the solenoid valves allows very fine dosage of the brake pressure, which also allows variable brake force adjustment EBV. The pressure reduction rate is not yet optimal because it is highly non-linear. In addition, in the event of a m-jump or with a low coefficient of friction, the pressure reduction rate is determined by the relatively low pumping power, which leads to large control deviations and thus to a loss of braking distance.
A brake system is known from DE 3342552. In this brake system, the master brake cylinder is used to generate a pedal-dependent pressure which serves as a command variable for an electronic control and regulating device which regulates the output pressure of an electrohydraulic servo device directly connected to the brake circuit to a value determined by the command variable. If the control device or the servo device itself fails, the pressure in the brake circuit is generated by the master cylinder. Instead of the reference variable generated by the master brake cylinder during normal operation, it is possible to have a reference variable generated as part of an anti-lock braking system or as part of a slip control of the drive control of the motor vehicle act on the electronic control and regulating device and thus on the electrohydraulic servo device. The servo device has an electrically operated hydraulic piston-cylinder unit, whose working chamber is connected to the brake circuit and whose piston is axially adjustable by means of an electric motor. The rotary motion of the electric motor is converted into a longitudinal motion of the piston via a spindle connected to the piston.
From WO2004/005095 A1, a braking system is well-known in which an electric motor drives the pistons of a piston cylinder system via a spindle drive. The pistons are not rigidly coupled to the spindle, so that the maximum piston speed when the spindle is retracted and thus the maximum pressure reduction speed is determined by the strength of the compression springs in the piston cylinder system. The brake pressure to be set in the wheel brakes is determined by means of a pressure sensor, whereby the pressure is the controlled variable of the brake pressure control.
The DE 3723916 A1 and DE 3440972 A1 show systems with a hydraulic brake booster, which realizes the ABS function in addition to the pure brake booster. In the pressure line connecting the piston cylinder system and the respective wheel brake there is only one valve each, which is open to change the pressure in the wheel brake and closed to maintain the wheel brake pressure. The pressure is also the controlled variable for this brake pressure control.
From DE 19500544 A1, an electronically controllable brake actuation system for anti-lock automotive brake systems is already known, in which a master brake cylinder can be actuated by a brake pedal. By means of a sensor, the actuation travel of the brake pedal is determined, which represents an input variable for a control unit which controls several brake pressure generators to which the vehicle brakes are connected directly or via solenoid valves by means of hydraulic lines. The connection of the hydraulic lines to the master brake cylinder can be shut off by a valve device. In order to increase functional safety, especially in the event of an electrical defect or failure of the vehicle electronics, the piston of the master brake cylinder can be adjusted in the release plane directly by means of the brake pedal to build up pressure in the wheel brakes, whereby the valve device is opened for this purpose. The brake pressure generators each have an electric drive which adjusts a piston in a cylinder so that a pressure is set in the brake circuit which is determined by means of a pressure sensor and fed to the control unit as an input variable. The pressure is also the controlled variable for this brake pressure control. A similarly working brake system is known from DE 4239386 A1, DE 10304145 A1 and DE 19543583 C1.
A brake system for motor vehicles is known from DE 4445975 A1, in which the brake pressure in a wheel brake is controlled by means of an electrically driven piston of a piston-cylinder system, whereby a pressure sensor is also provided in this brake system to measure the controlled variable. A 2/2-way valve is used to maintain the brake pressure in the wheel brake, by means of which the hydraulic line between the piston-cylinder system and the wheel brake can be shut off.
DE 10318401 A1 reveals a motor-driven vehicle braking device in which the position of the brake pedal is determined by means of a displacement sensor and transmitted to a control unit. Depending on the driving condition and the position of the brake pedal, the control unit controls an electromotive drive of a piston-cylinder system, which serves to build up pressure in the brake circuits. A mechanical connection between the piston of the piston-cylinder system and the brake pedal is not provided, so that in the fallback plane no pressure can be built up in the wheel brakes by means of the brake pedal. The pressure in the wheel brakes is regulated by means of inlet and outlet valves assigned to the respective wheel brakes. A similar brake system is known from DE 3241662 A1 and DE 19750977 A1.
The US 6079792 shows a conventional hydraulic brake system in which the piston-cylinder unit operated by an actuator can release the wheel brakes by mechanical action along the piston-cylinder axis.
DE 19936433 A1 and DE 10057557 A1 reveal brake systems in which a supporting force can be applied to the pistons of the master brake cylinder, which are adjustable by the brake pedal, by means of electromagnetic drives. In these brake systems, too, the pressure in the master brake cylinder is the controlled variable of the brake pressure control process.
From DE 69515272 T2, a brake control method for controlling an actuator for electrohydraulic brake pressure modulation is known, in which the brake pressure is set exclusively by means of the actuator via a spindle and piston-cylinder unit and effects from the controlled system can be compensated.
Furthermore, EP 0840441 B1 provides a field-oriented control system that allows a rotating field machine operating in both drive and generator mode to be controlled in such a way that unfavorable overcurrents are avoided. It describes the possibilities of a controller-based solution for controlling a given generator (braking) torque. This makes it possible to set a desired braking torque for a polyphase machine by means of the voltage components, knowing the stator currents, the motor speed, the rotor speed and the flux-dependent inductors.
Further reference is made to DE 19939950 A1 and DE 4239386 A1.
Abandonment of the invention
The purpose of the present invention is to provide an improved method for controlling a brake system.
This task is advantageously solved by a procedure with the characteristics of claim 1. Further advantageous designs of the brake system according to claim 1 result from the characteristics of the subclaims.
The brake system according to the invention is advantageously characterized by the fact that a variable pedal characteristic is controlled, an assignment between pressure and pedal force is freely variable, in the case of recuperable braking by means of a generator, a correction value is used to take into account the braking effect of the generator, and the drive device of the brake system is a brushless electric motor in which three lines are controlled by means of a microcontroller, taking into account a position measurement of the rotor of the electric motor and a current measurement. The piston-cylinder unit can also be used to build up and reduce brake pressure, to implement ABS and anti-slip control, and in the event of a power failure or malfunction of the drive device. This results in a small, integrated and cost-effective unit for brake booster and control, which saves installation space, assembly costs and additional hydraulic and vacuum connection lines. In addition, due to the short overall length, the spring dome does not affect the master cylinder and the pedal mechanism in the event of a frontal crash.
By the advantageous provision of a sensor system as well as a travel simulator, a variable pedal characteristic such as brake-by-wire function, i.e. brake pressure increase independent of pedal actuation, can be controlled freely variably, also taking into account the braking effect of the generator with recuperable brakes.
It is known from specialist literature, e.g. Schröder D., Elektrische Antriebe- Regelung von Antriebssystemen, 2nd edition, Springer-Verlag Berlin Heidelberg, 2001, that a generator braking torque can be estimated by knowing the fixed geometric variables (number of pole pairs), the motor-typical inductances (from the motor data sheet) and the measurable currents. A similar braking system that takes into account a regenerative braking torque is known from DE 19939950 A1.
Likewise, with the corresponding design, there is no disadvantageous falling through of the brake pedal in case of failure of the drive, since the pedal acts directly on the piston of the system. This also has the advantage of lower pedal forces in the event of a power failure, since the pistons have a smaller effective area than conventional master brake cylinders. This is possible by separating the piston travel when the reinforcement is intact and when it has failed. This is known as a transmission ratio jump, which reduces the pedal force by up to 40% for the same braking effect. The reduction of the total effort including the electrical connections also results in a reduction of the failure rate.
Furthermore, the electromotive drive allows an improvement of the ABS/ESP control by finely dosed pressure control with variable pressure rise and especially pressure drop speeds. A pressure reduction below 1 bar in the vacuum range is also possible for operation with the smallest friction coefficients, e.g. wet ice. Similarly, a rapid pressure increase at the start of braking, e.g. 0 - 100 bar, can be achieved in less than 50 ms, resulting in a considerable reduction in braking distance.
Due to the advantageous provision of a 2/2 directional control valve for the brake force amplification and the control function, the brake system according to the invention requires considerably less energy.
It is also possible to provide a separate piston-cylinder system for each brake circuit or wheel brake, each with its own drive. It is also possible to use a piston-cylinder system in which two pistons are arranged axially displaceably in a cylinder, whereby the cylinders are hydraulically coupled and only one piston is mechanically driven by the drive unit via electric motor.
In the following, various designs of the brake system according to the invention are explained in more detail using drawings.
- Fig. 1: A first version of a brake system with one brake circuit for two wheel brakes;
- Fig. 2: A second version of the brake system with two piston-cylinder systems for two brake circuits for two wheel brakes each;
- Fig. 3: A travel simulator for the invented braking system;
- Fig. 4: A piston-cylinder system with one cylinder and two pistons;
- Fig. 5 and 5a: Connection between actuator and piston-cylinder systems;
- Fig. 6: A side view of the integrated unit with housing;
- Fig. 7: Characteristic curves of the brake system;
- Fig. 8 and 8a: Piston drive via a crank rocker
- Fig. 9: Piston drive via a spindle
- Fig. 10: Piston actuation with superimposed pedal force
Fig. 1 shows a section of the integrated unit which is responsible for pressure generation or brake force amplification. Here, piston 1 with the usual seals 2 and 3 in the cylinder housing 4 is moved parallel to the piston via a specially designed gear rack 5a. Seal 2 is designed to seal even when there is negative pressure in piston chamber 4'. This gear rack 5a transmits the force to the front crowned end of piston 1, which has a collar pin 1a at this point, via which gear rack 5a with return spring 9 moves the piston to its initial position. Here the gear rack rests against the cylinder housing 4a. This external spring has the advantage that the cylinder is short and has little dead space, which is advantageous for venting. Due to the transverse forces, the rack is supported in rollers 10 and 11 with sliding piece 12. Figure 1 clearly shows that the parallel arrangement of the rack to the piston results in a short overall length. The assembly unit must be very short to be outside the crash zone. The gear rack is very rigid due to an H-profile shown in Fig. 5a. The arrangement of the rollers is selected so that the rack has a relatively short bending length in the end position 5b (shown as dashed lines) with the greatest bending force due to the offset pressure force. The gear rack is driven via tooth profile 5a' and gear 6 via gear 7 by the pinion of motor 8. This motor with a small time constant is preferably a brushless motor as a bell rotor with ironless winding or preferably a motor according to the PCT patent applications PCT/EP2005/002440 and PCT/EP2005/002441. It is controlled by the power amplifiers 21 preferably via three lines by a microcontroller (MC) 22. For this purpose a shunt 23 measures the current and a sensor signal 24 and indicates the position of the rotor and via corresponding counters the position of the piston. The current and position measurement is used in addition to motor control for indirect pressure measurement, since the motor torque is proportional to the pressure force. For this purpose, a characteristic map must be created in the vehicle during commissioning and also during operation, in which the position of the piston is assigned to the different flow rates. During operation, the piston is then moved to a position corresponding to the amplifier characteristic curve described later, which corresponds to a certain pressure according to the characteristic map. If position and motor torque do not quite match, e.g. due to temperature influence, the characteristic map is adapted during operation. Thus the map is continuously adapted. The output characteristic map is formed from preferably pressure-volume characteristic of the wheel brake, engine characteristic value, transmission efficiency and vehicle deceleration. With the latter, a vehicle deceleration proportional to pedal force can be achieved so that the driver does not have to adjust to different braking effects.
Piston 1 generates a corresponding pressure in line 13, which reaches the wheel brake 15 via the 2/2 solenoid valve (MV) 14 and the wheel brake 17 via solenoid valve MV 16. This described arrangement has several advantages. Instead of the two inexpensive small solenoid valves, another piston-motor unit could be used as shown in Fig. 4. However, this means considerably more costs, weight and installation space.
It is sufficient to use one piston-motor unit for each brake circuit.
The second advantage is the very low energy requirement and also the design of the motor for impulse operation only. This is achieved by closing the solenoid valves when the setpoint value of the pressure or motor torque is reached and then operating the motor only at low current until a new setpoint value is set by the brake pedal. This makes the energy requirement or the average power extremely small. For example, in a conventional design, motor 3 would draw a high current during emergency braking from 100 km/h. According to the invention, the engine requires only approx. 0.05 s current for the piston travel, which is 1.7 %. If the values are related to the power output, in the conventional case the vehicle electrical system would be loaded with >1000 W for at least 3 s and in the proposed pulsed operation only approx. 50 W of average power. An even greater energy saving is achieved by emergency braking at 250 km/h with braking times of up to 10 s on dry roads. To relieve the impulse load on the vehicle electrical system, a storage capacitor 27 can be used in the power supply here, which can also be used for the other electric motors in accordance with the line with the arrow.
In the pressure line 13, pressure transmitters can be used before or after the solenoid valve, which are not shown, as these correspond to the state of the art.
The piston 1 is supplied with liquid from the reservoir 18 via the sniffer hole. A solenoid valve 19 is switched on in this line. If the piston moves quickly to reduce the pressure, the seal 3 could sniff liquid from the reservoir, especially at low pressures, which is known to be disadvantageous. For this purpose, the low-pressure solenoid valve 19 is switched on and the connection to the reservoir tank is interrupted. This circuit can also be used to achieve negative pressure in wheel circles 15/17, which is beneficial for wheel control at very low friction coefficients, e.g. on wet ice, since no braking torque is generated in the wheel brake. On the other hand, the re-sniffing can be consciously used in case of steam bubble formation, where the piston is already at the stop without the corresponding pressure being reached. In this case the pistons are controlled accordingly by the solenoid valves so that the oscillating piston builds up pressure. If this function is omitted, a sniff-proof seal 3 can be used instead of the solenoid valve 19.
The solenoid valves 14, 16, 19 are controlled by microcontroller 22 via power amplifier 28.
In case of failure of the power supply or the electric motor, the piston is moved by a lever 26 of the actuating device. Between this and the piston there is a clearance which prevents the lever from hitting the piston before the motor moves the piston when the pedal is pressed quickly.
The control function with respect to wheel speed and wheel pressure for ABS / ASR or yaw rate and wheel pressure for ESP has been described in various publications, so there is no need to describe it again. The essential functions of the new system shall be shown in a table:
||Wheel brake 15
||Solenoid valve 14 1
||Wheel brake 17
||Solenoid valve 15 1
||P = constant
||P = constant
||P = constant
||P = constant
||P = constant
||P = constant
The height of the partial flow depends on the speed of pressure increase or decrease desired by the BKV or the brake control. Decisive for this is an extremely small time constant of the electric motor, i.e. a temporally fast torque increase and torque reduction via small moving masses of the entire drive, since the piston speed determines the speed of pressure change. In addition, a fast and precise position control of the pistons is necessary for brake control. In the case of rapid torque reduction, the pressure force from the brake calipers also has a supporting effect, but this is low at low pressures. However, it is precisely here that the rate of pressure drop should also be high in order to avoid large control deviations from the wheel speed on e.g. ice.
This concept has a decisive advantage over conventional pressure control via solenoid valves, since the piston speed determines the rate of pressure change. For example, with a small differential pressure at the outlet valve that determines the pressure reduction, the flow rate and thus the pressure reduction rate is low. As already mentioned, the piston unit can be used separately for each wheel with and without a solenoid valve. In order to take advantage of the low energy consumption, the electric motor would have to be extended with a fast electromagnetic brake, which is more expensive. The shown version with one piston unit and two solenoid valves is preferable in terms of installation space and costs. From a control engineering point of view, however, the restriction applies here that when the pressure on one wheel is reduced, the other wheel cannot build up pressure. However, since the pressure reduction time is approx. < 10% of the pressure build-up time in the control cycle, this restriction is without significant disadvantage. The control algorithms have to be adapted accordingly, e.g. after a phase of constant pressure from opening the solenoid valve, the electric motor has to be excited with a current to which the appropriate pressure in the wheel brake is assigned according to the BKV characteristic curve or is e.g. 20% higher than the preceding locking pressure in the control cycle. Alternatively, e.g. an adaptive pressure level can be applied during the control cycle which is 20% higher than the highest locking pressure of the axle or the vehicle. The locking pressure is the pressure at which the wheel runs unstably with greater slip.
The concept also offers new possibilities for pressure reduction in terms of control technology. In terms of control engineering, the pressure reduction and braking torque reduction are essentially proportional to the rotational acceleration of the wheel, the hysteresis of the seal and inversely proportional to the moment of inertia of the wheel. From these values, the amount of the required pressure reduction can be calculated in each case and the piston can already provide the corresponding volume when the MV is closed, taking into account the described characteristic diagram. When the MV then opens, a very rapid pressure reduction occurs practically into the vacuum. This is based on the fact that the MV has a smaller throttling effect due to the corresponding opening cross-sections in contrast to current solutions. Here, the pressure can be reduced faster than with conventional solutions by means of a specially provided chamber volume in accordance with the pressure volume characteristic curve. Alternatively, pressure can be reduced to a chamber volume that is slightly larger than the required pressure reduction, e.g. by adjusting the piston speed accordingly. For exact control of the pressure reduction, a very short switching time for closing the solenoid valve is necessary here, which can preferably be solved by pre-excitation and/or overexcitation. In addition, for special cases of control, it is advantageous to bring the solenoid armature of the 2/2 solenoid valve into an intermediate position using known PWM methods in order to generate a throttling effect.
The very fast pressure reduction can possibly generate pressure oscillations which affect the wheel. To avoid this damaging effect, the piston travel can be controlled accordingly as a further alternative, e.g. 80% of the required pressure reduction (fast pressure reduction). The remaining required 20% of the pressure reduction can then be achieved slowly by a subsequently controlled slow piston movement or, in the alternative with the pressure reduction control via solenoid valves, by clocking the solenoid valve and stepped reduction. This prevents harmful wheel vibrations. The slow pressure reduction can be continued until the wheel accelerates again with the ABS control.
This allows very small control deviations of the wheel speed. The method described above can also be applied to the pressure build-up. The speeds of the pressure increase can be optimized according to control engineering criteria. Thus the goal can be achieved that the wheel is braked in the immediate vicinity of the friction force maximum and thus optimum braking effect with optimum driving stability is achieved.
Special cases of regulation where a throttling effect is advantageous were mentioned above. This is the case, for example, when a pressure reduction is necessary for both wheels at the same time. In this case, the throttling effect is advantageous until the set piston has provided such a large chamber volume that the pressure can then be quickly released into the vacuum from different pressure levels. A similar procedure can be used, i.e. if the solenoid valves have a built-in throttle in the valve cross-section and pressure is to be built up at both wheel circuits simultaneously. However, the individual alternating pressure build-up is preferable because of the metered pressure build-up with evaluation of the characteristic diagram and controlled adjustment speed of the piston. The same alternating procedure can be used as an alternative to the above mentioned one with the throttling effect for pressure reduction. As a further possibility, the piston can already be retracted with a control signal with a lower response threshold than the control signal for pressure reduction. According to the state of the art, this is the signal at which the controller detects a tendency to lock and switches the MV to pressure hold (see brake manual p. 52-53). This signal is output 5-10 ms before the signal for pressure reduction. The proposed fast drive is capable of providing a chamber volume for 10 bar pressure reduction within approx. 5ms.
Based on the piston position for pressure reduction, the controller can decide whether sufficient chamber volume is available for simultaneous pressure reduction for both wheel brakes.
These remarks show that the concept with the fast and variably controlled electromotive piston drive and the solenoid valve with the evaluation of pressure and characteristic map represents a high potential for the controller, which enables additional braking distance reductions and driving stability.
Fig. 2 shows the complete integrated unit for BKV and control functions. The unit consists of two piston units with associated electric motors and gears according to Fig. 1 for two brake circuits and four wheel brakes. The piston units are located in housing 4. This housing is attached to the front wall 29.
The brake pedal 30 transmits the pedal force and movement via the bearing pin 31 to a fork 32, which acts on the actuating device 33 via a ball joint. This has a cylindrical extension 34 with a rod 35.
Cylinder 34 and rod 35 are mounted in a bushing 37. This bushing accommodates the travel simulator springs 36 and 36a, whereby one spring acts weakly and the other spring acts strongly progressive in force increase. The travel simulator can also be made up of even more springs or rubber elements. This determines the pedal force characteristics. The pedal travel is measured by a sensor 38, which in the drawn example is constructed according to the eddy current principle, in which the rod 35 with a target is immersed. The pedal movement is transmitted to the elements 32 and 33, the piston 34 moves with the rod 35 in the bushing 37. A lever 26 is rotatably mounted on the actuating device, which hits the pistons if the power supply fails. The pedal travel sensor supplies the travel signal to the electronic control unit, which causes the pistons to move via the electric motor according to the BKV characteristic curve, as described in Fig. 7. The parameters of this characteristic curve are described in Fig. 7. A clearance is provided between lever 26 and the two pistons 1 as shown in Fig. 1. The actuating device has an anti-rotation device via pin 39, which is shown offset, and a return spring 40, which supports the pedal return spring not shown. According to the state of the art, many travel simulator solutions are known, which are also partially hydraulically operated by pistons and shut off by solenoid valves if the power supply fails. This solution is complex and hysteresis-prone. Also known are solutions in which the travel simulator travel is lost when the energy supply fails, when the pistons are actuated to generate brake pressure.
The aim of the invention is a simple solution in which the path simulator is switched off if the power supply fails. For this purpose, a counterforce is exerted on the bushing 37 when the power supply is intact via the armature lever 41 with a high transmission ratio and the holding magnet 42. This counterforce is eliminated when the electrical power supply fails. Two-stage levers can also be used to reduce the magnet. This is described in detail in Fig. 3. In this case, the lever comes into contact with the two pistons via the brake pedal after passing through the clearance and can thus transfer the pedal force to the pistons. The pistons are dimensioned in such a way that they generate a pressure at full pedal stroke which still produces a good braking effect, e.g. 80 %. However, the piston stroke is considerably greater than the pedal stroke and can generate much higher brake pressures when the energy supply and electric drive are intact. However, the driver cannot apply the corresponding pedal force. This design is referred to as a transmission ratio jump, which is possible by decoupling the actuation unit with travel simulator from the piston. In conventional design, in which the BKV and master brake cylinder with piston are connected in series, the required pedal force increases up to a factor of 5 for the same wheel brake pressure if the power supply fails. With the new design, for example, the factor can be reduced to 3. This case is relevant, for example, when towing a vehicle if the battery fails.
The lever 26 is pivoted to allow for tolerances in the movement of the pistons, e.g. due to different venting. This compensation can also be limited so that the lever comes to rest on a stop 33a of the actuating device.
However, other error cases must be considered.
Failure of an electric motor.
In this case the amplification and control is fully effective with the adjacent intact piston actuator. Via lever 26, brake pressure is generated in the failed circuit after it is applied at stop 33a. Here, the amplifier characteristic curve of the second circuit can also be increased, which reduces the required pedal force. However, this can also be done without stop.
Failure of one brake circuit.
Here the piston moves to the stop in the housing 4. The intact second circuit is fully effective. Unlike conventional systems of today, there is not a falling pedal, which is known to irritate the driver very much. The irritation can also lead to a complete loss of braking effect if he does not depress the pedal.
Fig. 3 describes the function of the path simulator lock. In borderline cases, the driver can apply high pedal forces, which the lock must apply via the anchor lever 41. To avoid that the magnet 42 with excitation coil 43 has to apply these forces completely, the upper crowned end 41a of the lever engages asymmetrically at the bushing 37. If the pedal is now deflected until the rod 35 hits the floor 37b, this lever action causes a slight twisting of the bushing 37, which creates friction in the guide, whereby the nose 37a can also be supported by the housing 4. Thus the magnetic force can be kept relatively small. The magnet is also designed as a holding magnet 42, so that due to the small air gap a small holding power is necessary. If the power supply fails, the armature lever 41 is deflected by the bushing 37 to the dotted line position 41'. When the actuating device 33 returns to its initial position, the return spring 44 returns the armature lever to its initial position.
The sensor 38 has been moved to the end of the bore of the socket in housing 4, which has advantages for contacting the el. control unit, as shown in Fig. 6. The same applies to the brake light switch 46. In this design example, target 45 is drawn for the eddy current sensor.
The locking of the travel simulator via socket 37 can be changed to avoid the pedal reaction with ABS described in Fig. 7. For this purpose, the lever 41 with its bearing and magnet 42 with receptacle 42a can be moved via an electric motor 60 which drives a spindle 60a via a gear 60b. The lever is mounted on the extension of the spindle and the magnet housing is attached.
Fig. 4 shows a principle representation of a solution with only one electric motor 7a. This description is based on Fig. 1 and Fig. 2. The drive pinion of the motor moves the gear rack 5c, which can also be moved in parallel like Fig. 1. This is connected to a piston 1a which builds up pressure in brake circuit 13a and at the same time moves piston 1a which builds up pressure in brake circuit 13. This piston arrangement corresponds to a conventional master brake cylinder for whose piston and seal designs many variants exist. As in the figures above, the 2/2-way solenoid valves 14, 14a, 15, 15a are arranged in the brake circuits. The ABS pressure modulation takes place in the manner described above. The BKV function is performed by a parallel arranged travel simulation 36 and travel sensor 38. Here, too, a clearance or idle stroke s0 is provided between piston 1a and brake pedal. The brake fluid enters the piston chambers from reservoir 18, 18a. This arrangement is cost-effective. The dynamics of the BKV function in pressure build-up is less than in the dual-motor version, since the electric motor has to provide twice the torque. In addition, the redundancy function of the 2nd motor as described in Fig. 7 is omitted, including a failing pedal in case of brake circuit failure.
Fig. 5 shows the view from the front wall to the integrated unit, whose flange 4b is screwed to the front wall using screws 47. The view shows the actuation unit 33, lever 26 and a pin 39 (not drawn offset) as an anti-rotation device. For size comparison, the outline contour of a 10" vacuum BKV is drawn here. This shows an important advantage in the overall height with the cover 48 of the reservoir. According to the distance A the front wall could be lowered, which is what the designers want. On the left side of the flange, with reference to Fig. 5a, the drive of the rack 5 is drawn as a dashed line. This detail is shown enlarged as Fig. 5a on the right half of the picture. The pinion of the gearwheel 6 meshes with the H-shaped design of the gear rack 5 on both sides. The transverse forces described are supported by the roller 10 or 11 according to Fig. 1 with bearing 10a. For cost reasons, the gear rack can be made of plastic. Since the surface pressure of plastic is not sufficient, hard metal strips 49 are inserted here, which adapt to the rollers when the support is slightly crowned. The gear wheel 7 is pressed into the pinion 6, which is in mesh with the motor pinion. Preferably the pinion is mounted in the motor housing 8a.
Fig. 6 shows the side view of the integrated unit with housing 4, fork 32 for brake pedal 30, actuation unit 33, flange 45, fastening screws 47, cover 48. This view shows the short overall length with the electronic control unit 50 mounted on the front. According to the state of the art, this is connected to the coils or part of the magnetic circuit of the solenoid valves 14 and 16 in order to additionally save contact and electrical connection lines. This feature can be extended by contacting all electrical components such as electric motor 8, solenoid coil 43, travel sensor 38, brake light switch 46, brake fluid level sensor 53 directly with the control unit without electrical connection lines. In this case, the control unit would have to be installed from above direction 50a. However, it is also possible in direction 50b, which results in a different arrangement of the magnetic coil.
The solenoid valves are preferably mounted on a carrier plate 51, since these are pressed into aluminum with high breaking elongation for cost reasons. The screw plugs 52 for the brake lines are screwed into this carrier plate. In the middle part of the control unit the contact is drawn, which contains a redundant power supply in area 54, the bus line in area 55 and the sensors for ABS and ESP in 56.
Fig. 7 shows the essential characteristics of the brake system. Shown are pedal force Fp, brake force pressure p and pedal travel at the actuation unit. Usually, a transmission ratio of 4 to 5 is selected from here to the pedal foot. The pedal travel has its maximum at SP and the pistons, as already mentioned, at a higher value SK. 57 is the so-called pressure-displacement characteristic curve, which here corresponds to a brake circuit, for example. The non-linear curve results from various elasticities such as those of the brake caliper, seals, lines, residual air inclusions and compressibility of the fluid. This line shows the mean value of a scatter band, which is also temperature-dependent, especially in the case of a brake caliper. For this reason, a characteristic map must be created for the current proportional pressure control.
The characteristic curves 59 show the failure of the electric actuator, where the pistons are actuated after the clearance S0. In order to achieve e.g. 100 bar, the described considerably higher pedal forces FPA of approx. 600 N are necessary here, which corresponds to a more than 40% lower pedal force compared to today's solutions.
From the pedal position and the brake pressure, it can be seen that the pressure modulation of 10 bar at locking pressures > 50 bar does not affect the pedal, since the pedal hits the lock at Ss. At lower locking pressures, when the pressure is lowered and built up, there is a reaction on the pedal when the pedal is fully depressed, and is therefore comparable with today's ESP and ABS systems. However, it is possible to reduce or avoid this reaction by using an electric motor 60 described in Fig. 4, which adjusts the locking of the travel simulator via a drive. The pedal is moved back via the piston drive 6 to reduce the pressure. At this point the motor adjusts the drive with small force. This also allows a pedal movement to warn the driver, e.g. in case of traffic jams or similar. Even without this additional motor, a reaction is possible if the pedal movement is greater than the clearance So and the pistons are moved back briefly to warn the driver.
The thicker lines are the booster lines 58 and 58a, which show the assignment of pedal force FP to brake pressure. At approx. 50% of the maximum pedal travel, the travel simulator is fully controlled at SS. This has the advantage that full braking with short pedal travel is possible. The pedal travel is recorded by sensor 38. The assignment of the pressure to the pedal force is freely variable and can, for example, take into account the vehicle deceleration in the engraved line by including it as a correction value in the amplification, so that a higher pressure is applied at the same pedal force when the brake is fading. This correction is also necessary for systems with recuperation of braking energy via the generator, as the braking effect of the generator must be taken into account.
The same applies to panic braking at high pedal speed. Here, a much higher pressure can be fed in disproportionately to the pedal force, which follows the shown static characteristic curve again with a time delay (extended line).
For Fp1, a foot force of 200 N is usually specified for the brake pressure of 100 bar. This pressure corresponds to the blocking limit in dry road conditions. In this range, the travel simulator characteristic is almost linear to ensure good dosing. As a rule, a maximum pressure of 160 bar is sufficient, according to which the endurance strength of the elements is dimensioned. However, a reserve R can be kept for rare loads, which can become effective, for example, if the blocking limit is not yet reached at 160 bar.
The electric drive can be regarded as more fail-safe than the vacuum BKV in the event of a power supply failure, since at least two electric motor drives are used for the proposed invention, i.e. one acts redundantly and, as is known, the total failure rate is λg = λ1 λ2. A failure of the power supply while driving can almost be excluded, since the generator and battery practically do not fail at the same time. A breakdown of the electrical power supply is prevented by the redundant power supply described in Fig. 7. The vacuum BKV is not redundant with amplifying elements, supply lines and, if necessary, pump.
Fig. 8 shows another solution of the piston drive. Instead of the gear rack, a crank rocker 60 can be used, which is connected to the piston via a tie rod 61 via the bearing pin 62. The return spring 9 acts on the crank rocker whose initial position is determined by the stop 65. The crank arm is driven by the engine 11 via a multi-stage gear 63.
Fig. 8a shows a two-armed crank rocker 60 and 60a with two tie rods 61 and 61a. This means that only slight lateral forces act on the piston. The gear 63 is here encapsulated in an extended motor housing 64 and is driven by the drive pinion 11a of motor 11. The advantage of this solution lies in the encapsulation of the gear unit, which allows oil or grease filling, allows helical gearing and is therefore more resilient and quieter.
Fig. 9 shows another alternative with a spindle drive, which is located inside the rotor of the electric motor. This arrangement is known from DE 195 11 287 B4, which refers to electromechanically operated disc brakes. In the solution presented, the nut 67 is located as a separate component in the bore of the rotor 66 and is supported on the flange 66a of the rotor. The spindle drive also acts as a reduction gear, with the spindle 65 transmitting the force to the piston. All the drives shown so far have a reduction gear which is firmly coupled to the piston. If the power supply fails, the reduction gear must be moved by the brake pedal and accelerated by the motor when the pedal is pressed quickly. These mass inertia forces prevent fast pedal actuation and irritate the driver. To avoid this, the nut is axially movable in the bore of the rotor, so that the ball screw drive is switched off when the pedal is engaged. For normal operation with electric motor, the nut is fixed by a 70 mm lever, which is effective when the piston returns quickly, especially when there is a vacuum in the piston chamber. This lever is supported by the shaft 71 in the rotor and, when the motor is not rotating, is moved by the spring 72 to a position where the nut is free. Since the drive motor accelerates extremely fast, centrifugal force acts on the lever and the nut is enclosed by the lever for the movement of the piston.
This movement can also be achieved by a dashed electromagnet, where the lever represents a rotating armature. The torque on the spindle generated by the nut is absorbed by two bearing pins 69 and 69a. The rotor is preferably mounted in a ball bearing 74, which absorbs the axial forces of the piston, and in a plain bearing 75, which can also be a roller bearing. This solution requires a greater overall length, which is clear in comparison with Fig. 9, since the immersion length of the spindle in the nut is equal to the piston stroke. To keep this extension small, the motor housing 74 is flanged directly to the piston housing 4. This has the additional advantage of the different material selection of motor and piston housing.
The nut 67 can also be connected directly to the rotor 66, e.g. by injection. A plastic nut with a low coefficient of friction can be used for the required forces.
In case of motor or energy supply failure, the pedal not drawn acts on the fork according to Fig. 2 and via lever 26 after the idle stroke in such a way on spindle 65 or piston 1. As a blocking of the drive can be eliminated with this solution, the stop 33 can have a smaller distance to the lever. This has the advantage that the pedal force is fully effective on the piston if, for example, an electric motor fails. As soon as the lever rests on the opposite end during rotation, only half the pedal force acts on the piston. In the design, spindle and piston are decoupled, which was not done separately.
Of importance is the return of the piston to the starting position. If the motor fails in an intermediate position, the piston return spring can be additionally supported by a spiral spring 66a, which is located at the end of the rotor 66 and the motor housing 74 and coupled to it. This is intended to compensate for the detent and frictional torque of the motor. This is particularly advantageous for small restoring forces of the pistons, which act on the pedal in case of power failure, in connection with the clutch lever described in Fig. 9.
Fig. 10 shows a further simplified version with an electromotive piston drive, in which piston 1 in turn performs the brake force amplification and pressure modulation for ABS. The piston chambers 1' are connected to the wheel brakes (not shown) and to the solenoid valves (not shown) via lines 13 and 13a as shown in Figs. 1 to 9. The design corresponds to Fig. 8 with spindle drive 65 and with rotor 66, fixed nut 67, separation of motor and piston, housing 74 or 4, piston return springs 9 and bearing pin 69, spiral spring 66a for motor return. The pedal force is transmitted similar to Fig. 2 from a fork 26 to an actuator 34 with rod 35. This is mounted in the motor housing 74 and carries a target 45 in the extension, e.g. for an eddy current sensor 38, which measures the pedal travel. The actuating device is reset by a spring 79. The actuating device 35 is again supported by a lever 26, which carries at the end in the connection to the piston preferably leaf springs 76, which are connected with a displacement transducer 77 in case of a strong leaf spring or with a force transducer 77a in case of a softer spring. In both cases the force transmitted by the lever or pedal is to be measured here. The function of the leaf spring 76 is to avoid a hard reaction before the motor starts running when the pedal is operated. The function is performed in such a way that in a certain function of this pedal force the motors exert a reinforcing force on the piston, whereby this force can in turn be determined from current and piston travel or a pressure transducer. The pedal travel can be processed via the travel sensor 38 in this amplifying function or characteristic curve. This sensor can also take over the amplifier function at the beginning of braking at low pressures in conjunction with the return spring 76. Here the spring 79 takes over the function of the travel simulator spring.
The motor housing has a flange for fastening the unit via the screw bolts 78 in the front wall. This simplified concept does not have the effort of the displacement simulator and locking device. Disadvantages are the limited pedal travel characteristic of the amplifier characteristic, a failure of the pedal in case of brake circuit failure and higher pedal forces in case of amplifier failure, because pedal travel and piston travel are identical. This design is mainly suitable for small vehicles.
In the version shown in Fig. 10, safety valves 80 are drawn in place of all solutions, which become effective if, for example, a piston drive jams when the pedal returns to its initial position. When the pedal is moved, a conical extension of the actuating device 35 actuates the two safety valves 80 which close the connection from brake circuit 13 or 13a to the return flow. This ensures that no brake pressure is built up in the brake circuit when the pedal is in its initial position. These valves can also be electromagnetically actuated.
Safety-relevant systems usually have a separate shutdown option for faults in the output stages, e.g. full current flow due to alloying. For this case a shutdown possibility, e.g. by a conventional relay, is built in. The diagnostic part of the electrical circuit detects this error and switches off the relay, which normally supplies the output stages with power. The concepts proposed here must also include a shutdown option, which is implemented by a relay or a central MOSFET.
In view of the pulse control of the electric motors, a fuse can also be used, as the pulse-off ratio is very high.
The following are examples of design according to the invention.
- Execution example 1:
- Brake system, comprising an actuating device, in particular a brake pedal, and a control and regulating device, wherein the control and regulating device controls an electromotive drive device by means of the movement and/or position of the actuating device, wherein the drive device adjusts a piston of a piston-cylinder system via a non-hydraulic transmission device, so that a pressure is set in the working chamber of the cylinder, wherein the working chamber is connected to a wheel brake via a pressure line, characterised in that in the event of failure of the drive device, the actuating device adjusts the piston (1) or the drive device.
- Execution example 2:
- Brake system according to design example 1, characterized in that a sensor device determines the position of the actuating device.
- Execution example 3:
- Brake system according to design example 1 or 2, characterized in that a device, in particular a haptic device, for presetting or setting a force/displacement characteristic of the actuating device is in operative connection with the latter.
- Execution example 4:
- Brake system according to one of the design examples 1 to 3, characterized in that a valve (14, 16) controlled by the control and regulating device (22) is arranged in the pressure line (13) to the wheel brake (15, 17).
- Execution example 5:
- Brake system according to design example 4, characterized in that the valve (14, 16) closes after the required brake pressure in the brake cylinder (15, 17) has been reached and is opened to set a new brake pressure.
- Execution example 6:
- Brake system according to one of the design examples 1 to 5, characterized in that the piston (1) generates the required pressure change for the brake force booster (BKV) and the anti-lock braking system (ABS).
- Execution example 7:
- Brake system according to one of the design examples 1 to 6, characterized in that a spring (9) applies force to the piston (1) or the drive device, the spring force acting in the direction that the working space is increased.
- Execution example 8:
- Braking device according to one of the design examples 1 to 7, characterized in that the drive device comprises at least one electric motor (8) with, in particular, a small time constant and/or a large acceleration capacity.
- Execution example 9:
- Brake system according to design example 8, characterized in that when the valve (14, 16) is closed, the electric motor (8) is supplied with an excitation current which is sufficient to hold the piston (1) in position against the spring force.
- Execution example 10:
- Brake system according to one of the preceding examples, characterized in that each brake circuit comprises a piston-cylinder system.
- Execution example 11:
- Brake system according to one of the preceding design examples, characterized in that the working chamber (4') is connected to a plurality of brake cylinders (15, 17) via two or more pressure lines (13), one valve (14, 16) being arranged in each pressure line (13).
- Execution example 12:
- Brake system according to one of the design examples 4 to 11, characterized in that the valve (14, 16) is a 2/2-way valve.
- Execution example 13:
- Brake system according to one of the preceding examples of design, characterized in that the piston-cylinder system comprises a first and a second piston (1a, 1b) which are arranged axially displaceably in a cylinder, the first piston (1a) being mechanically coupled to the electromotive drive device (7a, 6, 5c) and the second piston (1b) being hydraulically coupled to the first piston (1a), wherein the two pistons (1a, 1b) form between them a working chamber (4a') which is connected via at least one pressure line (13a) to at least one brake cylinder, and the second piston (1b) forms with the cylinder a second working chamber (4b') which is connected via at least one further pressure line (13) to at least one further brake cylinder.
- Execution example 14:
- Brake system according to design example 13, characterized in that valves (14, 15, 14a, 15a) controlled by the control and regulating device, in particular 2/2-way valves, are arranged in the pressure lines (13, 13a).
- Execution example 15:
- Brake system according to one of the preceding examples of design, characterized in that, when generating the brake booster, the actuating device is not or not in direct mechanical connection with the piston or the drive device, and only in the event of failure of the drive device or upon activation of the ABS is the piston in mechanical connection with the actuating device.
- Execution example 16:
- Brake system according to one of the preceding examples of design, characterized in that two piston-cylinder systems with a respective associated drive device are arranged next to one another, in particular parallel to one another, the actuating device (30) adjusting at least one of the two pistons directly or via interposed means in the event of failure of at least one drive device.
- Execution example 17:
- Brake system according to design example 16, characterised in that the actuating device adjusts a lever or the pivot point of a rocker (26) parallel to the adjustment path of the pistons (1) of the piston-cylinder systems, and each free end of an arm of the rocker (26) is associated with a respective piston (1).
- Execution example 18:
- Brake system according to design example 17, characterized in that a limiting element (33) limits the pivoting range of the rocker (26).
- Execution example 19:
- Brake system according to one of the examples of design 16 to 18, characterized in that the rocker (26) is mounted on a piston (34) which is mounted in a cylinder so as to be displaceable parallel to the pistons (1) driven by the drives, the piston (34) being pressurized in the direction of the brake pedal by means of at least one, in particular non-linear spring (36, 36a), and the spring together with the piston forming a so-called travel simulator, and a sensor determining the position of the piston.
- Execution example 20:
- Brake system according to design example 19, characterized in that the piston stroke of the piston (34) connected to the rocker (26) is limited by a stop, the stop being able to be switched off via an in particular electromagnetic adjusting device.
- Execution example 21:
- Brake system according to one of the preceding design examples, characterized in that a duct connects the working chamber (4') of the piston-cylinder unit to a reservoir (18), the piston (1) closing the duct (20) when it is retracted into the cylinder and the duct (20) being open in the starting position, i.e. only when the piston (1) is almost or completely retracted.
- Execution example 22:
- Brake system according to design example 21, characterized in that a shut-off valve, in particular a 2/2-way valve (19), is arranged in the duct (20).
- Execution example 23:
- Brake system according to design example 22, characterized in that the seal of the piston does not sniff any liquid due to vacuum in the working chamber when the piston is quickly returned from the reservoir.
- Execution example 24:
- Brake system according to one of the preceding examples of design, characterized in that the drive drives a toothed rack (5a) which is mounted displaceably and in particular with low friction parallel to the displacement path of the piston (1), in particular next to the piston, the toothed rack being connected in particular fixedly to the piston (1) via a coupling member (5).
- Execution example 25:
- Brake system according to design example 16, characterized in that a spring (9) applies pressure to the coupling member or to the rack.
- Execution example 26:
- Brake system according to one of the preceding design examples, characterized in that the control system adjusts a corresponding brake force amplification as a function of the movement and/or force application to the brake pedal and/or the driving state and/or braking effect of an electrical machine.
- Execution example 27:
- Brake system according to one of the preceding design examples, characterized in that the control system determines the brake pressure in the working chamber of the cylinder from the drive current of the drive.
- Execution example 28:
- Brake system according to one of the preceding design examples 1 to 16, characterized in that a pressure sensor is provided to determine the brake pressure in the working chamber of the cylinder.
- Execution example 29:
- Braking system according to one of the preceding design examples, characterized in that the control and regulating device has a memory in which a characteristic map with various parameters for controlling the drive is stored.
- Execution example 30:
- Brake system according to one of the preceding examples of execution, characterized in that the control system determines the piston position by means of at least one sensor, in particular an incremental encoder of the electric motor.
- Execution example 31:
- Brake system according to one of the previous examples, characterized in that the actuator moves the piston out of the cylinder so that it comes into mechanical contact with the brake pedal and exerts a force on the brake pedal.
- Execution example 32:
- Brake system according to one of the preceding design examples, characterized in that the control system for generating a rapid pressure reduction in the wheel brake before opening the respective valve generates a vacuum by means of the associated piston by enlarging the working space.
- Execution example 33:
- Braking system according to one of the preceding examples of design, characterized in that the control and regulating device for building up an increased locking pressure before the opening of the respective valve supplies the electric motor of the drive device with approximately 120% of the locking pressure preceding in the control cycle.
- Execution example 34:
- Braking system according to one of the preceding design examples, characterized in that fast energy storage devices for storing electrical energy, in particular capacitors with a large capacity, are provided for generating pulse currents.
- Execution example 35:
- Brake system according to one of the preceding design examples, characterized in that an additional drive adjusts the actuating device or the stop of the travel simulator in such a way that in normal operation the actuating device is not in mechanical connection with the piston.
- Execution example 36:
- Brake system according to design example 35, characterized in that the additional drive acts on a travel simulator, whereby at a low locking pressure the additional drive moves the travel simulator back to the initial position during the pressure reduction, in such a way that the actuating device is not mechanically connected to the piston.
- Execution example 37:
- Brake system according to one of the preceding design examples, characterized in that the control and regulating device pre-excites the valve for rapid closing, so that the valve closes immediately by a small excitation amplification.
- Execution example 38:
- Brake system according to one of the preceding examples of execution, characterized in that the drive device has at least one piston rocker (60, 61) by means of which the piston is adjustable.
- Execution example 39:
- Brake system according to design example 38, characterized in that the piston rocker is a double-armed crank rocker (60, 60a).
- Execution example 40:
- Brake system according to design example 38 or 39, characterized in that the transmission is an encapsulated transmission and in particular is mounted in the motor housing.
- Execution example 41:
- Braking system according to one of the design examples 1 to 37, characterized in that the piston is driven by means of a spindle drive arranged inside the rotor of an electric motor.
- Execution example 42:
- Brake system according to design example 41, characterised in that the rotor drives the piston via a nut mounted axially displaceably in the rotor, the nut being held in axial position by a lever actuated in particular by means of an electromagnet or centrifugal force when the rotor rotates, and the spindle together with the nut being axially displaceable in the rotor if the electrical drive fails.
- Execution example 43:
- Brake system according to design example 41 or 42, characterized in that the spindle is secured against rotation by means of two bearing pins outside the piston, which also receive the piston return springs.
- Execution example 44:
- Brake system according to one of the design examples 41 to 43, characterized in that a torsion spring returns the motor back.
- Execution example 45:
- Brake system according to one of the preceding design examples, characterized in that the brake system adjusts a gain proportional to the pedal force, the brake system determining the pedal force at the piston.
- Execution example 46:
- Brake system according to one of the preceding design examples, characterized in that a damping element, in particular in the form of a leaf spring, is arranged between the actuating device and the respective piston, the leaf spring being arranged in particular on the rocker (26), and in that a force and/or displacement sensor for measuring pedal force is arranged on the rocker or the damping element.
- Execution example 47:
- Brake system according to one of the preceding examples of design, characterized in that a duct connects the working chamber (1') to the reservoir, in which a safety valve (80) is arranged, which opens in the event of a jammed piston and connects the working chamber (1') to the reservoir (18) for pressure reduction in the working chamber.
- Execution example 48:
- Brake system according to design example 47, characterized in that the safety valve is a mechanical-hydraulic or an electromagnetic valve.