Brake system, having an actuating device and a control and regulating device, wherein the control and regulating device controls at least one electromotive drive device on the basis of the movement and/or position of the actuating device in such a way that a pressure is set in a working chamber of the piston-cylinder system, the working space being connected to at least one wheel brake via a pressure line and at least one valve being arranged between a brake cylinder of the at least one wheel brake and the working space of the piston-cylinder system, the control and regulating device opening the at least one valve to build up pressure in the brake cylinder and closing it to maintain the pressure in the brake cylinder, wherein, during the control by means of the piston-cylinder system, an initial pressure is regulated, wherein the control and regulating device is designed to set a first initial pressure (91) when the coefficient of friction is high and to set a second initial pressure (91a) when the coefficient of friction is low, wherein the first initial pressure (91) is significantly higher than the second initial pressure (91a).

[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 device on the basis of the movement and/or position of the actuating device, the drive device displacing 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.

Prior art

[0002] Modern brake systems consist of a brake booster, i.e. the conversion of the pedal force into a corresponding amplified 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 as the means of transmitting the brake pressure from the pedal force.

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

[0004] The hydraulic unit (HU) consists of solenoid valves, multi-piston pumps for dual-circuit braking 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 used with 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 in gasoline engines, the internal combustion engine is used, 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 system and Bosch used the ABS3 system, which were integrated units containing all components for brake booster and control with a hydraulic BKV, see Bosch Automotive Engineering Manual 1986, 20th edition. These systems did not catch on for cost reasons, except for use in armored vehicles. The same applies to fully electric braking systems, known as 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 power supply is also necessary for this. The design of the wedge brake, which requires additional rollers for hysteresis reasons, which in turn require integration into the brake caliper, has not yet been solved. 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 very fine metering of the brake pressure, which also allows variable brake force coordination EBV. The pressure reduction speed is not yet optimal because it is highly non-linear. Furthermore, in the case of a µ-jump or a small friction coefficient, the pressure reduction rate is determined by the relatively low pump output, which leads to large control deviations and thus to a loss of braking distance.

[0008] A generic braking system is known from DE 33 42 552 A1. 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 for motor vehicles with a hydraulic braking pressure generator that can be driven by an electric motor is known from DE 44 45 975 A1. A control scheme or control program is known from DE 44 10 299 A1, by means of which, in driving situations with a low coefficient of friction, when the initial pressure is low and wheel instability occurs at the same time, maintaining the cornering forces is given priority over achieving a short braking distance.

[0010] A circuit arrangement for adapting the control of an anti-lock braking system for road vehicles, by means of which the brake pressure control valves can be controlled, is known from DE 35 00 745 C2. These are inserted into the pressure medium paths.

Aim of the invention

[0011] The present invention has the purpose of providing a novel brake system that is small and compact in its dimensions.

[0012] This problem 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.

[0013] The advantage of the invention braking system is that it realizes the brake booster 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 brake pressure build-up and brake pressure reduction, for ABS and anti-skid control, as well as in the event of a 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, 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.

[0014] The advantageous provision of a sensor system and a displacement simulator allows variable pedal characteristics such as a brake-by-wire function, i.e. brake pressure increase independently of pedal actuation, to be freely variable, also taking into account the braking effect of the generator in the case of recuperative brakes.

[0015] 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 when the amplification is either intact or has failed. 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 required, including the electrical connections, also results in a reduction of the failure rate.

[0016] The electric motor drive also makes it possible to improve 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.

[0017] The brake system according to the invention requires considerably less energy due to the advantageous provision of a 2/2-way valve for brake force boosting, maintaining the regulated braking pressure and the control function.

[0018] The valves, together with the hydraulic lines, are to be designed with the smallest possible flow resistance, i.e. large flow cross sections, so that a pressure build-up or pressure reduction that is as fast and variable as possible can be realized by means of the piston-cylinder system(s), since then in particular the valves and also the connecting channels and piping no longer have a throttling effect. This ensures that the piston-cylinder system alone determines the rate of pressure build-up or pressure reduction.

[0019] It is also possible to provide a separate piston-cylinder system with an associated 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.

[0020] 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 diagram represents a high potential for the controller, which enables additional reductions in braking distance and driving stability. It is advantageous to use pressure-compensated seat valves or slide valves with low temperature dependence and short switching times, so that shorter dead times can be achieved and thus short cycle times can be achieved.

[0021] The brake system is also to be controlled by a control-related process in such a way that the switchover times are as short as possible, i.e. it must be possible to switch over to another control channel or another wheel brake as quickly as possible in order to control it. In this context, it has proven advantageous to evaluate the signals for the next wheel brake to be adjusted while the pressure for one wheel brake is being adjusted, so that it is possible to switch immediately to the other wheel brake after the adjustment process for the first wheel brake has been completed.

[0022] The invention shows a special method for controlling constant and variable gradients by evaluating the pressure volume characteristic of the corresponding brake via piston displacement, current or pressure.

[0023] Furthermore, it has been found to be advantageous to provide larger brake pipe diameters and heatable brake pipes.

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

[0025] The figures show:

Fig. 1: a first embodiment of a braking system with one braking 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 third embodiment of a brake system, wherein the piston-cylinder system has a cylinder and two pistons;

Fig. 5: principle structure of the braking system according to the invention;

Fig. 6: pressure curve during pressure reduction from a level P₀, which corresponds, for example, to the blocking limit of a dry road, for conventional and braking systems according to the invention;

Fig. 7: pressure reduction and pressure build-up at high and low µ for conventional braking system

Fig. 7a: pressure reduction and pressure build-up at high and low µ for an invention-based braking system

Fig. 8: time curve of wheel speed and pressure for a conventional and an invention-based braking system;

Fig. 9 to Fig. 10a: pressure curves and valve positions during pressure reduction;

Fig. 11: time curve of several control cycles.

[0026] Fig. 1 shows a section of the integrated unit responsible for pressure generation and brake boost. In this process, the piston 1 with the usual seals 2 and 3 is moved in the cylinder housing 4 parallel to the piston via a specially designed rack 5a. The seal 2 is designed in such a way that it also seals in the piston chamber 4′ when there is a vacuum. This rack 5a transfers the force to the front ball end of the piston 1.

[0027] This has a collar bolt 1a at this point, via which the rack 5a with return spring 9 brings the piston into the initial position. Here the rack rests against the cylinder housing 4a. This external spring has the advantage that the cylinder is short and has little dead space, which is advantageous for venting. Due to the transverse forces, the rack is supported by the rollers 10 and 11 with the sliding block 12. Fig. 1 clearly shows that the rack’s parallel alignment to the piston results in a short overall length. The unit must be very short in order to be outside the crash zone. To achieve this, the rack must be designed with a very rigid H-profile. The rollers are arranged so that the rack has a relatively small bending length in the end position 5b (shown by the dashed line) 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 7 via the toothed profile 5a and gear wheel 6. 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) 2 2. 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 map must be created in the vehicle when it is put into operation and also during operation, in which the various flow rates are assigned to the position of the piston. During operation, the booster characteristic lines described later are used to move the piston to a position that corresponds to a certain pressure according to the map. If the position and motor torque do not quite match, e.g. due to temperature influences, the map is adapted during operation. This means that the map is continuously adapted. The initial characteristic diagram is formed from the pressure-volume characteristic of the wheel brake, the engine characteristic value, the transmission efficiency and the 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.

[0028] Piston 1 generates a corresponding pressure in line 13, which is transmitted to wheel brake 15 via 2/2-way solenoid valve (MV) 14 and to wheel brake 17 via 2/2-way solenoid valve (MV) 16. The arrangement described above 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.

[0029] It is sufficient to use a piston-motor unit for each brake circuit.

[0030] The second advantage is the very low energy requirement and also the design of the motor only for impulse 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 with a low current until a new setpoint value is specified by the brake pedal. This makes the energy requirement and the average power extremely small. For example, in a conventional design, the motor would consume a high current of 3 A during an emergency stop from 100 km/h. According to the invention, the motor requires only approx. 0.05 A 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 3s 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 pulse 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.

[0031] 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.

[0032] The piston 1 is supplied with fluid from the reservoir 18 via the bleed hole. A solenoid valve 19 is connected in this line. If the piston moves quickly to reduce the pressure, the seal 3 could bleed fluid from 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 vapour locks occur, where 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 dispensed with, a snifting-proof seal 3 can be used instead of the solenoid valve 19.

[0033] The solenoid valves 14, 16, 19 are controlled by the microcontroller 22 via output stages 28.

[0034] 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 certain amount of play is built into this lever and the piston to prevent the lever from hitting the piston before the motor moves it when the pedal is depressed quickly.

[0035] 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.

[0036] The essential functions of the new system should be shown in a table:

Pressure Pressure

Functions Electric motor Wheel brake 15 Solenoid valve 14 Wheel brake 17 Solenoid valve 16

BKV On Structure 0 Structure 0

Partially energized P = constant 1 P = constant 1

Partially energized Reduction 0 Reduction 0

Brake control On Build-up 0 Build-up 0

partially energized P = constant 1 P = constant 0

On Build-up 0 P = constant 1

partially energized Reduction 0 P = constant 1

partially energized Reduction 0 Reduction 0

[0037] The level of the partial flow 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.

[0038] 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, at a small differential pressure at the outlet valve that determines the pressure drop, the flow rate and thus the rate of pressure drop are low. As already mentioned, the piston unit can be used separately for each wheel with and 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 one 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: when there is a pressure drop at 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 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 e.g. 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.

[0039] The concept also offers new possibilities for pressure reduction in terms of control engineering. In terms of control engineering, the pressure reduction and braking torque reduction are essentially proportional to the rotational acceleration of the wheel, the hysteresis of the seal and inversely proportional to the moment of inertia of the wheel. 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.

[0040] 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 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.

[0041] 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 ensures that the wheel is braked in the immediate vicinity of the maximum friction force, thus achieving optimum braking efficiency with optimum driving stability.

[0042] 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 in 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 and the MV switches to pressure holding (see brake manual p. 52-53). This signal is issued 5-10 ms before the signal for pressure reduction. The proposed fast drive is able to provide a chamber volume for a 10 bar pressure reduction within approx. 5 ms.

[0043] 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.

[0044] 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.

[0045] 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. This has a cylindrical extension 34 with a rod 35.

[0046] Cylinder 34 and rod 35 are mounted in a sleeve 37. This sleeve holds the travel simulator springs 36 and 36a, 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 determines the pedal force characteristic. The pedal travel is detected by a sensor 38, which in the example shown is constructed according to the eddy current principle, into which the rod 35 with a target is immersed.

[0047] 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 operating device and comes into contact with the piston in the event 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 in accordance with the BKV characteristic. A play s_o is provided between the lever 26 and the two pistons 1, as shown in Fig. 1. The actuating device has an anti-twist device via the bolt 39, which is shown offset, and a return spring 40, which supports the pedal return spring, which is not shown. According to the state of the art, many travel simulator solutions are known which are also partially hydraulically actuated by pistons and shut off by solenoid valves when the energy supply fails. This solution is complex and subject to hysteresis. Solutions are also known in which the travel simulator travel is lost when the energy supply fails when the pistons are actuated to generate braking pressure.

[0048] The aim of the invention is to provide a simple solution in which the path simulator is switched off when the power supply fails. To this end, when the power supply is intact, a counterforce is exerted on the bushing 37 via the armature lever 41 with a high transmission ratio and the holding magnet 42, which is eliminated when the electrical power supply fails. Two-stage levers can also be used to reduce the magnet. This is described in detail in Fig. 3. In this case, the lever comes into contact with the two pistons via the brake pedal after the play has been overcome and can thus transmit the pedal force to the pistons. The pistons are dimensioned so that at full pedal travel they generate a pressure that still results in a good braking effect, e.g. 80%. However, the piston travel is considerably greater than the pedal travel and, with an intact energy supply and electric drive, can generate much higher braking pressures. 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, for example, the factor can be reduced to 3. This case is relevant, for example, when towing a vehicle with a flat battery.

[0049] The lever 26 is rotatably 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 on a stop 33a of the actuating device.

[0050] However, further fault cases must be considered.

Failure of an electric motor.

[0051] 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 has reached the 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 one brake circuit.

[0052] In this case, the piston moves to the stop in the housing 4. The intact second circuit is fully effective. Unlike in conventional systems today, there is no “falling” pedal, 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.

[0053] 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 the magnet 42 with the exciter coil 43 having to apply these forces fully, the upper spherical 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 sleeve 37 into the dashed-dotted position 41′. When the actuating device 33 returns to its initial position, the return spring 44 returns the armature lever to its initial position.

[0054] The sensor 38 has been moved to the end of the bore of the sleeve in the housing 4, which has advantages for the electrical connection to the 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 shown.

[0055] The locking of the travel simulator via the socket 37 can be changed to avoid the pedal feedback 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.

[0056] 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 figures above, the 2/2-way solenoid valves 14, 14a, 16, 16a are arranged in the brake circuits. The ABS pressure modulation is carried out in the manner described above. The brake balance function is performed by a parallel path simulation 36 and a path sensor 38. Here too, a play or empty stroke s₀ is provided between piston 1a and the brake pedal. The brake fluid passes from the reservoir 18, 18a into the piston chambers. This arrangement is cost-effective. The dynamics of the BKV function in pressure build-up are lower than in the variant with two motors, since the electric motor has to provide twice the torque. Furthermore, the redundancy function of the 2nd motor, as described in Fig. 7, is omitted, including a pedal failure in the event of a brake circuit failure.

[0057] Fig. 5 shows the pressure modulation device described in Figs. 1 and Fig. 2, which contains an electric motor 8 that is controlled via the shunt 23 for pressure-proportional current measurement via output stages 21. The latter are shown in a simplified form. The piston displacement is detected by a rotary encoder 72 or a piston displacement sensor 74, which is also used for controlling an EC motor. This motor actuates the piston, which moves the pressure medium to the corresponding wheel brakes via the 2/2-way solenoid valves 14, 14a. The corresponding brake fluid reservoir 18 is connected to the piston housing. A cost-effective central actuating device can also be used for four wheel brakes and additional 2/2-way solenoid valves 14‘ and 14a‘. A piston displacement sensor 74 or a displacement or angle of rotation sensor together with pressure transmitters 73 and 73a in the wheel circuits can be used to control the piston.

[0058] In ABS, EHB and ESP systems, the solenoid valves for pressure control and for pressure build-up and reduction are designed as throttle valves (ATZ Automobiltechnische Zeitung 101 (1999) 4, p. 224). In principle, the aim is to make the pressure build-up and pressure reduction gradients as steep as possible so that the braking torque surplus is quickly equalized during control. However, the solenoid valves used in the prior art have dead times, which means that an additional pressure change occurs after the control command – e.g. closing. As a rule, this is approximately 3 bar if the gradient is 1500 bar/s and the switching time is 2 ms. This closing process also causes, among other things, pressure oscillations, which affect the wheel behavior and, among other things, cause noise. This means that the solenoid valves with their switching characteristic determine the maximum gradient for pressure build-up or pressure relief. Due to the fixed throttle resistance of the valves used, the pressure build-up and pressure relief gradient is highly non-linear and approximately follows the function where Δ_P is the differential pressure. However, a variable and constant pressure gradient is advantageous for optimal and simple control.

[0059] It is essential for the invention that the design and dimensioning of the 2/2 solenoid valves is such that they have almost no throttling effect, so that the actuating device determines the pressure gradient. Preferably, pressure-relieved seat valves with low temperature dependence are used.

[0060] Knowledge of the pressure-volume characteristic of the wheel brake, as shown in Fig. 5a, is important for controlling the gradient. In the upper part, the dependence of pressure (current) on volume absorption, which is proportional to the piston travel or angle of rotation α, is shown. This is known to be non-linear. For constant pressure gradient control, the pressure-volume characteristic must be evaluated for corresponding speed control of the piston.

[0061] For the method in which a central actuating device serves several control channels, it is very important to keep the dwell time on a control channel as short as possible, since the other control channels are not served during this time. In this case, a fast pressure gradient, especially during pressure reduction, and a short switching time of the 2/2-way solenoid valves are of great importance. This is described in more detail in the following figures.

[0062] Fig. 6 describes the pressure reduction from a level P₀, which corresponds, for example, to the blocking limit on a dry road. In the event of a µ-step on ice or aquaplaning, the pressure level must be reduced to the level of line 89. In the systems mentioned at the beginning, the pressure reduction according to line 86 is not linear with very small gradients at a low pressure level. In systems with a state-of-the-art storage chamber, this fills up at 88. The very slow pressure curve determined by the performance of the return pump is shown by the dashed line. By contrast, the system according to the invention proposal causes an almost constant gradient – line 87, which can be selected larger or higher than in conventional systems (line 86) due to the design. A transition region is present only in the lower course at very low pressures 80, due to the positioning speed of the piston. The decisive factor for this favorable pressure reduction is the dimensioning of the solenoid valves and pipelines, which should not form any significant flow resistance for the corresponding pressure gradients even at low temperatures, so that only the adjustment speed of the piston is dominant. A larger diameter can be used for the brake pipe, or alternatively the brake pipe could be heated electrically.

[0063] Fig. 7 shows the pressure drop and pressure build-up on the left side at high and on the right side at low µ. The dashed line is intended to represent the so-called

[0064] correspond to the initial pressure 91 that the driver generates in the master cylinder. As already explained, the pressure reduction gradient p_ab/dt depends on the pressure level and the build-up gradient p_au/dt on the differential pressure with respect to the initial pressure. In particular, when controlling at a low pressure level, high differential pressures arise and thus high pau/dt. The valves are clocked to a stepped pressure build-up. This generates pressure oscillations 92a and 93a due to the fast closing of the solenoid valve, which cause considerable noise and even affect the wheel behavior.

[0065] Fig. 7a shows the pressure-time behavior at high and low µ in the new system. The pressure gradients p_ab/dt and p_au/dt can be the same regardless of the pressure level. The pressure build-up gradient p_au/dt can be different within the control cycle, e.g. large during the first pressure build-up p_au1 and smaller during the second pressure build-up p_au2.

[0066] The variable pressure gradients can be used to create a transition region 94 and 94a in the pressure reduction and pressure build-up, which avoids pressure oscillations. The pre-pressure of the system can also be controlled by corresponding control of the actuating element such that the pre-pressure is 20% higher than the maximum regulated pressure. This saves a corresponding amount of electrical energy for controlling the actuating element.

[0067] Fig. 8 shows the variation with time of wheel speed and pressure. The curves are highly linearized. During the braking process, the wheel speed increases up to point 95, at which the locking limit is exceeded, which manifests itself in the fact that the wheel acceleration increases. Before the pressure reduction begins, a differential speed ΔV₀ is waited for. It makes sense to keep the pressure constant during this phase. At time 96, the pressure reduction occurs according to curve 101 in the conventional system. This occurs after a dead time of t_VA. Here the small gradient is shown as a small µ. At time 102 the torque surplus, which causes the blocking tendency, is equalized by the corresponding pressure reduction. The wheel speed V_R1,2 increases again, so that the blocking tendency disappears. For the sake of simplicity, it is assumed that both wheel speeds are synchronous and controlled simultaneously when considering the conventional system. In this case, the so-called control deviation ΔV₃ occurs in the conventional system.

[0068] In the system according to the invention, the advantageously rapid pressure reduction also sets in at time 96 after t_VA, which is terminated at 97 and a much smaller control deviation ΔV₁ arises, after which the first wheel no longer locks (v_r1‘ increases again). Now the system switches to the second control channel, which leads to a pressure reduction after t_VA, which is completed at 98. This results in a control deviation ΔV₂, which is still smaller than in the conventional system with ΔV₃ despite the offset control.

[0069] In the first control cycle, the pressure reduction can also occur simultaneously in both wheels in the new system if both wheels become unstable and exceed points 95/96 because the initial pressure level is the same. This is of great importance, since when braking at a high pressure increase rate, the torque surplus is greater than in the following control cycles, in which the mean pressure increase is considerably smaller due to the stepped pressure build-up.

[0070] As shown, larger control deviations arise in the conventional system in the first control cycle by a factor of 2 to 3, which, as is well known, means longer braking distances and loss of lateral force.

[0071] The above descriptions show that, in the case of simultaneous instability and time-delayed pressure reduction, it is important to keep this time delay as short as possible. It should be noted, however, that this case rarely occurs in practice.

[0072] Fig. 9 and Fig. 10 show the main parameters influencing the time lag. The pressure curve over time is shown in a linearized form.

The abbreviations mean:

t_VA: delay or dead time of the control device or actuator

t_VM: delay time of 2/2 solenoid valve

t_C: Sampling time or sampling rate of the computer; the computer requires this time to calculate the speed when switching from one wheel to the next

t_ab: Pressure drop time

ΔT: Time offset

[0073] In Fig. 9, the controller issues the command to reduce the pressure at 103, marked with a triangle, which occurs after t_VA and is completed after tab. In this phase, the second control channel (dashed line) is kept at a constant pressure by closing the 2/2-way solenoid valves. After tab at 104, tc acts simultaneously 2 × t_VM. After t_VA or parallel opening of the 2/2-way solenoid valves via t_VM, the next p_ab occurs and after renewed t_C the next pressure change, which can cause pressure increase or decrease. It should be noted that the pressure build-up is less critical because the time factor 10 – 20 is greater than tab in the control cycle, since many phases of constant pressure, see Fig. 7a, are switched on. Fig. 9 shows a quantitative analysis of Δt = 17 ms as a time offset.

[0074] Fig. 9a shows one way of reducing Δt. At 103 the control command is given again. In this case, the necessary pressure reduction is calculated during t_VA, mainly from wheel acceleration and wheel moment of inertia, so that after t_VA the computer is switched to the next control channel, so that after tab and t_VA or t_VM, Δt has already been reached for the next pressure reduction. As shown on the left, Δt is reduced from 17 to 12 ms – 40%.

[0075] Fig. 10 and Fig. 10a correspond to Figs. 9 and Fig. 9a, respectively, with the difference that tab is chosen to be smaller by a factor of 2, which means that with the method according to Fig. 9a, Δt can be reduced from 17ms to 7ms. This is such a small value that the time offset has a negligible effect on the control deviation and makes it possible to operate four control channels with one actuator. Further potential can be utilized in the reduction of delay or dead times tVA and tVM.

[0076] As shown, t_VA and the rate of pressure reduction largely determine the switchover time Δt, i.e. t_VA should be small and the rate of pressure reduction as large as possible.

[0077] The dead time of the 2/2 solenoid valves can vary within certain limits, since a small switching delay in the pressure reduction is not noticeable because the piston is already being moved by the control signal. As soon as the solenoid valve opens, the fluid flows into the piston chamber with practically no throttling. The end of the pressure reduction can be seen in the performance diagram and can be taken into account with the corresponding advance. During the pressure build-up, the EC motor is activated slightly earlier than the expected opening time of the solenoid valve. The motor’s start-up indicates the actual opening time, since the pressure medium only reaches the brakes and the piston can move when the valve is open. If necessary, the activation time must be corrected. The closing time can be checked in a similar way.

[0078] The time of the end of the pressure build-up is known from the control algorithm and the characteristic diagram. If the intended pressure build-up is not achieved, the solenoid valve closes too early and its control time is corrected for later closing. The motor/piston remains for a short time after reaching the pressure build-up to be sure that the solenoid valve is closed.

[0079] Fig. 11 describes the timing of several control cycles. It shows the speed curve of two wheels, V_R1 and V_R2, with the associated pressure curve, p₁ and p₂. The pressure increase is shown. As is well known, a differential speed of V_R1 and V_R2 with respect to the vehicle speed, V_F, is formed here. This is referred to as slip. The so-called speed for optimum friction V_opt is also shown, which usually has a slip of 10%, but can also fluctuate between 5% and 30%, for example. This means that V_opt is usually 90% of V_F, for example. After the pressure increase, V_opt is exceeded at 105, and after Δ_v has elapsed (see Fig. 8), pressure is released at both wheels in the first control cycle because both have the same initial pressure level and exceed V_opt. According to the method shown in Fig. 9a or Fig. 10a, a pressure reduction is initiated here, preferably in proportion to the wheel acceleration and the moment of inertia, which is different for V_R1 and V_R2.

[0080] This calculation process runs independently of the calculation of the wheel speed or acceleration. The data for the pressure reduction can be stored, for example, in a characteristic diagram, so that no significant computing power/time is required. At time 107 the pressure reduction is completed at V_R1 and at 109 at V_R2. This is determined in each case so that the friction torque on the wheel is greater than the braking torque, resulting in wheel re-acceleration. At time 109, V_R1 exceeds Vopt and V_R2 exceeds 100. Here, a pressure build-up occurs whose amount is again proportional to the wheel acceleration and wheel moment of inertia, somewhat reduced compared to the pressure reduction, e.g. 90%.

[0081] After phases of pressure stabilization over, for example, 30 ms, a small pressure build-up of a few bar occurs at 101. However, this can be set higher if the wheel is at low slip values.

[0082] At time 102, the next pressure reduction to 113 occurs at V_R1. At 114, the pressure build-up occurs identically to 109 and 100, and at 115 and 116, the pressure reduction occurs at V_R2. At time 119, the larger pressure build-up corresponding to 100 and the smaller pressure build-up corresponding to 111 coincide. The large pressure build-up has priority here, the small one occurs offset by tv. At V_R2, at time 117, a large wheel acceleration is already evident at the point of greater slip. This results in pressure building up according to the conditions of 100. The same occurs again at 119.

[0083] In this representation, the delay times t_VA and t_VM and calculation time t_c were not taken into account in favor of a clear representation.

[0084] In the context of this invention, the term “control cycle” refers to the control process that initiates the pressure reduction after the speed falls below the speed for optimum friction or exceeds a corresponding slip value, see points 105 and 106 of Fig. 11. The end of the “control cycle” is given, see point 109 or 110 of Fig. 11, when the speed for optimum friction is exceeded or the slip value is again undershot. A “control cycle” for a wheel brake thus always consists of a phase in which pressure is built up or released and a subsequent phase in which the pressure is kept constant.

[0085] The following are examples of embodiments according to the invention.

Example 1:

[0086] 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 at least one electromotive drive device on the basis of the movement and/or position of the actuating device, the drive device displacing a piston of a piston-cylinder system via a non-hydraulic transmission device, so that a pressure is set in the working space of the cylinder, the working space being connected to a wheel brake via a pressure line, characterized in that a valve is arranged between the brake cylinder of the wheel brake and the working space of the piston-cylinder system, wherein the control and regulating device opens the valve to reduce or build up pressure in the brake cylinder and closes it to hold the pressure in the brake cylinder.

Embodiment 2:

[0087] Brake system according to embodiment 1, characterized in that the drive apparatus has an electromotive or electromechanical drive for adjusting the piston of the piston-cylinder system.

Embodiment 3:

[0088] Brake system according to embodiment 1 or 2, characterized in that the drive device drives a piston which, together with a hydraulically coupled further piston, is arranged in a cylinder (tandem piston-cylinder system).

Example 4:

[0089] Brake system according to one of the embodiments 1 to 3, characterized in that the brake system has two piston-cylinder systems arranged parallel to one another, and each piston is assigned a drive device which adjusts the respectively associated piston.

Embodiment 5:

[0090] Brake system according to one of the previous embodiments, characterized in that a valve, in particular a 2/2-way valve, is arranged between each wheel brake and the working space of the piston-cylinder system.

Embodiment 6:

[0091] A brake system according to one of the preceding embodiments, characterized in that the hydraulic lines connecting the working chamber of the piston-cylinder system to the brake cylinder have negligible flow resistance.

Example 7:

[0092] Brake system according to example 6, characterized in that the valve has a large flow cross section, such that the valve has no throttling function, the valve being in particular a 2/2 slide valve.

Example 8:

[0093] Brake system according to embodiment 6, characterized in that the valve is a pressure-compensated 2/2-way seat valve.

Embodiment 9:

[0094] A brake system according to one of the preceding embodiments, characterized in that the actuating device, in the event of failure, adjusts the at least one piston of the at least one piston-cylinder system directly or via a gear mechanism.

Embodiment 10:

[0095] Brake system according to one of the preceding embodiments, characterized in that the pressure in the working space of the piston-cylinder system and/or the brake cylinders of the wheel brakes is determined by means of sensors.

Embodiment 11:

[0096] Brake system according to one of the previous embodiments, characterized in that at least sections of the hydraulic lines connecting the wheel brakes to the piston-cylinder system can be heated by means of heating devices, in particular electric heating elements.

Embodiment 12:

[0097] Brake system according to one of the previous embodiments, characterized in that the control and regulating device has a knowledge database, in particular in the form of a characteristic diagram, which is in particular designed to be adaptive.

Embodiment 13:

[0098] Method for setting a pressure in at least one brake cylinder of a brake system according to one of the preceding embodiments, characterized in that the pressure in one or more brake cylinders is adjusted simultaneously or in succession by means of the at least one piston-cylinder system and the valves associated with the wheel brakes.

Embodiment 14:

[0099] Method according to embodiment 13, characterized in that the rate of change of the pressure build-up and/or the pressure reduction in the wheel brakes is regulated as a function of the driving state or the brake control of the vehicle or the respective wheel to be braked by means of the piston-cylinder system.

Embodiment 15:

[0100] Method according to Embodiment 13 or 14, characterized in that the rate of change of the pressure build-up and/or the pressure reduction in a wheel brake changes during a control cycle.

Embodiment 16:

[0101] Method according to embodiment 15, characterized in that the rate of change of the pressure reduction and/or pressure build-up during the time in which a valve of a wheel brake is open changes, in particular is initially high and is reduced towards the end of the pressure reduction or pressure build-up phase.

Embodiment 17:

[0102] Method according to one of the embodiments, characterized in that the control and regulating device determines the required pressure build-up, pressure reduction, pressure holding phases and/or the optimum slip for the respective wheel or all braked vehicle wheels, at least from the respective wheel speed, the vehicle acceleration and the pressure present in the respective brake cylinder of the wheel brake.

Embodiment 18:

[0103] Method according to one of the embodiments, characterized in that the control and regulating device, during the pressure build-up or pressure reduction for a first wheel brake, the control and regulating device opens the valve associated with the wheel brake, and immediately after setting the pressure determined by the controller for the first brake, the control and regulating device closes the valve associated with the first wheel brake and, by opening the valve for the second wheel brake, regulates the necessary pressure for the second wheel brake by means of the piston-cylinder system.

Embodiment 19:

[0104] Method according to embodiment 18, characterized in that the necessary pressure reduction or pressure build-up for the wheel to be adjusted next is calculated in particular on the basis of the measured wheel acceleration and the wheel moment of inertia, in particular from the characteristic diagram.

Embodiment 20:

[0105] Method according to embodiment 19, characterized in that the pressure reduction or pressure build-up to be newly set for the second wheel brake is calculated during the adjustment of the pressure for the first wheel brake.

Embodiment 21:

[0106] Method according to one of the embodiments 13 to 20, characterized in that the pressure reduction in two wheel brakes is carried out by opening the respectively associated valves at the same time, in particular when approximately the same pressure level initially prevails in the brake cylinders of the two wheel brakes or in the first control cycle of a braking operation.

Embodiment 22:

[0107] Method according to embodiment 21, characterized in that the valve for a first wheel is closed earlier than the valve of the second wheel.

Embodiment 23:

[0108] Method according to one of the embodiments 13 to 22, characterized in that the control and regulating device has a memory in which the pressure and/or the pressure signal, such as, for example, motor current or piston position, adjusted at the time of closing the associated valve of a wheel brake, is stored.

Embodiment 24:

[0109] Method according to any one of embodiments 13 to 23, characterized in that during the control by means of the piston-cylinder system, an initial pressure is controlled which is approximately 10-30%, in particular 20%, above the pressure to be controlled.

Embodiment 25:

[0110] Method according to any of embodiments 13 to 24, characterized in that the motor or piston, respectively, is held in its position for a short time after reaching the pressure build-up in order to ensure that the last opened solenoid valve is completely closed.

Embodiment 26:

[0111] Method according to one of the embodiments 13 to 25, characterized in that the response time tVA of the drive means is small in order to achieve a large pressure decrease rate (dp_down/dt) and/or pressure build-up rate (dp_up/dt), in particular such that the pressure change rate is greater than 1500 bar per second.

Embodiment 27:

[0112] Method according to any one of embodiments 13 to 26, characterized in that the rate of pressure reduction (dp_ab/dt) is selected or set to be very high if a plurality of wheels are simultaneously designated by the controller for pressure reduction.

Embodiment 28:

[0113] Method according to any one of embodiments 13 to 27, characterized in the controller calculates the optimum pressure for achieving the optimum slip for the braked wheel and the pressure build-up for the associated wheel brake is carried out up to a pressure which is slightly smaller, in particular 1-20%, preferably 5-10% smaller, than the calculated optimum pressure, in such a way that renewed exceeding of the optimum slip is avoided.

Embodiment 29:

[0114] Method according to embodiment 28, characterized in that, in order to achieve slip which is as optimal as possible, the pressure build-up takes place in steps, with a large pressure increase according to the control cycle taking place first, followed by pressure holding phases alternating with pressure build-up phases with small pressure changes in each case.

Embodiment 30:

[0115] Method according to embodiment 29, characterized in that during a pressure holding phase for a first wheel, a pressure build-up phase for a second wheel is initiated or carried out by means of the same piston-cylinder system.

Embodiment 31:

[0116] Method according to any one of embodiments 13 to 30, characterized in that the control means, when controlling the 2/2-way valves, takes into account their response times or dead times, in such a way that the 2/2-way valves receive a command to open or close earlier by the response time of the valve, so that the valve is actually open or closed at the calculated time.

Embodiment 32:

[0117] Method according to any one of embodiments 13 to 31, characterized in that the control device controlling the 2/2-way valves derives the response time of the valves from the reaction of the activated drive device or the piston adjustment path that has occurred after corresponding activation and stores it in a memory for the subsequent control.

Embodiment 33:

[0118] Method according to any one of embodiments 13 to 32, characterized in that, for pressure build-up or pressure reduction in at least one wheel brake, the piston of the piston-cylinder system is already being adjusted and the associated valve or valves opens or open later.