The present invention relates to a 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 on the basis 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.
State of the art:
 Modern braking systems consist of brake force amplification, i.e. conversion of the pedal force into a correspondingly amplified braking torque at the wheel brakes and brake force control via open or closed control 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 widely used method is to divide the brake force booster (BKV) or brake force control and brake force regulation into units in a hydraulic unit (HE). This configuration is mainly used for systems such as anti-lock braking systems (ABS), anti-slip systems (ASR), electronic stability programs (ESP) or electrohydraulic brakes (EHB).
 The hydraulic unit (HE) consists of solenoid valves, multi-piston pumps for 2-circuit braking systems, electric motor for pump drive, hydraulic accumulator and several pressure transmitters. The pressure is controlled in such a way that, in order to reduce braking torque, pressure fluid is drained from the wheel brakes into an accumulator via solenoid valves and pumped back into the master brake cylinder by the pump, which causes a pedal movement. Both pressure increase and decrease is controlled by solenoid valves, with pressure transmitters being used in some cases for solenoid valve control. Except for the EHB, the brake force is amplified by 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 force amplification 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 for the power supply is necessary to ensure the braking capability of a brake circuit in the event of a fault.
 The 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 currently not solved. 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 speed of pressure reduction is not yet optimal because it is highly non-linear. In addition, with a µ-jump or with a small coefficient of friction, the pressure reduction speed 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 system and control device, which controls the output pressure of an electro-hydraulic servo device directly connected to the brake circuit to a value determined by the reference variable. In case of failure of the control device or the servo device itself, 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 brake system is familiar in which an electric motor drives the pistons of a piston cylinder system via a spin-drive. The pistons are not firmly coupled to the spindle, so that the maximum piston speed when the spindle is retracted and thus the maximum speed of pressure reduction 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.
 DE 3723916 A1 shows a brake system with a hydraulic brake booster, which in addition to pure brake booster also realizes the ABS function. 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 195 00544 A1, an electronically controllable brake actuation system for anti-lock motor vehicle brake systems is well-known, in which a master brake cylinder can be actuated by means of a brake pedal. By means of a sensor, the actuation travel of the brake pedal is determined, which represents an input value for a control unit that 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 sensors 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 already known from DE 4239386 A1.
 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 a piston of a piston-cylinder system driven by an electric motor, 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.
 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 can be adjusted by brake speed, 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.
 DE 695 15 272 T2 is a brake system in which a piston position is adjusted depending on the pedal position. The piston position is adjusted by setting a current, whereby piston position errors are detected by corresponding sensors.
 Based on the DE 195 00 544 A1, the task is to provide an improved brake system.
 This task is advantageously solved by a brake system with the features 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 characterized by the fact that it realizes the brake force amplification and the servo device on smallest space per brake circuit by means of only one piston-cylinder unit. The piston-cylinder unit is used to build up and reduce brake pressure, to implement ABS and anti-slip control, as well as in the event of power supply failure or malfunction of the drive unit. This results in a small, integrated and cost-effective unit for brake booster and control, which saves installation space, installation 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.
 The advantageous provision of a sensor system as well as a travel simulator allows a variable pedal characteristic such as brake-by-wire function, i.e. brake pressure increase independent of pedal actuation, also taking into account the braking effect of the generator with recuperable brakes.
 Furthermore, with the corresponding design, there is no disadvantageous falling 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 reducing 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 expenditure including the electrical connections also results in a reduction of the failure rate.
 The electromotive drive also makes it possible to improve ABS/ESP control by means of finely dosed pressure control with variable pressure rise and, in particular, pressure drop speeds. It is also possible to reduce the pressure below 1 bar in the vacuum range for operation with the smallest frictional force coefficients, e.g. wet ice. Likewise, 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 booster and the control function, the brake system requires considerably less energy.
 It is also possible to provide a separate piston-cylinder system for each brake circuit or each 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 one cylinder, whereby the cylinders are hydraulically coupled and only one piston is mechanically driven by the drive device by means of an electric motor.
 In the following, various designs of the brake system according to the invention are explained in more detail using drawings.
 Show it:
- 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 brake system according to the invention;
- Fig. 4: a piston-cylinder system with one cylinder and two pistons;
- Fig. 5 and Fig. 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 Fig. 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 that is responsible for generating pressure or boosting braking power. 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 a vacuum in piston chamber 4'. This gear rack 5a transmits the force to the front crowned end of the piston 1, which has a collar pin 1a at this point, via which the gear rack 5a with return spring 9 brings the piston into the 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 gear rack is supported in rollers 10 and 11 with sliding piece 12. Fig. 1 clearly shows that the parallel arrangement of the gear 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 (dashed line) with the greatest bending force due to the offset pressure force. The rack is driven via tooth profile 5a' and gear 6 via gear 7 by the pinion of motor 8. The 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, which 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 applied 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 correspond completely, 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 pe- dal 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 or 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 only for pulse operation. 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 needs only approx. 0.05 s current for the piston stroke, 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 results when braking at 250 km/h with braking times of up to 10 s on dry roads. In order 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 according to 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.
 Piston 1 is supplied with liquid from 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. 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, this can be deliberately used in the case of steam bubble formation, where the piston is already at the stop without the corresponding pressure being reached. Here the pistons are controlled 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 lever and the piston there is a clearance which prevents the lever from hitting the piston before the motor moves the piston.
 The control function with respect to wheel speed and wheel pressure for ABS / ASR or yaw rate and wheel pressure for ESP has been presented in various publications, so that a new description will be provided. The essential functions of the new system will 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. The decisive factor here is an extremely small time constant of the electric motor, i.e. a fast torque increase and torque reduction over 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. For fast torque reduction, the pressure force from the brake calipers also provides support, 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 small differential pressure at the outlet valve that determines the pressure reduction, the flow rate and thus the pressure reduction speed 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 however 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 limitation is without significant disadvantage. The control algorithms must be adapted accordingly, e.g. after a phase of constant pressure from the opening of the solenoid valve, the electric motor can be excited with a current to which the appropriate pressure in the wheel brake is assigned according to the BKV characteristic curve or, for example, is 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 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 inertia moment of the wheel. From these values, the amount of the required pressure reduction can be calculated and the piston can provide the corresponding volume when the MV is closed, taking into account the described characteristic diagram. When the MV opens, the pressure is reduced very quickly into the vacuum. This is based on the fact that the MV has a smaller throttle 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 the exact regulation of the pressure reduction a very small 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 as a further alternative, e.g. 80% of the required pressure reduction (fast pressure reduction). The remaining 20% of the required 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. Accordingly, 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 dosed 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 volumes are available for simultaneous pressure reduction for both wheel brakes.
 These remarks show that the concept with the fast and variably controlled electromotoric piston actuator and the solenoid valve with the evaluation of the pressure and characteristic diagram represents a high potential for the controller, which enables additional reductions in braking distance and driving stability.
 Fig. 2 shows the entire 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, one spring being weak and the other spring being 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 example shown is based on the eddy current principle, into 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 in case of power failure. 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 path simulator path is lost when the energy supply fails and 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 in case of power failure. 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. 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 longer 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 the travel simulator from the piston. In conventional design, in which the BKV and master brake cylinder with piston are connected in series, the pedal force required for the same wheel brake pressure increases up to a factor of 5 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.
 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, there are other cases of error to be considered.
Failure of an electric motor.
 In this case the amplification and control is fully effective with the adjacent intact piston drive. Brake pressure is generated in the failed circuit via lever 26 after it is applied to stop 33a. In this case, the amplifier characteristic of the second circuit can be increased additionally, which reduces the required pedal force. However, this can also be done without a 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. 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 travel simulator lock. In borderline cases, the driver can apply high pedal forces, which the locking device must apply via the anchor lever 41. In order to avoid that the magnet 42 with excitation coil 43 has to apply these forces fully, 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.
 Sensor 38 has been moved to the end of the bore of the socket in housing 4, which has advantages for the contact with the el. control unit, as shown in Fig. 6. The same applies to the brake light switch 46. In this design example, target 45 for the eddy current sensor is drawn.
 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 holder 42a can be moved by 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 the brake circuit 13a and at the same time moves piston 1a which builds up pressure in the 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 way described above. The BKV function is performed by a parallel arranged displacement simulation 36 and displacement sensor 38. Here, too, a clearance or idle stroke s0 is provided between piston 1a and brake pedal. The brake fluid flows from reservoir 18, 18a into the piston chambers. This arrangement is cost-effective. The dynamics of the BKV function in pressure build-up is less than in the variant with two motors, 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 with screws 47. Here, the actuation unit 33, lever 26 and a pin 39, which is not offset, can be seen as an anti-rotation device. For size comparison, the outline of a 10" vacuum BKV is drawn here. This shows an important advantage in the overall height with the cover 48 of the storage tank. 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 figure. The pinion of the gear 6 engages the H-shaped design of the gear rack 5 on both sides. The described transverse forces 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 rolls 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 control unit is connected to the coils or part of the magnetic circuit of the solenoid valves 14 and 16 in order to save additional contact and electrical connection lines. This feature can be extended by connecting all electrical components such as electric motor 8, solenoid coil 43, travel sensor 38, brake light switch 46, brake fluid level sensor 53 directly to the control unit without electrical connection lines. In this case, the control unit would have to be installed from above in 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 elongation at break from cost green. 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. It shows pedal force FP, brake pressure p and pedal travel at the actuation unit. Usually a transmission ratio of 4 to 5 to the pedal foot is selected from here. The pedal travel has its maximum at SP and the pistons, as already mentioned, at a higher value sK. The so-called pressure/stroke characteristic curve is shown at 57, which here corresponds to a brake circuit, for example. The non-linear course results from different 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 the brake caliper. Therefore, a characteristic map must be created for the pressure control proportional to the current.
 The characteristic curves 59 show the failure of the electric actuator, in which the pistons are actuated after the clearance s0. In order to achieve e.g. 100 bar, the considerably higher pedal forces FPA of approx. 600 N described here are necessary, 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 encounters the locking 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. Piston drive 6 is used to move the pedal back to reduce pressure. At this time, the motor adjusts the drive with little force. This also allows the pedal to be moved to warn the driver, e.g. in case of a traffic jam or similar chemical problems. 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, since the braking effect of the generator must be taken into account. The same applies to panic braking with 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).
 FP1 usually specifies a foot force of 200 N 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 curve 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 during the journey can almost be excluded, since generator and battery at the same time practically not fail. A break in the electrical power supply is prevented by the redundant power supply described in Fig. 7. The vacuum BKV is not redundant with amplifying elements, feeders and, if necessary, pump.
 Fig. 8 shows another solution for the piston drive. Instead of the gear rack, a crank arm 60 can be used which is connected to the piston via a tie rod 61 and bearing pin 62. The return spring 9 acts on the crank arm, 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 electro-mechanically 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, whereby the spindle 65 transmits 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 is 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 the event of a motor or power supply failure, the undrawn pedal acts on the clutch as shown in Fig. 2 and on the spindle 65 or piston 1 via lever 26 after the idle stroke. Since blocking of the drive can be eliminated with this solution, the stop 33 can be at a smaller distance from 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.
 The return of the piston to the starting position is important. If the motor fails in an intermediate position, the piston return spring can be additionally supported by a spiral spring 66a, which is arranged 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 electric motor-driven piston drive, in which piston 1 in turn performs the brake force amplification and pressure modulation for ABS. According to Fig. 1 to Fig. 9, the piston chambers 1' are connected to the wheel brakes (not shown) via lines 13 and 13a and to the solenoid valves, which are also not shown. 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. A lever 26 is mounted on the actuating device 35, which carries at the end in the connection to the piston preferably leaf springs 76, which are connected to a travel sensor 77 in case of a strong leaf spring or to a force sensor 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 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. A disadvantage is the limited pedal travel characteristic of the amplifier curve, a failure of the pedal in case of brake circuit failure and higher pedal forces in case of amplifier failure, since 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 as representatives for all solutions. These valves become effective, for example, if a piston drive jams when the pedal returns to the 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 the 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 realized by a relay or a central MOSFET.
 In view of the pulse control of the electric motors, a fuse can also be used because the pulse-off ratio is very high.