||Wheel brake se 15
||Solenoid valve 14 1
||wheel brake 17
||Solenoid valve 15 1
||P = konstant
||P = konstant
||P = konstant
||P = konstant
||P = konstant
||P = konstant
 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 movable measures 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 a 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 with a pressure drop. If you build a wheel on one wheel, 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 opening the solenoid valve, the electric motor must be energized 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 necessary 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 be solved preferably 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.
 The above mentioned special cases of regulation where a throttling effect is advantageous. 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 electromo- toric 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, whereby one spring acts weakly and the other spring acts strongly progressive in the increase of force. The travel simulator can also be made up of even more springs or rubber elements. This determines the pe- dal 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 piston in case of a 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 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.
 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 error cases 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 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 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 manner described above. The BKV function is performed by a parallel arranged displacement simulation 36 and displacement sensor 38. also here 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 using screws 47. Here, the actuation unit 33, lever 26 and a pin 39, which is not offset, are shown 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 described considerably higher pedal forces FPA of approx. 600 N are necessary here, which corresponds to a pedal force that is more than 40% lower than 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 a small 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 becomes 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 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.
 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 mounting 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 of all solutions. These valves become effective, for example, if 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 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.
 Considering the pulse control of the electric motors, a fuse can also be used, since the pulse-off ratio is very high.
 The following are examples of design according to the invention.
Execution example 1:
 Brake system, having an actuating device, in particular a brake pedal, and a control and regulating device, the control and regulating device controlling an electromotive drive device on the basis of the movement and/or position of the actuating device, the drive device adjusting 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, the working chamber being connected to a wheel brake via a pressure line, characterised in that if the drive device fails, 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, whereby the spring force acts in the direction that increases the working space.
Execution example 8:
 Brake system according to one of the design examples 1 to 7, characterized in that the drive device has at least one electric motor (8) with, in particular, a small time constant and/or a high 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 has a piston-cylinder system.
Execution example 11:
 Brake system according to one of the preceding 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 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 design examples, 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), in particular 2/2-way valves, controlled by the control and regulating device are arranged in the pressure lines (13, 13a).
Execution example 15:
 Brake system according to one of the preceding design examples, characterized in that, when generating the brake force amplification, the actuating device is not or not directly mechanically connected to the piston or the drive device, and the piston is in mechanical connection with the actuating device only in the event of failure of the drive device or upon activation of the ABS.
Execution example 16:
Brake system according to one of the preceding examples, characterized in that two piston-cylinder systems each with an 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 intermediate means in the event of failure of at least one drive device.
Execution example 17:
 Brake system according to design example 16, characterized 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 assigned to one piston (1) in each case.
Execution example 18:
 Brake system according to design example 17, characterized in that a limiting element
(33) limits the swivel range of the rocker (26).
Execution example 19:
Brake system according to one of the design examples 16 to 18, characterized in that the tip (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 closes the channel (20) when entering the cylinder and the channel (20) is open in the initial 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 channel (20).
Design example 23
0099] Brake system according to design example 22, characterized in that the piston seal does not sniff any fluid 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 design examples, 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:
0101] Brake system according to design example 16, characterized in that a spring (9) pressurizes the coupling member or the rack.
Execution example 26:
0102] Brake system according to one of the preceding design examples, characterized in that the control system regulates a corresponding brake force amplification as a function of the movement and/or force application of the brake pedal and/or the driving state and/or braking effect of an electric machine.
Execution example 27:
0103] 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:
0104] 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:
0105] 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 design examples, 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 preceding examples, characterized in that the drive moves the piston out of the cylinder so that it mechanically engages 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 the opening of the respective valve generates a vacuum by means of the associated piston by enlarging the working chamber.
Execution example 33:
 Brake system according to one of the preceding design examples, 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 approx. 120% of the preceding locking pressure in the control cycle.
Execution example 34:
 Brake 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, characterised in that the additional drive acts on a travel simulator, whereby at a low locking pressure the additional drive moves the travel simulator back into 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, 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 arm is a double-arm crank rocker arm (60, 60a).
Execution example 40:
 Brake system according to design example 38 or 39, characterized in that the gear unit is an encapsulated gear unit and is, in particular, mounted in the motor housing.
Execution example 41:
 Brake 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 is rotated, 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 torsionally secured via 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, characterized in that a channel connects the working chamber (1' ) with 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' ) with 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.