BRAKE SYSTEM WITH ELECTROMOTIVELY DRIVEN PISTON/CYLINDER SYSTEM

[0001] 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 , 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:

[0002] Modern brake systems consist of a brake booster, i.e. conversion of the pedal force into a corresponding increased braking torque at the wheel brakes, and brake force control via open or closed control circuits. With a few exceptions, hydraulic lines are used in the passenger car sector to transmit the pedal force to generate the brake pressure.

[0003] A common approach is to divide the hydraulic unit (HU) into units for brake force boosting (BKV) or brake force control and brake force regulation. This configuration is used primarily in systems such as anti-lock braking systems (ABS), traction control systems (TCS), electronic stability programs (ESP) and electro-hydraulic braking (EHB).

[0004] The hydraulic unit (HU) consists of solenoid valves, multi-piston pumps for dual-circuit brake systems, an electric motor for driving the pump, a hydraulic accumulator and several pressure transmitters. The pressure control is carried out in such a way that, to reduce the braking torque, pressure medium is released from the wheel brakes into an accumulator via solenoid valves and pumped back from the pump into the main brake cylinder, which causes a pedal movement. Both the pressure increase and decrease are controlled by solenoid valves, in which pressure transmitters are used in part for solenoid valve control. Except for the EHB, the brake booster is provided by the vacuum BKV, which partly contains switching means and sensors for the so-called brake assistant function and also for the detection of the so-called control point. As a source of energy for the vacuum, the combustion engine is used in gasoline engines, but as a direct injector, especially at higher altitudes, it only provides a weak vacuum. In diesel engines, a mechanically or electrically driven vacuum pump is used. The latest ESP systems are able to achieve additional brake boost or, if the BKV fails, a brake boost with a longer time constant by switching the solenoid valves and pump. These systems and functions are described in detail in the 2003 edition of the braking manual published by Vieweg Verlag.

[0005] In the mid-1980s, Teves used the Mark II and Bosch used the ABS3, which, as integrated units, contained all components for brake booster and control with hydraulic BKV, see Automotive Engineering Manual Bosch 1986, 20th edition. For cost reasons, these systems have not caught on, except for use in armored vehicles. The same applies to fully electric braking systems, so-called EMB, with electric motors on the wheel brakes, which have been developed intensively in conjunction with the 42 V electrical system. In addition to the additional costs, a new redundant electrical system is necessary for the energy supply in order to ensure the braking capability of a brake circuit in the event of a fault.

[0006] The wedge brake with electric motor drive also belongs to the category of EMB systems. Despite the lower energy requirement, a redundant on-board electrical system is also necessary for this. The design of the wedge brake, which requires additional rollers for hysteresis reasons, necessitating integration into the brake caliper, has not yet been resolved. The wedge brake with its electric motor drives with sensors must withstand the harsh environmental conditions (dust, water, high temperatures).

[0007] The systems for BKV and HE are very well developed, in particular the control and regulating functions for ABS to ESP. For example, the pressure-controlled operation of the solenoid valves allows a very fine metering of the brake pressure, which also enables a variable braking force coordination EBV. The pressure reduction speed is not yet optimal because it is highly non-linear. Furthermore, with a µ-jump or with a small friction coefficient, the pressure reduction speed is determined by the relatively low pump power, which leads to large control deviations and thus to a loss of braking distance.

[0008] A braking system is known from DE 3342552. In this braking system, the main brake cylinder is used to generate a pedal-dependent pressure, which serves as a reference variable for an electronic control and regulating device, which regulates the output pressure of an electro-hydraulic servo device, which is directly connected to the brake circuit, to a value determined by the reference variable. If the control device or the servo device itself fails, the pressure in the brake circuit is generated by the main cylinder. Instead of the command variable generated by the master brake cylinder during normal operation, it is possible to allow a command variable generated by an anti-lock braking system or by a traction control system of the motor vehicle to act on the electronic control and regulating device and thus on the electro-hydraulic servo device. The servo device has an electrically operated hydraulic piston-cylinder unit, the working chamber of which is connected to the brake circuit and the piston of which can be adjusted axially by means of an electric motor. The rotary movement of the electric motor is converted into a longitudinal movement of the piston via a spindle connected to the piston.

[0009] A braking system is known from WO2004/005095 A1, in which an electric motor drives the pistons of a piston-cylinder system via a spindle 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 pressure reduction speed is determined by the strength of the pressure springs in the piston-cylinder system. The brake pressure to be set in the wheel brakes is determined by means of a pressure sensor, with the pressure being the controlled variable of the brake pressure control.

[0010] DE 3723916 A1 shows a brake system with a hydraulic brake booster, which in addition to the 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, which is open to change the pressure in the wheel brake and closed to hold the wheel brake pressure. In this brake pressure control, too, the pressure is the controlled variable.

[0011] An electronically adjustable brake actuation system for anti-lock automotive brake systems is known from DE 195 00544 A1, in which a brake pedal is used to actuate a master brake cylinder. A sensor is used to determine the distance the brake pedal has been depressed, which is an input variable for a control unit that controls several braking pressure transmitters, 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 means of a valve device. In order to achieve an increase in functional safety, in particular in the event of an electrical defect or failure of the vehicle electronics, the piston of the master brake cylinder can be adjusted directly by means of the brake pedal to build up pressure in the wheel brakes, with the valve device being open for this purpose. The brake pressure generators each have an electric drive, which adjusts a piston in a cylinder so that a pressure is set in the brake circuit, which is determined by means of a pressure sensor and supplied to the control unit as an input variable. The pressure is also the controlled variable in this brake pressure control. A similarly operating brake system is known from DE 4239386 A1.

[0012] A braking system for motor vehicles is known from DE 4445975 A1, in which the braking pressure in a wheel brake is regulated by means of an electromotively driven piston of a piston-cylinder system, wherein a pressure sensor is also provided in this braking system to measure the controlled variable. A 2/2-way valve is used to maintain the braking 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.

[0013] DE 10318401 A1 discloses an engine-driven vehicle braking device in which the position of the brake pedal is determined by means of a position sensor and transmitted to a control unit. The control unit controls an electromotive drive of a piston-cylinder system, which serves to build up pressure in the brake circuits, as a function of the driving condition and the brake pedal position. No mechanical connection is provided between the piston of the piston-cylinder system and the brake pedal, so that, as a fallback, 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.

[0014] DE 19936433 A1 and DE 10057557 A1 disclose brake systems in which an assisting force can be applied to the piston of the master brake cylinder, which is adjustable by means of the brake pedal, by means of electromagnetic drives. In these brake systems, too, the pressure in the master brake cylinder is the controlled variable of the brake pressure control process.

[0015] DE 695 15 272 T2 describes a brake system in which a piston position is set depending on a pedal position. The piston position is set by specifying a current, with piston position errors being detected by corresponding sensors.

[0016] Based on DE 195 00 544 A1, the task is to provide an improved brake system.

[0017] This task is advantageously solved by a brake system with the features of claim 1. Further advantageous configurations of the brake system according to claim 1 result from the features of the subclaims.

[0018] The advantage of the braking system according to the invention is that it realizes the brake force boosting and the servo device in the smallest space per brake circuit by means of only one piston-cylinder unit. The piston-cylinder unit is used for building up and releasing brake pressure, for implementing the ABS and anti-skid control, and in the event of a power failure or malfunction of the drive unit. This results in a small, integrated and cost-effective unit for brake booster and control, which saves on installation space, assembly costs and additional hydraulic and vacuum connection lines. In addition, due to the short overall length, the spring dome, for example, does not act on the main cylinder and the pedal mechanism in the event of a frontal crash.

[0019] The advantageous provision of a sensor system and a travel simulator allows a variable pedal characteristic such as a brake-by-wire function, i.e. brake pressure increase freely variable independently of pedal actuation, also taking into account the braking effect of the generator in the case of recuperable brakes.

[0020] Furthermore, in the corresponding design, there is no adverse brake pedal drop in the event of a drive failure, since the pedal acts directly on the system piston. 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 path with intact and failed amplification. This is referred to as a transmission step, which reduces the pedal force for the same braking effect by up to 40 %. The reduction of the overall effort, including the electrical connections, also results in a favorable reduction of the failure rate.

[0021] The electric motor drive also makes it possible to improve the ABS/ESP control by means of finely dosed pressure control with variable pressure increase and, in particular, pressure drop rates. It is also possible to lower the pressure to below 1 bar in the vacuum range for operation at the lowest friction force coefficients, e.g. wet ice. Likewise, a rapid pressure increase at the start of braking, e.g. 0 – 100 bar in less than 50 ms, can be achieved, which results in a significant reduction in braking distance.

[0022] Due to the advantageous provision of a 2/2-way valve for the brake booster and the control function, the brake system according to the invention requires considerably less energy.

[0023] It is also possible to provide a separate piston-cylinder system with a dedicated drive for each brake circuit or wheel brake. Likewise, it is possible to use a piston-cylinder system in which two pistons are arranged axially displaceable in a cylinder, wherein the cylinders are hydraulically coupled and only one piston is mechanically driven by the drive device in the form of an electric motor.

[0024] Various configurations of the braking system according to the invention are explained in more detail in the following drawings.

[0025] Showing:

• Fig. 1: a first embodiment of a brake system with a brake circuit for two wheel brakes;
• Fig. 2: a second embodiment of the brake system with two piston-cylinder systems for two brake circuits for two wheel brakes in each case;
• Fig. 3: a path simulator for the brake system according to the invention;
• Fig. 4: a piston-cylinder system with a cylinder and two pistons;
• Fig. 5 and Fig. 5a: connection between the actuation device and the piston-cylinder systems;
• Fig. 6: a side view of the integrated unit with housing;
• Fig. 7: characteristic curves of the braking 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

[0026] Fig. 1 shows a section of the integrated unit responsible for generating pressure and boosting the braking force. Here, the piston 1 with the usual seals 2 and 3 in the cylinder housing 4 is moved parallel to the piston via a specially designed rack 5a. Seal 2 is designed to seal even when there is a vacuum in piston chamber 4′. This rack 5a transfers the force to the front ball-shaped end of piston 1 . At this point, the piston has a collar bolt 1a , which the rack 5a uses to return the piston to the initial position with return spring 9. Here the rack is in contact with the cylinder housing 4a. The advantage of this external spring is that the cylinder is short and has little dead space, which is advantageous for venting. Due to the transverse forces, the rack is supported in the rollers 10 and 11 with a sliding block 12. Fig. 1 clearly shows that the rack is arranged parallel to the piston, resulting in a short overall length. The assembly must be very short to be outside the crash zone. The rack is very resistant to bending thanks to the H-profile shown in Fig. 5a. The rollers are arranged so that the rack has a relatively small bending length in the end position 5b (shown dashed) with the greatest bending force due to the offset compressive force. The rack is driven by the pinion of the motor 8 via the gear wheel 6 and the gear wheel 7. This motor with a small time constant is preferably a brushless motor as a bell-type rotor with an ironless winding or preferably a motor according to the PCT patent applications PCT/ EP2005/002440 and PCT/ EP2005/002441. This is controlled by the output stages 21, preferably via three strands, 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 meters, the position of the piston. In addition to controlling the motor, the current and position measurement is used for indirect pressure measurement, since the motor torque is proportional to the pressure force. To do this, a characteristic map must be created in the vehicle when it is put into operation and also during operation, in which the position of the piston is assigned to the various flow rates. During operation, the piston is then moved to a position that corresponds to a certain pressure according to the characteristic map, in line with the amplifier characteristic described later. If the position and motor torque do not quite match, e.g. due to temperature influence, the characteristic map is adapted during operation. As a result, the characteristic map is continuously adapted. The initial characteristic map is formed from, preferably, the pressure-volume characteristic of the wheel brake, engine characteristic value, transmission efficiency and vehicle deceleration. With the latter, a pedal-force-proportional vehicle deceleration can be achieved so that the driver does not have to adjust to different braking effects.

[0027] Piston 1 generates a corresponding pressure in line 13, which is applied to wheel brake 15 via 2/2-way solenoid valve (MV) 14 or to wheel brake 17 via solenoid valve MV 16. This arrangement has several advantages. Instead of the two low-cost small solenoid valves, a further piston-motor unit could be used as shown in Fig. 4. However, this means significantly higher costs, weight and space requirements.

[0028] It is sufficient to use a piston-motor unit for each brake circuit.

[0029] The second advantage is the very low energy requirement and also the design of the motor only for pulsed operation. This is achieved by closing the solenoid valves when the setpoint pressure or motor torque is reached and then operating the motor only at a low current until a new setpoint is specified by the brake pedal. This makes the energy requirement or average power extremely small. For example, in a conventional design, the motor 3 would draw a high current during an emergency stop from 100 km/h. According to the invention, the motor requires only approx. 0.05 s of current for the piston stroke, which amounts to 1.7 %. If the values are related to the power, in the conventional case the on-board power supply would be loaded with >1000 W for at least 3 s and with the proposed pulse operation only approx. 50 W of average power. An even greater energy saving results from an emergency braking maneuver from 250 km/h with braking times of up to 10 s on a dry road. To relieve the impulse load on the on-board power supply, a storage capacitor 27 can be used in the power supply, which can also be used for the other electric motors, as shown by the arrow in the line.

[0030] Pressure transmitters, which are not shown because they correspond to the state of the art, can be used in the pressure line 13 before or after the solenoid valve.

[0031] The piston 1 is supplied with liquid from the reservoir 18 via the snifting hole. A solenoid valve 19 is connected in this line. If the piston moves quickly to reduce the pressure, the seal 3 could snort liquid out of the reservoir, especially at low pressures, which is known to be disadvantageous. To prevent this, the low-pressure solenoid valve 19 is switched on and the connection to the reservoir is interrupted. This circuit can also be used to achieve a vacuum in wheel circuits 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, snifting can be deliberately used when vapor locks are forming, in which the piston is already at the stop without the corresponding pressure being reached. In this case, the pistons are controlled by the solenoid valves so that the oscillating piston builds up pressure. If this function is not used, a snifting-proof seal 3 can be used instead of solenoid valve 19.

[0032] The solenoid valves 14, 16, 19 are controlled by the microcontroller 22 via output stages 28.

[0033] In the event of a power failure or failure of the electric motor, the piston is moved by a lever 26 of the actuating device. A play is built in between this and the piston, which prevents the lever from hitting the piston before the motor moves the piston when the pedal is depressed quickly.

[0034] The control function with regard to wheel speed and wheel pressure in the case of ABS/ASR or yaw rate and wheel pressure in the case of ESP has been described in various publications, so that it will not be described again here. The main functions of the new system should be shown in a table:

Functions Electric motor Wheel brake 15 Solenoid valve 14 Wheel brake 17 Solenoid valve 15
1 1
One structure 0 Structure 0
BKV partially energized P = constant 1 P = constant 1
partially energized Removal 0 Removal 0
One structure 0 Structure 0
partially energized P = constant 1 P = constant 0
Brake control One structure 0 P = constant 1
partially energized Removal 0 P = constant 1
partially energized Removal 0 Removal 0

[0035] The level of partial energization depends on the pressure increase or reduction speed desired by the BKV or the brake control. Crucial for this is an extremely small time constant of the electric motor, i.e. a fast instantaneous increase and torque reduction over small moving masses of the entire drive, since the piston speed determines the pressure change speed. In addition, fast and precise position control of the pistons is necessary for brake control. During the fast torque reduction, the compressive force from the brake caliper also has a supporting effect, but this is low at low pressures. However, it is precisely here that the pressure drop rate should also be high in order to avoid large control deviations of the wheel speed on, for example, ice.

[0036] 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, if the differential pressure at the outlet valve that determines the pressure drop is small, the flow rate and thus the rate of pressure drop is low. As already mentioned, the piston unit can be used separately for each wheel with or without a solenoid valve. To take advantage of the low energy consumption, the electric motor would have to be equipped with a fast electromagnetic brake, which is more expensive. The design shown with a piston unit and two solenoid valves is preferable in terms of space and cost. However, in terms of control technology, there is a restriction here: if there is a pressure drop at one wheel, the other wheel cannot build up pressure. However, since the pressure drop time is approx. < 10% of the pressure build-up time in the control cycle, this restriction is not a 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 must be energized with a current that is assigned the appropriate pressure in the wheel brake according to the BKV characteristic or is 20% higher than the previous locking pressure in the control cycle. Alternatively, an adaptive pressure level can also be introduced during the control, for example, 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 unstable with increased slip.

[0037] The concept also offers new possibilities for pressure reduction in terms of control technology. In terms of control technology, the pressure reduction and braking torque reduction are essentially proportional to the rotational acceleration of the wheel, the hysteresis of the seal and inversely proportional to the moment of inertia of the wheel. The amount of pressure reduction required can be calculated from these values and the piston can already provide the corresponding volume when the MV is closed, taking into account the described characteristic diagram. When the MV then opens, the pressure is reduced very quickly, practically into a vacuum. This is based on the assumption that the MV has a smaller throttling effect than current solutions due to the corresponding opening cross-sections. In this case, the pressure drop can take place faster than with conventional solutions via a specially provided chamber volume according to the pressure volume characteristic. Alternatively, a pressure drop into a chamber volume that is slightly larger than the necessary pressure drop is possible, e.g. by adjusting the piston speed accordingly. To control the pressure reduction accurately, a very short switching time is required here to close the solenoid valve, which can preferably be achieved by pre-excitation and/or overexcitation. Furthermore, for special control cases, it is advantageous to bring the armature of the 2/2 solenoid valve to an intermediate position using known PWM methods in order to create a throttle effect.

[0038] The very fast pressure reduction can possibly generate pressure oscillations that react on 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 pressure reduction required can then be carried out slowly by a subsequent controlled slow piston movement or, in the case of the alternative with pressure reduction control via solenoid valves, by pulsing the solenoid valve and staged reduction. This is how harmful wheel vibrations are avoided. The slow pressure reduction can continue until the wheel accelerates again in the ABS control.

[0039] This allows for very small control deviations in wheel speed. The method described above can also be applied to the pressure build-up. The rates of pressure increase can be optimized according to control criteria. This achieves the goal of braking the wheel in the immediate vicinity of the maximum friction force, thus achieving optimal braking efficiency with optimal driving stability.

[0040] Special cases of the control were mentioned above, in which a throttle effect is advantageous. This is the case, for example, when a simultaneous pressure reduction is necessary for both wheels. Here, the throttle effect is advantageous until the actuating piston has provided such a large chamber volume that the subsequent rapid pressure reduction can take place 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 simultaneously on both wheel circuits. However, the individual alternating pressure build-up is to be preferred because of the metered pressure build-up with evaluation of the characteristic map and controlled adjustment speed of the piston. The same alternating procedure can be used as an alternative to the above with the throttle effect for pressure reduction. As a further option, 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 up and the MV switches to pressure holding (see brake manual p. 52-53). This signal is issued 5-10 ms before the signal to reduce pressure. The proposed fast drive is able to provide a chamber volume for a 10 bar pressure reduction within approx. 5 ms.

[0041] Based on the piston position for pressure reduction, the controller can decide whether sufficient chamber volume is available for simultaneous pressure reduction for both wheel brakes.

[0042] These explanations show that the concept with the fast and variably controlled electromotive piston drive and the solenoid valve with the evaluation of the pressure and characteristic map represents a high potential for the controller, which enables additional braking distances to be shortened and driving stability to be improved.

[0043] Fig. 2 shows the entire integrated unit for the BKV and control functions. The unit consists of two piston units with associated electric motors and gearboxes as shown in Fig. 1 for two brake circuits and four wheel brakes. The piston units are housed in the housing 4. This housing is attached to the front wall 29.

[0044] The brake pedal 30 transmits the pedal force and movement via the bearing pin 31 to a fork piece 32, which acts on the actuating device 33 via a ball joint. The actuating device has a cylindrical extension 34 with a rod 35.

[0045] Cylinder 34 and rod 35 are mounted in a bushing 37. This accommodates the travel simulator springs 36 and 36a, with one spring acting weakly and the other spring acting strongly progressively in terms of force increase. The travel simulator can also be constructed from even more springs or rubber elements. This travel simulator specifies the pedal force characteristic. The pedal travel is detected by a sensor 38, which in the illustrated example is constructed according to the eddy current principle, into which the rod 35 with a target enters.

[0046] The pedal movement is transmitted to the elements 32 and 33, the piston 34 moves with the rod 35 in the sleeve 37. A lever 26 is rotatably mounted on the actuating device and, in the event of a power failure, strikes the pistons. The pedal travel sensor supplies the displacement signal to the electronic control unit, which causes the pistons to move via the electric motor in accordance with the BKV characteristic, as described in Fig. 7. The parameters of this characteristic are described in more detail in Fig. 7. There is a play s between the lever 26 and the two pistons 1

[0047] The aim of the invention is a simple solution in which the path simulator is switched off when the power supply fails. For this purpose, a counterforce is exerted on the bushing 37 when the energy supply is intact via the armature lever 41 with a large transmission ratio and the holding magnets 42, which is eliminated when the electrical energy supply fails. Two-stage levers can also be used to reduce the magnet. This is described in detail in Fig. 3. In this case, the lever comes into contact with the two pistons via the brake pedal after the play has been run through, and can thus transfer the pedal force to the pistons. The pistons are dimensioned in such a way that, when the pedal is fully depressed, they generate a pressure that still results in a good braking effect, e.g. 80%. However, the piston stroke is considerably greater than the pedal stroke and can generate much higher braking pressures with an intact energy supply and electric drive. 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 the conventional design, in which the BKV and the main brake cylinder with piston are connected in series, the required pedal force increases by up to a factor of 5 for the same wheel brake pressure if the energy supply fails. With the new design, the factor can be reduced to 3, for example. This case is relevant, for example, when towing a vehicle with a flat battery.

[0048] The lever 26 is pivotally mounted so that it can take into account tolerances in the movement of the pistons, e.g. due to different ventilation. This compensation can also be limited so that the lever comes to rest against a stop 33a of the actuating device.

[0049] However, further fault cases must be considered.

Failure of an electric motor.

[0050] In this case, the amplification and control of the neighboring intact piston drive is fully effective. Brake pressure is generated in the failed circuit via lever 26 after it comes into contact with stop 33a. In addition, the booster characteristic of the second circuit can be increased, which reduces the required pedal force. However, this can also be done without a stop.

Failure of a brake circuit.

[0051] In this case, the piston moves to the stop in the housing 4 . The intact second circuit is fully effective. Unlike in conventional current systems, there is no pedal drop, which is known to irritate the driver. The irritation can also lead to a complete loss of the braking effect if the driver does not depress the pedal.

[0052] Fig. 3 describes the function of the travel simulator lock. In the limit case, the driver can apply high pedal forces, which the locking mechanism has to apply via the armature lever 41. To avoid that the magnet 42 with the exciter coil 43 has to apply these forces fully, the upper ball end 41a of the lever engages asymmetrically with the bushing 37. When the pedal is depressed until the rod 35 touches the base 37b, this lever action causes the bushing 37 to twist slightly, which generates friction in the guide, while the lug 37a can also be supported on the housing 4. This means that the magnetic force can be kept relatively low. The magnet is also designed as a holding magnet 42, so that a small holding power is necessary due to the small air gap. If the power supply fails, the armature lever 41 is deflected by the bushing 37 into the dashed-dotted position 41′. When the operating device 33 returns to its initial position, the return spring 44 returns the armature lever to its initial position.

[0053] The sensor 38 has been moved to the end of the bore of the sleeve in the housing 4, which has advantages for the contact with the electrical control unit, as shown in Fig. 6. The same applies to the brake light switch 46. In this embodiment, the target 45 for the eddy current sensor is drawn.

[0054] The locking of the travel simulator via the socket 37 can be changed in order to avoid the pedal feedback effect in ABS described in Fig. 7. To do this, the lever 41 with its bearing and magnet 42 with the receiver 42a can be moved via an electric motor 60, which drives a spindle 60a via a gear 60b. The lever is mounted on the extension of the spindle and the magnet housing is attached.

[0055] Fig. 4 shows a schematic representation of a solution with only one electric motor 7a. This description is based on Fig. 1 and Fig. 2. The motor’s drive pinion moves the rack 5c, which, similar to Fig. 1, can also be moved in parallel. This is connected to a piston 1a, which builds up pressure in the brake circuit 13a and at the same time displaces the piston 1a via the pressure, which builds up pressure in the brake circuit 13. This piston arrangement corresponds to a conventional master brake cylinder for whose pistons and seals there are many variants. As in the previous figures, the 2/2-way solenoid valves 14, 14a, 15′, 15a are arranged in the brake circuits. The ABS pressure modulation is carried out in the manner described above. The BKV function is carried out by a path simulation 36 and a path sensor 38 arranged in parallel. Here too, there is a play or free stroke between piston 1a and brake pedal s

[0056] Fig. 5 shows the view from the front wall onto the integrated unit, the flange 4b of which is screwed to the front wall by means of screws 47. The actuating unit 33, lever 26 and a bolt 39, which is not shown in the offset position, can be seen here as an anti-rotation device. The outline of a 10” vacuum-BKV is shown here for size comparison. This shows an important advantage in the overall height with the cover 48 of the storage tank. The front wall could be lowered according to distance A, which the designers want. On the left side of the flange, with reference to Fig. 5a, the drive of the rack 5 is shown with a dashed line. This detail is shown enlarged as Fig. 5a on the right half of the picture. The pinion of the gear wheel 6 engages on both sides in the H-shaped design of the rack 5 . The transverse forces described are supported by the roller 10 or 11 according to Fig. 1 with bearing 10a. For cost reasons, the rack can be made of plastic. Since its surface pressure is insufficient, hard sheet metal strips 49 are inserted here, which adapt to the rollers when the support is slightly convex. The gear wheel 7 is pressed into the pinion 6 and is in mesh with the motor pinion. The pinion is preferably mounted in the motor housing 8a.

[0057] Fig. 6 shows a side view of the integrated unit with housing 4, fork piece 32 for brake pedal 30, operating unit 33, flange 45, fastening screws 47, cover 48. This view shows the short overall length, with the electronic control unit 50 attached to the front. This is connected to the coils or part of the magnetic circuit of the solenoid valves 14 and 16 using state-of-the-art technology, in order to save additional contacting and electrical connection lines. This feature can be extended by directly connecting all electrical components such as the electric motor 8 , solenoid 43 , position sensor 38 , brake light switch 46 , brake fluid level sensor 53 to the control unit without electrical connecting lines. In this case, the control unit would have to be installed from above in the direction of 50a. However, it is also possible in the direction of 50b, which results in a modified arrangement of the solenoid coil.

[0058] The solenoid valves are preferably mounted on a carrier plate 51, since these are pressed in for cost reasons in aluminum with high elongation at break. The screw plugs 52 for the brake lines are screwed into this carrier plate. The contact is drawn in the middle part of the control unit, which contains a redundant power supply in the area 54, the bus line in the area 55, and the sensors for ABS and ESP at 56.

[0059] Fig. 7 shows the essential characteristic curves of the braking system. The pedal force F

[0060] The characteristic curves 59 show the failure of the electric drive, in which, according to the play s

[0061] From the pedal position and the brake pressure, it can be seen that the pressure modulation of 10 bar at blocking pressures > 50 bar does not affect the pedal, since the pedal at S

[0062] The thicker lines are the amplifier lines 58 and 58a, which show the assignment of pedal force F

[0063] At F

[0064] the electric drive can be regarded as more fail-safe than the vacuum BKV in the event of a power failure, since at least two electric motor drives are used for the proposed invention, i.e. one acts redundantly and, as is well known, as the total failure rate λ

[0065] Fig. 8 shows a further solution for the piston drive. Instead of the rack, a crank rocker 60 can be used, which is connected to the piston via a tie rod 61 and the bearing bolt 62. The return spring 9 acts on the crank rocker, whose initial position is given by the stop 65. The crank rocker is driven by the motor 11 via a multi-stage gear 63.

[0066] Fig. 8a shows a two-arm crank rocker 60 and 60a with two tension struts 61 and 61a. This means that only slight transverse forces act on the piston. The gearbox 63 is encapsulated here in an extended motor housing 64 and is driven by the drive pinion 11a of the motor 11. The advantage of this solution lies in the encapsulation of the gearbox, which allows for oil or grease filling, permits helical gearing and is therefore more resilient and quieter.

[0067] Fig. 9 shows a further alternative with a spindle drive, which is arranged within the rotor of the electric motor. This arrangement is known from DE 195 11 287 B4, which relates to an electromechanically actuated disc brake. 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 pressure forces of the piston 1 act on this. The spindle drive also acts as a reduction gear, with the spindle 65 transmitting the force to the piston. All the drives shown so far have a reduction gear that is firmly coupled to the piston, which has to be moved from the brake pedal if the power supply fails and accelerated with the engine if the pedal is pressed quickly. These mass inertia forces prevent rapid operation of the pedal and irritate the driver. To avoid this, the nut in the rotor bore is axially movable so that the ball screw is switched off when the pedal is engaged. The nut is fixed for normal operation with an electric motor by a lever 70 mm, which is effective when the piston is quickly pushed back, especially when there is a vacuum in the piston chamber. This lever is mounted in the rotor via the shaft 71 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 quickly, the centrifugal force acts on the lever, and the nut is enclosed by the lever for the movement of the piston.

[0068] This movement can also be accomplished by an electromagnet drawn with a dashed line, in which the lever represents a rotating armature. The twisting moment generated by the nut on the spindle is absorbed by two bearing pins 69 and 69a. These pins also support the return spring 9. 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 larger 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 materials of construction of the motor and piston housing.

[0069] The nut 67 can also be connected directly to the rotor 66, e.g. by injection molding. A plastic nut with a low coefficient of friction can be used for the required forces.

[0070] If a motor or the power supply fails, the non-drawn pedal acts on the fork according to Fig. 2 and via the lever 26 after the free travel so on the spindle 65 or piston 1. Since a blocking of the drive can be eliminated with this solution, the stop 33 can have a smaller distance to the lever. This has the advantage that the pedal force acts fully on the piston if, for example, an electric motor fails. As soon as the lever is supported on the opposite end when it is twisted, only half the pedal force acts on the piston. In the design, the spindle and piston are decoupled, which was not designed separately.

[0071] The return of the piston to the initial 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 them. This is intended to compensate for the detent and friction torque of the motor. This is particularly advantageous for small resetting forces of the pistons, which act on the pedal in the event of a power failure, in conjunction with the coupling lever described in Fig. 9.

[0072] Fig. 10 shows a further simplified design with an electromotive piston drive, in which the piston 1 again performs the brake booster and pressure modulation for ABS. The piston chambers 1′ are connected to the wheel brakes (not shown) and to the solenoid valves (also not shown) via lines 13 and 13a, as shown in Fig. 1 to Fig. 9. The design corresponds to Fig. 8 with spindle drive 65 and with rotor 66, fixed nut 67, separation of motor and piston, housing 74 or 4, piston return springs 9 and bearing pin 69, spiral spring 66a for motor return. The pedal force is transmitted similarly to Fig. 2 from a fork piece 26 to an actuating device 34 with a 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. In turn, a lever 26 is mounted on the actuating device 35, which preferably carries leaf springs 76 at the end in the connection to the piston, which are connected to a travel sensor 77 in the case of a strong leaf spring or to a force sensor 77a in the case of a softer spring. In both cases, the force transmitted by the lever or pedal should be measured here. The leaf spring 76 has the task of avoiding a hard reaction when the pedal is pressed before the engine starts. The function is carried out in such a way that, at a certain function of this pedal force, the engines exert an amplifying force on the piston, whereby this force can in turn be determined from the current and piston travel or a pressure sensor. In this case, the pedal travel can be processed via the travel sensor 38 in this amplifier function or characteristic. 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.

[0073] 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 travel simulator and locking device. A disadvantage is the limited pedal travel characteristic of the amplifier characteristic, a drop in the pedal when the brake circuit fails and higher pedal forces when the amplification fails, since pedal travel and piston travel are identical. This design is mainly suitable for small vehicles.

[0074] In the embodiment according to Fig. 10, safety valves 80 are shown as representative of all solutions, which become effective if, for example, 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. This ensures that when the pedal is in the initial position, no brake pressure is built up in the brake circuit. These valves can also be actuated electromagnetically.

[0075] Safety-related systems usually have a separate shutdown option for faults in the output stages, e.g. full current flow through conduction. In this case, a shutdown option is provided, e.g. by means of a conventional relay. The diagnostic part of the electrical circuit detects this fault and switches off the relay that normally supplies the power stages with power. The concepts proposed here must also include a shutdown option, which is implemented by a relay or a central MOSFET.

[0076] In view of the pulse control of the electric motors, a fuse can also be used, since the pulse-off ratio is very high.