ELECTROMECHANICAL BRAKE DEVICE FOR A MOTOR VEHICLE

Description

State of the art

[0001] The invention relates to an electromechanical braking device for a motor vehicle, comprising an actuating device driven by at least one electric motor, of which a brake part is adjustable, wherein the motor has a stator and a rotor rotatably mounted therein about a motor axis, wherein the stator has a stator outer diameter and a stator inner diameter and comprises a number of radially inwardly projecting stator teeth distributed around the circumference, which have a tooth width in the circumferential direction on the stator outer diameter and between which a stator gap is arranged on the stator inner diameter, having a gap width, and which carry at least one stator winding, wherein the rotor comprises a rotor diameter and a number of poles of rotor magnets distributed around the outer circumference, which differ from the number of stator teeth and which are designed as permanent magnets and extend axially parallel to the rotor axis in a rod-like manner.and each having a radial magnet thickness and a circumferential magnet width, and the rotor magnets are each arranged at a magnet distance on the rotor diameter, wherein a radial air gap with a radial air gap width is formed between the stator teeth and the rotor magnets, wherein the ratio of the magnet thickness to the magnet width is between 0.4 and 0.55.

[0002] Such a braking device of a motor vehicle is designed as a friction brake, in which a brake element supported on the chassis and stationary relative to the rotation of the wheel to be braked can be brought into braking engagement by means of an adjusting device with a counter-braking element that rotates with the wheel. In braking engagement, a frictional contact is created between the brake element and the counter-braking element, whereby the braking torque generated by friction is greater the higher the adjusting force exerted by the adjusting device in the direction of adjustment.

[0003] A common design is the disc brake, which is known in principle, where the counter-braking element is formed by a brake disc rotating with the wheel and axially gripped on both sides by a brake caliper. A brake component, usually a brake pad, can be adjusted in an axial direction by means of at least one preferably linear actuating drive axially supported on the brake caliper, thereby bringing it into frictional contact with an axial side of the brake disc. During braking, the brake disc is frictionally clamped between the adjusted brake component and another brake component axially opposite, also supported on the brake caliper.

[0004] From DE 10 2017 123 266 A1, it is known that the adjusting device has two actuating drives arranged in series in the adjustment direction. Each actuating drive has a drive element on the drive side and an output element on the output side that is linearly adjustable relative to it in the axial adjustment direction. To realize an adjustment movement, each drive element has a drive wheel, preferably a gear wheel such as a toothed gear or the like, which can be driven to rotate about its axis by an electric motor, the actuating motor. The rotation of the drive wheel is converted in each actuating drive into a relative adjustment movement or stroke of the output element relative to the drive element in the axial adjustment direction. In the prior art of this type, the two drive wheels of the first and second actuating drives are arranged coaxially on a common axis lying in the axial adjustment direction.

[0005] An actuating drive is a lifting or adjusting device that acts axially in the direction of adjustment. For example, an actuating drive can have a spindle drive in which the drive element has a spindle nut and the output element has a threaded spindle engaging with it, or vice versa. Other designs of actuating drives can also be used, which may include, for example, ramp bearings, cam or cam discs, rocker pin assemblies, or the like, and which also convert a rotation of the drive element into a linear adjustment of the output element.

[0006] Each actuator is driven by a motor designed as an internal permanent magnet synchronous motor. This motor has a stator with stator windings wound on a plurality of stator teeth distributed around its circumference and directed radially inwards towards the motor shaft. The number of stator teeth corresponds to a multiple of the number of phases of the motor current supply. The rotor, which is rotatably mounted around the motor shaft, has a number of permanent magnets distributed around its outer circumference corresponding to the number of poles. These permanent magnets form the rotor magnets and are located radially inwards opposite the stator teeth at a distance equal to the circumferential air gap.

[0007] A braking device of the type mentioned above is known, for example, from US 2016 / 294235 A1.

[0008] For a braking device of this type, high operational reliability and efficiency are essential, along with the lowest possible weight and installation space requirement for the motor(s). Motors known in the prior art have disadvantages in this respect.

[0009] In view of the problems explained above, it is an object of the present invention to enable an efficient drive and a compact design of a braking device of the generic type. Description of the invention

[0010] This problem is solved according to the invention by the electromechanical braking device with the features of claim 1. Advantageous further developments are set out in the dependent claims.

[0011] In an electromechanical braking device for a motor vehicle, comprising an actuating device driven by at least one electric motor, of which a brake part is adjustable, wherein the motor has a stator and a rotor rotatably mounted therein about a motor axis, wherein the stator has a stator outer diameter and a stator inner diameter and comprises a number of radially inwardly projecting stator teeth distributed around the circumference, which have a tooth width in the circumferential direction at the stator outer diameter and between which a stator gap is arranged at the stator inner diameter, having a gap width, and which support at least one stator winding, wherein the rotor comprises a rotor diameter and a number of poles of rotor magnets distributed around the outer circumference, which differ from the number of stator teeth and which are designed as permanent magnets and extend axially parallel to the rotor axis in a rod-like manner,and each having a radial magnet thickness and a circumferential magnet width, and the rotor magnets are each arranged at a magnet distance on the rotor diameter, wherein a radial air gap with a radial air gap width is formed between the stator teeth and the rotor magnets, wherein the ratio of the magnet thickness to the magnet width is between 0.4 and 0.55, it is provided according to the invention that the ratio of the air gap width to the tooth width is between 0.45 and 0.65.

[0012] The radial inner surfaces of the stator teeth define a cylindrical inner diameter of the stator. The inner cross-section of the stator teeth is correspondingly cylindrical.

[0013] The tooth width corresponds to the tooth root width measured circumferentially in a radially outer area of ​​a stator tooth, facing away from the air gap. The stator teeth can be connected there to a preferably hollow cylindrical stator body, which is continuous or segmented in the circumferential direction.

[0014] In their radially inner end region facing the air gap, adjacent stator teeth are spaced apart in a stator gap which has a gap width measured in the circumferential direction.

[0015] The rotor magnets are attached to the outside of the rotor in such a way that their outer surfaces are radially opposite the inner surfaces of the stator teeth at a distance equal to the air gap width.

[0016] The rod-shaped rotor magnets have a solid profile elongated in the axial direction parallel to the motor axis. They preferably have a continuous cross-section and are preferably made of a hard magnetic material with high coercivity, for example SmCo, NdFeB alloys or the like.

[0017] A dimensional ratio is defined for the cross-section of the magnet material in the radial and circumferential directions. This allows for an optimized distribution of the magnetic field, advantageously reducing the dimensions and weight of the rotor. The result is the ability to achieve a relatively high motor torque with a relatively small rotor diameter. Consequently, the stator's inner diameter, and therefore its overall dimensions and weight, can also be reduced, enabling a more powerful, lightweight, and compact brake design.

[0018] The advantageous reduction in installation space and weight can be particularly high in the design of a braking device that has at least two motors. Preferably, each of the motors is designed according to the invention.

[0019] According to the invention, the ratio of the air gap to the tooth width is between 0.45 and 0.65. A ratio of the air gap width to the tooth width is particularly preferred and is between 0.5 and 0.6. This allows for an optimized magnetic flux, enabling high motor torque and smooth running with compact dimensions and low weight.

[0020] The rotor magnets can be designed to have a rectangular cross-section. The solid profile of the rotor magnet, which is rectangular in cross-section along its axial length, has a radially outer surface tangential to a circumferential direction (i.e., perpendicular to the diameter) and an inner surface parallel to this radially inner surface, as well as side surfaces parallel to each other and parallel to the diameter of the rotor. The outer surface, the inner surface, and the side surfaces can preferably be essentially planar. Such rotor magnets can advantageously be produced cost-effectively with optimized magnetic properties and integrated into the rotor.

[0021] Alternatively, the rotor magnets can be designed to have a hollow cylindrical segment cross-section. This hollow cylindrical segment cross-section rotor magnet has a cylindrical outer surface coaxial with the rotor axis, the diameter of which corresponds to the rotor diameter, and a coaxial cylindrical inner surface, the diameter of which corresponds to the rotor diameter minus the magnet thickness. The flat side surfaces bounding the rotor magnet circumferentially are parallel to each other and parallel to the rotor diameter. The cylindrical outer surface preferably has a diameter corresponding to the stator's inner diameter minus the air gap width. The advantage of this design is that the air gap width can be constant over the entire circumference of the rotor, thus enabling optimized magnetic flux.

[0022] As an alternative to the two aforementioned designs, the rotor magnets can have a segmented cross-section with a cylindrical outer surface and a flat inner surface. The rotor magnet has a cylindrical outer surface coaxial with the rotor axis, the diameter of which corresponds to the rotor diameter, and a substantially flat inner surface tangential to one circumferential direction, i.e., perpendicular to the diameter. The advantage is that the air gap width can be kept constant over the entire circumference, allowing for a relatively high magnetic flux.

[0023] An advantageous embodiment is that a rotor magnet extends along the rotor diameter with respect to the motor axis over a magnetic angular segment, and the rotor magnets are arranged offset from each other by a pole angle, with the ratio of the magnetic angular segment to the pole angle being between 0.65 and 0.75. This allows for an optimized magnet spacing between the rotor magnets in the circumferential direction, which is advantageous for a compact design.

[0024] It is preferably possible for the number of poles to be 8 or 10. The number of poles is determined by the number of rotor magnets distributed around the circumference. A high number of poles can result in a more even motor torque and improved smoothness of operation, although this is more complex to manufacture. Conversely, a low number of poles can be achieved with less effort, but with less smooth operation and a less even motor torque. It has been found that the number of poles according to the invention is optimal for use in the braking device.

[0025] It is advantageous for the number of stator teeth to be 6 or 9. The number of stator teeth is a multiple of the number of phases of the current supplying power to the motor. A combination of 6 stator teeth with 8 rotor magnets is advantageous for use in a braking device, and a combination of 9 stator teeth with 10 rotor magnets is particularly advantageous.

[0026] It is advantageous that the ratio between the stator outer diameter and the stator inner diameter is between 0.55 and 0.65. This allows optimized magnetic properties to be achieved in a compact design.

[0027] An advantageous embodiment may provide that the ratio between the air gap width and the stator inner diameter is between 0.015 and 0.03. This ratio is advantageous with regard to high efficiency and a compact design.

[0028] The invention further comprises an electromechanical braking device for a motor vehicle, comprising a drive carrier to which an electric motor and an actuating device are attached. The actuating device is gear-coupled to a motor shaft and a brake element is adjustable by the actuating device. The motor has a motor housing in which the axially extending motor shaft is supported in an end-face bearing cap and projects axially from the bearing cap. The motor housing is fixed to the drive carrier. Preferably, the drive carrier has a recess in which the motor housing can be positively received in the axial and radial directions, and the motor housing can be clamped against an axial support surface of the recess at its end face.

[0029] The recess provides a mounting for the motor, in which the motor housing can be precisely aligned axially and radially relative to the drive carrier. The recess has an opening through the drive carrier, through which the motor shaft passes perpendicularly.

[0030] The support surface can be formed by a radially inwardly projecting support projection into the opening of the recess, for example, by a step or similar feature located at the edge of the opening. The open cross-section of the recess is matched to the outer cross-section of the motor housing in such a way that the housing can be inserted axially – by definition forwards in the direction of the motor axis defined by the motor shaft – into the recess with minimal radial play until it abuts axially against the support surface. The motor shaft protrudes on the side of the drive carrier facing away from the motor.

[0031] The bearing cover, located on the front face of the motor housing, can be supported against the support surface and clamped to the motor housing.

[0032] One advantage is that the motor housing is held in the recess with a positive fit in the radial direction when inserted, and is also positively supported in the axial direction when it comes to rest against the support surface. Thus, the motor, mounted in the recess, is positioned in a spatially defined manner relative to the drive carrier. Clamping to fix the motor in place can be performed after insertion into the recess. This significantly simplifies assembly and allows for a defined gear engagement with the actuator, for example, via meshing gear wheels on the motor shaft and the actuator.

[0033] It is possible that the support surface is formed on a projection that extends radially inwards within the recess. The projection can preferably have a step or the like circumferentially around at least part of the circumference of the recess, on which the support surface is formed parallel to the planar extent of a mounting section of the drive carrier.

[0034] It is preferred that the bearing cap is attached to the end face of the motor housing. The bearing cap can initially be provided as a separate part and then, during assembly, joined to the motor housing after the rotor and motor shaft have been inserted. This allows for efficient motor assembly.

[0035] It may be provided that the bearing cover is axially supported against the support surface.

[0036] The bearing cap, which is mounted axially at the front of the motor housing, can rest against the support surface with its front face, the side facing away from the motor housing. On its rear side, the side facing axially away from the front, the bearing cap is connected to the motor housing.

[0037] It is advantageous that the bearing cap can be clamped between the support surface and the motor housing. The bearing cap has at least one section that is axially positioned between the motor housing and the support surface. Thus, the bearing cap can be axially clamped against the motor housing by clamping the latter against the support surface. In other words, the bearing cap is at least partially within the force path of the clamping of the motor housing to the drive bracket. This allows the fixing of the motor to the drive bracket and the fixing of the bearing cap to the motor housing to be carried out in a single assembly step during clamping, thus advantageously reducing the effort involved.

[0038] For example, the bearing cap can have an axial projection that fits into an axial opening in the motor housing and a circumferential collar that projects radially outward beyond the cross-section of the opening. The collar can have essentially the same outer cross-section as the motor housing and is positioned axially between the support surface and the motor housing. When the motor housing is clamped against the support surface, the bearing cap can thus be simultaneously supported on the support surface and secured to the motor housing. An advantage of this is that the bearing cap only needs to be temporarily connected to the motor housing before the motor is installed in the brake assembly, and the final fixing can be achieved in a single assembly step when clamping the motor to the drive bracket. The clamping mechanism therefore fulfills a dual function: fixing the motor to the drive bracket and fixing the bearing cap to the motor housing.For example, the bearing cap with an axial projection can simply be inserted into the motor housing or placed onto the motor housing using a friction-fit mechanism. This eliminates the need for complex screw connections, welds, or other joining processes between the bearing cap and the motor housing, thus simplifying the motor's construction and saving weight.

[0039] It may preferably be provided that the motor housing has a flange element which is axially spaced from the end face and projects radially beyond the recess.

[0040] The flange element projects radially outwards from the motor housing and protrudes beyond the recess. It can be connected to the drive carrier. Fasteners can be provided for this purpose, which can be connected to the drive carrier outside the recess to axially clamp the motor housing against an outer surface of the brake housing. For example, the flange element can preferably have several axial flange bores distributed around its circumference, through which fasteners such as screws or the like can be passed and screwed into corresponding threaded bores in the drive carrier. Because the flange element has an axial distance from the front face of the bearing cap, the axial clamping of the flange element against the drive carrier clamps the motor housing, which is inserted into the recess, together with the bearing cap, against the support surface located in the recess.In this way, the flange element allows the motor to be fixed to the drive carrier in the position defined by the recess, and at the same time the bearing cover can be clamped and firmly connected to the motor housing.

[0041] The axial distance of the flange element from the end face of the bearing cap, which is mounted at the front of the motor housing, is preferably greater than the depth of the recess, measured from the support surface to the outer surface of the drive carrier in the area of ​​the flange element. This allows the bearing cap to be clamped axially between the support surface and the motor housing by clamping the flange element. The advantage is that the bearing cap can be securely connected to the motor housing without additional fasteners. This allows for a simpler and lighter motor design.

[0042] The motor housing may be designed to have at least a partial hollow cross-section to which the bearing cap can be positively locked. The motor housing may, for example, be cup- or pot-shaped and axially closed at the front by the bearing cap mounted on it. The motor shaft, which carries the rotor, may be supported at one end inside the motor housing and rotatably mounted at the other end in the bearing cap, passing through it to the outside. The hollow cross-section may, for example, be a substantially cylindrical tube section, from the open end face of which the bearing cap can be secured by axial clamping.For example, a cylindrical extension of the bearing cap can be positively inserted into the opening of the hollow cross-section, and an essentially ring-shaped collar can be clamped between the support surface and the end face of the pipe section in the manner described above.

[0043] It can be advantageous for the motor housing, bearing cover, and / or drive carrier to have a cast component. The cast component can be an injection-molded part made of a thermoplastic material, which may optionally be fiber-reinforced to increase strength, or a die-cast part made of a metallic material, for example, aluminum, magnesium, or zinc alloys. Complex shapes can be efficiently realized using the casting process. For example, the recess according to the invention and optionally other functional elements can be integrally formed on the drive carrier. Similarly, the flange element and optionally other functional elements can be integrally formed on the motor housing. For example, a projection for connection to the motor housing, a bearing receptacle for the motor shaft, or the like can be integrally formed on the bearing cover.

[0044] It is possible for an elastic retaining element and / or sealing element to be arranged between the recess and the motor housing. For example, an elastically deformable O-ring made of a rubber or polymer material can be radially clamped between a circumferential inner surface of the recess and an outer surface of the motor housing. This allows the motor housing to be temporarily held in position on the drive carrier by simply inserting it axially into the recess, either by force or friction, thus simplifying the subsequent clamping process. It can be advantageous to provide an O-ring that is received in a circumferential groove of the recess or the motor housing and is thereby positively locked in the axial direction.

[0045] Furthermore, an O-ring or other elastic sealing element can be used to effectively seal the motor housing in the recess against the ingress of moisture or contaminants.

[0046] Preferably, the drive carrier may have at least two recesses. A motor can be fixed in each of the recesses, each of which can drive one actuator of the positioning device. This makes it possible to mount two motors on the drive carrier to drive two actuators of the positioning device. An advantage of the invention is that both motors can be positioned and mounted easily and securely relative to the actuators.

[0047] An advantageous embodiment may provide that the braking device comprises an adjusting device and a brake part connected thereto, which is adjustable by the adjusting device along an axis and can be brought into brake engagement with a counter-brake part, wherein the adjusting device has a first adjusting drive and a second adjusting drive coupled serially thereto, wherein the first adjusting drive has a rotatably driven first drive wheel, and the second adjusting drive has a rotatably driven second drive wheel coaxial to the first drive wheel, wherein a coupling device is arranged between the first drive wheel and the second drive wheel.

[0048] The actuator can be driven by at least one electric actuator motor. This motor is preferably engaged with at least one drive wheel via a gearbox. Preferably, one actuator motor can be provided for the first and one for the second drive wheel. According to the invention, the actuator motor(s) can be controlled by a wheel brake control unit associated with the brake device.

[0049] It may be provided that the coupling device is designed as a friction coupling with a friction element that can be frictionally connected to a counter-friction element in the coupling engagement.

[0050] In the following, the first and second drive wheels will also be referred to together as the two drive wheels or simply as the drive wheels.

[0051] The drive wheels can each be designed as a gear, for example as a spur gear, or as a belt or toothed belt wheel or worm gear, so that in general a gear wheel is provided via which a drive torque from an electric actuator can be coupled into the actuator.

[0052] A friction clutch is implemented between the drive wheels. This clutch comprises a friction element, which is torque-locked to one of the drive wheels, and a corresponding counter-friction element, which is torque-locked to the other drive wheel. The friction element can be engaged with the counter-friction element at any relative angular position. This creates a purely force-locking clutch, unlike a positive-locking detent connection. This allows the relative position of the drive wheels to be continuously preset, unlike the discrete detent positions of a detent connection. Accordingly, a smooth, continuous adjustment of the second actuating mechanism relative to the first actuating mechanism is possible, and the air gap can be continuously adjusted.This is particularly advantageous with regard to the consistent adjustment of the brake's optimal operating point to the continuous wear of the brake component during operation, i.e., the continuous wear of the brake lining. Compared to a purely incremental adjustment option, a consistently improved brake response can be achieved, resulting in increased operational safety and greater ease of use.

[0053] A further advantage over a detent clutch is that essentially no axial relative movement between the clutch elements engaged is required to actuate and release the clutch device. This is particularly true for the drive wheels and the detent elements, which must necessarily be movable relative to each other to create and release the detent-type positive engagement. In contrast, the frictional engagement between the friction and counter-friction elements according to the invention can be achieved simply by applying an axial actuating force, without requiring any axial movement relative to each other. This allows for a simpler and more reliable design of the clutch device.

[0054] It is preferably provided that the friction clutch has a defined, predefinable clutch torque. The clutch torque indicates the maximum differential torque that can be positively transmitted between the friction element and the counter-friction element during clutch engagement. When the clutch torque is exceeded, the clutch device slips, causing the two drive wheels to rotate relative to each other. An advantage of this is that the friction clutch according to the invention slips continuously, thus enabling improved, uniform readjustment of the air gap. Furthermore, there is no need to consider and compensate for axial movement of detent elements, as is the case with known detent clutches.

[0055] It is advantageous for the friction element and the counter-friction element to be arranged coaxially. This coaxial arrangement corresponds to the coaxial arrangement of the drive wheels. The friction element and the counter-friction element can be arranged in a simple and compact design in the area of ​​the axially opposing end faces of the drive wheels. Due to the purely frictional transmission of the clutch described above, no moving parts are required.

[0056] In an advantageous embodiment, the friction element and the counter-friction element can be conical. The friction element can have a conical section, converging at least partially in the axial adjustment direction, with a conical friction surface. This conical section can be configured as an outer or inner cone. The counter-friction element can be configured as an inner or outer cone, converging in the opposite direction, and has a conical counter-friction surface. To generate the clutch engagement, the outer cone engages with the inner cone, whereby the conical friction and counter-friction surfaces are frictionally engaged against each other by an axial actuating force of the clutch. An advantage of this design is that the cone allows the axially acting actuating force of the clutch to be converted into the normal force acting between the conical friction surfaces in frictional contact.Thus, a shallower slope allows a relatively small axial actuating force to be converted into a larger normal force in the friction contact, making it possible to achieve a high clutch torque even with a relatively small axial actuating force of the clutch.

[0057] Alternatively or additionally to the aforementioned design, the friction element and the counter-friction element can be designed planar. In this case, the corresponding friction surfaces are designed, at least in sections, as planar axial surfaces, similar to a disc clutch. This allows for a space-saving arrangement, particularly when only a relatively small clutch torque is required.

[0058] It is preferable for the friction element and the counter-friction element to be preloaded against each other. Preferably, the friction element and the counter-friction element are elastically or resiliently preloaded against each other. The friction and counter-friction surfaces are pressed against each other in frictional engagement with a predetermined axial preload force. An elastic preload element, such as a spring element or the like, is preferably provided to generate the preload force. The coupling torque of the friction clutch is determined by the actuating force acting perpendicular to the friction contact, i.e., the force applied axially between the friction and counter-friction elements, with the coupling torque being greater the greater the preload force. This offers the advantageous possibility of simply specifying the coupling torque by the preload force exerted by the preload element.For example, in the case of an axially elastic spring element, such as a compression spring, the applied preload force can be easily determined and adjusted by the spring constant and the compression of the spring.

[0059] The aforementioned embodiment can advantageously be realized by making the friction element and / or the counter-friction element axially displaceable and supported against the first or second drive wheel by an axially acting spring element. The friction element or the counter-friction element is connected to one of the drive wheels in a torque-locking and axially displaceable manner, for example, via radially projecting drivers that generate a circumferentially effective positive locking connection. The spring element, preferably designed as an axially acting compression spring, which is axially clamped between the friction element or the counter-friction element and one of the drive wheels, ensures that the friction or counter-friction element is axially pre-tensioned against the corresponding counter-friction or friction element, which is axially supported on the other drive wheel, i.e., pressed axially against it in frictional contact.The corresponding counter-friction or friction element is rotationally connected to the other drive wheel. Alternatively or additionally, the counter-friction element can be supported on one of the drive wheels by a spring element. An advantage of this arrangement is that this friction coupling can be integrated between the drive wheels in a simple and space-saving manner.

[0060] In an advantageous embodiment, the friction element and / or the counter-friction element can be arranged in the first or second drive wheel. For example, one drive wheel can be designed to be essentially drum-shaped, so that the friction or counter-friction element can be arranged in an interior enclosed by the rotating gear or ring gear. This enables a compact design protected against external influences. For example, the drive wheel of the first actuating drive can have a conical friction element that engages axially in a counter-friction element designed as an internal cone, which is arranged at least partially inside the second drive wheel.

[0061] A particularly compact design can be achieved – especially in the last-mentioned version – by arranging the drive wheels within the axial extent of the actuating drives, i.e., not protruding axially on one side.

[0062] It is preferred that the friction element and / or the counter-friction element have a friction lining. The friction and counter-friction elements preferably have a metallic base body, for example made of steel. To prevent metal-to-metal contact, a coating or lining can preferably be applied to create a friction pair with a defined frictional force, for example made of sintered, metal and / or ceramic friction materials, composite materials or the like. This ensures a defined, reproducible coupling torque.

[0063] It is possible for an actuating drive to include a spindle drive. In this configuration, a threaded spindle engages with a spindle nut in a manner known per se, and a relative rotating drive is transmitted via a drive wheel connected to the threaded spindle or the spindle nut. It is possible for the spindle nut to form the drive-side input element of the actuating drive, and the threaded spindle the output-side output element, which is linearly adjustable relative to it, or vice versa.

[0064] It is possible for an actuating drive to have a ball ramp arrangement, a wedge disc arrangement, or a rocker pin arrangement. In a ball ramp arrangement, also known as a ramp bearing, the drive and driven elements preferably have cam discs with raceways or ramps inclined to the axis, between which circumferentially rolling balls are arranged. A relative rotation, due to the balls rolling on the ramps, causes the driven element to be axially displaced relative to the drive element. In a rocker pin arrangement, which is known per se, rocker pins are arranged between the drive and driven elements and supported circumferentially in such a way that, depending on the direction of rotation, they are inclined more or less to the axis during a relative rotation, thus also allowing the distance between the drive and driven elements to be adjusted.

[0065] In the adjusting device, two identically acting actuators can be combined as first and second actuators, for example, two spindle drives. It is also possible to combine two different designs, for example, a ball ramp arrangement as the first actuator and a spindle drive as the second actuator for adjusting the air gap. The respective characteristic properties of each design can be optimally utilized. For example, a ball ramp arrangement can easily achieve a non-linear adjustment characteristic, and / or at least partially self-locking properties, and / or a defined dead center or extended position that enables a defined adjustment path. Achieving these positive properties may require, at least in part, precise specification of the air gap, which can be easily accomplished with the friction clutch according to the invention.

[0066] A braking device according to the invention can comprise an adjusting device and a brake part associated therewith, which is adjustable by the adjusting device along an axis and can be brought into brake engagement with a counter-brake part, wherein the adjusting device has a first adjusting drive and a second adjusting drive coupled serially thereto, wherein the first adjusting drive has a rotatably driven first drive wheel, and the second adjusting drive has a rotatably driven second drive wheel coaxial to the first drive wheel, wherein a coupling device is arranged between the first drive wheel and the second drive wheel.

[0067] The actuator can be driven by at least one electric actuator motor. This motor is preferably engaged with at least one drive wheel via a gearbox. Preferably, one actuator motor can be provided for the first and one for the second drive wheel. According to the invention, the actuator motor(s) can be controlled by a wheel brake control unit associated with the brake device.

[0068] In the last-mentioned embodiment of the brake device, it may preferably be provided that the clutch device is designed as a friction clutch with a friction element that can be frictionally connected to a counter-friction element in the clutch engagement.

[0069] This makes the advantages previously explained in connection with the braking system achievable.

[0070] To implement the method according to the invention, the brake device can be provided with an actuating device that can be coupled to an actuating motor, comprising a first actuating drive and a second actuating drive coupled serially thereto, and which acts on a brake element that can be brought into brake engagement with a counter-brake element in the direction of an axis, wherein the first actuating drive has a rotatably driven first drive wheel onto which a first drive torque can be applied for actuation, and the second actuating drive has a rotatably driven second drive wheel, coaxial to the first drive wheel, onto which a second drive torque can be applied for actuation, wherein a coupling device is arranged between the first drive wheel and the second drive wheel, wherein according to the invention the coupling device is designed as a friction clutch and has a predefinable coupling torque.Upon exceeding this threshold, the first drive wheel slips relative to the second drive wheel, whereby, to actuate the first actuating mechanism, the first drive wheel and the second drive wheel are driven synchronously, so that the second actuating mechanism remains unactuated, and, to actuate the second actuating mechanism, the second drive wheel is driven, and the first drive wheel is stopped relative to it, so that the friction clutch slips and the first actuating mechanism remains unactuated.

[0071] The features mentioned above in connection with the braking device according to the invention can be used individually and in combinations to implement the method according to the invention.

[0072] To adjust the first actuating drive, an actuating torque can be coupled into the first drive wheel by means of a first electric actuator, and accordingly the second actuating drive can be driven by a second electric actuator.

[0073] In normal braking operation, the first and second drive wheels rotate synchronously. This can be achieved either by driving the first and second drive wheels with synchronized drive torques from the first and second actuators, respectively. Alternatively, the second drive wheel can be driven synchronously by the clutch mechanism when the first drive wheel is driven, as long as the transmitted drive torque remains below the clutch torque. In this operating mode, the second actuator remains unactuated and rotates freely along with the brake element.

[0074] In this method, the coupling device can continuously and smoothly slip when the coupling torque is exceeded, thus adjusting the air gap. This can be achieved, for example, by locking the drive wheel of the first actuator, such as by a brake or by appropriately controlling the first drive motor, while a second drive motor applies a second drive torque to the second drive wheel, which is greater than the coupling torque. This rotates the second drive wheel relative to the first drive wheel, and by actuating the second actuator, the air gap can be continuously and precisely adjusted, optimally compensating for the progressive wear of the brake element or brake lining.

[0075] It is possible that the first drive wheel and the second drive wheel are torque-locked via the friction clutch to generate a synchronous drive.

[0076] This does not require synchronous drive of the two drive wheels by the actuators. Any torque differences can be compensated for within specified tolerances.

[0077] It can be advantageous to provide a higher coupling torque when the first actuating mechanism is engaged than when the second actuating mechanism is engaged. The first actuating mechanism is engaged by the synchronous drive of the first and second drive wheels. The friction element and the counter-friction element are pre-tensioned against each other by the spring force of the spring element, and the adjusting force of the first actuating mechanism acts in opposition to the spring force. This results in a relatively high coupling torque. If, however, only the second drive wheel is rotated to adjust the air gap, only the spring force acts, resulting in a lower coupling torque. This simplifies the adjustment of the air gap. Description of the drawings

[0078] Advantageous embodiments of the invention are explained in more detail below with reference to the drawings. Specifically, they show: Figure 1 shows a brake device according to the invention in a schematic perspective view, Figure 2 shows a side view of the brake device according to the invention. Figure 1 Figure 3 shows the adjusting device of the brake device according to the invention. Figure 1 Isolated in a schematic perspective view, Figure 4 shows a section QQ through the braking device according to Figure 1 , Figure 5 the first actuating drive of the brake device according to Figure 1 Isolated in a schematic perspective representation, Figure 6 shows an enlarged detail view of the adjusting device. Figure 4 Figure 7 shows a longitudinal section through a motor mounted on the drive carrier of the brake device, Figure 8 shows a cross-section through a motor according to the invention. Figure 7 Figure 9 shows an enlarged cross-section through the rotor of the motor according to Figure 8 Figure 10a,b,c Cross-sections through different designs of a rotor magnet of the motor according to Figure 8 or 9 . Embodiments of the invention

[0079] In the various figures, identical parts are always marked with the same reference symbols and are therefore usually only named or mentioned once.

[0080] Fig. 1 Figure 1 shows a complete brake device according to the invention, designed as a disc brake. This includes a brake disc 2, which forms a counter-brake element and is connected to a vehicle wheel (not shown) rotatable about a wheel axle R. A brake caliper 3 engages the two axial end faces of the brake disc 2.

[0081] Brake disc 2 is designed here as an unventilated brake disc made of solid material. Alternatively, it can also be designed as an internally ventilated brake disc.

[0082] An electric brake actuator 4 according to the invention is attached to the brake caliper 3, which is in Figure 3shown in a separate, isolated schematic perspective view, and in the Figures 4 to 6 will be explained in detail.

[0083] The brake actuator 4 comprises an actuating device 5 which extends axially in the direction of an axis A, which is parallel to the wheel axis R and indicates the adjustment direction V of the actuating device 5.

[0084] As shown in the sectional view of Figure 4 As can be seen along axis A, the brake disc 2 is arranged axially between two brake pads 31 and 32. One brake pad 31 is fixedly supported on the brake caliper 3 on the side facing away from the brake actuator 4. The other brake pad 32, which forms a brake component according to the invention, is attached to the adjusting device 5 and can be adjusted by it in the axial adjustment direction V given by axis A to generate the brake engagement on the brake disc 2, as shown in Figure 4 as indicated by the arrow.

[0085] In the unactuated state of the brake device 1, there is an axial air gap L between the brake disc 2 and the adjustable brake pad 32, which is Figure 4 is drawn schematically in an exaggeratedly wide way.

[0086] The construction of the adjusting device 5 is described in Figure 4 and in the enlarged section thereof in Figure 6 depicted.

[0087] The adjusting device 5 comprises a first adjusting drive 6, which has a ball ramp arrangement, also referred to as ramp bearing, and a second adjusting drive 7, which has a spindle drive and is coupled to it axially (with respect to the axis A) in series.

[0088] The first actuating drive 6, which in the example shown is designed as a ball ramp arrangement or ramp bearing, comprises an axially and rotationally fixed cam disk 61 on the drive side and an output-side cam disk 62. Balls 63 are arranged between the cam disks 61 and 62. As shown in the schematically isolated view of Figure 5 As can be seen, the cam discs 61 and 62 have axially opposing, ramp-like raceways 64, inclined to axis A, between which balls 63 can roll. A rotation of the output-side cam disc 62, in Figure 5 The upward movement, relative to the stationary drive-side cam disk 61 – as schematically indicated by the curved arrows – results in a linear adjustment of the output-side cam disk 62 in the adjustment direction V parallel to the axis A. This allows the brake pad 32 to move as shown in Figure 4shown by actuating the first actuating drive 6 into brake engagement.

[0089] The cam disk 62 is connected to a coaxial gear 65, which is designed as a spur gear and forms a drive gear in the sense of the invention.

[0090] The gear 65 engages with a first electric actuator 41, which is also referred to simply as motor 41. This enables the rotating drive of the cam disk 62 and thus the actuation of the first actuating drive 6.

[0091] The second actuating drive 7, which in the example shown is designed as a spindle drive, has a threaded spindle 71 on the output side, which engages in the internal thread of an input-side spindle nut 72. This internal thread is formed in the output-side cam disk 62 of the first actuating drive 6, so that the functions of the output-side cam disk 62 and the input-side spindle nut 72 are combined in one component.

[0092] The threaded spindle 71 is connected via a hub part 74 to a coaxial gear 75, which is rotatably mounted axially fixed in the brake actuator 4. The threaded spindle is coupled to the gear 75 in a torque-locking but axially displaceable manner via drivers 73, which may, for example, have radially projecting projections or teeth that engage axially displaceably in axial slots of the hub part 74.

[0093] The gear 75, like the gear 65, can be designed as a spur gear and is arranged coaxially adjacent to it. This gear 75 engages with a second electric actuator 42, which is also referred to simply as motor 42. This enables the rotational drive of the threaded spindle 71 and thus the actuation of the second actuating drive 7.

[0094] The threaded spindle 71 is axially connected via a thrust bearing 43, for example an axial roller bearing as shown, to a pressure piece 44, to which the movable brake pad 32 is attached, as shown in Figure 4 This is recognizable. The pressure piece 44 can also be referred to as a piston.

[0095] The coupling device has a friction element 8, which is directed as a coaxial, conical projection from the cam disk 62 towards the second actuating drive 7. The conical projection has a conical friction surface 81 arranged on the outside of an outer cone. The friction element 81 can preferably be formed integrally with the cam disk 62 / spindle nut 72.

[0096] The friction element 8 is frictionally coupled to a counter-friction element 9 in the clutch engagement. The conical projection of the counter-friction element 9 engages axially in a corresponding conical opening, which has a conical friction surface 91 arranged in an inner cone. In the clutch engagement, the friction surface 81 and the counter-friction surface 91 are in frictional contact with each other, as shown in Figure 6 is clearly recognizable.

[0097] The counter-friction element 9 is coupled to the gear 75 via drivers 92, which engage axially displaceably in corresponding slots 76 in the hub part 74 or the gear 75, in a torque-locking but axially displaceable manner.

[0098] A spring element 93 is arranged between the gear 75 or the hub part 74 connected thereto and the counter-friction element 9. Its axially acting spring force elastically preloads the counter-friction element 9 against the friction element 8. This generates a defined coupling torque for the friction clutch according to the invention, formed by the friction element 8 and the counter-friction element 9.

[0099] In Figure 3 The figure shows how the two motors 41, 42 and the actuator 5 are arranged relative to the brake caliper 3. The drive carrier 100 has been omitted from this figure for clarity.

[0100] Each of the motors 41, 42 has a motor shaft 411, 421 which is driven to rotate about a motor axis M and which lies parallel to the axis A. A gear 412 or 422 is mounted on this shaft, respectively, and each gear meshes with the gear 65 or 75 of the actuating device 5.

[0101] Each motor 41, 42 has a motor housing 413, 423, which in the example shown has a cylindrical basic shape. It is cup-shaped and on its Figure 3 The axial end face facing the viewer is closed by means of a bearing cover 414, 424, wherein the motor shaft 411, 412 carrying a rotor of the motor 41 is supported in the bearing cover 414, 424 and protrudes axially from it.

[0102] Figure 7shows a longitudinal section along the motor axis M through the motor 41 or 42, whereby for better clarity only the reference symbols for the motor 41 are shown, which are also present for the other motor 42.

[0103] The drive carrier 100 has a recess 101 that includes an opening 102 extending through the drive carrier 100. A projection 103, extending radially inwards into the opening cross-section in a stepped manner, has an axial support surface 104 that faces the motor 41. The recess 101 is bounded radially outwards by an inner surface 105 that circumferentially surrounds the motor axis M.

[0104] The inner surface 105 is adapted to the outer diameter of the motor housing 413 in such a way that the latter can be inserted axially in it and is held and supported radially, i.e. transversely to the motor axis M, in a form-fitting manner.

[0105] The bearing cap 414 is inserted axially from the front into the motor housing 413 by means of an axial projection 415. With its front end face – which by definition is at the front and in Figure 7 Pointing to the left - the bearing cover 414 rests axially against the support surface 104 of the recess 101.

[0106] The bearing cover 414 further has a circumferential, radially projecting collar 416, which is arranged axially between the motor housing 413 and the support surface 104.

[0107] The motor housing 413 has flange elements 416 projecting radially outwards beyond the recess 101, with axially through flange bores through which screws 417 serving as fastening elements are passed and screwed into corresponding threaded bores in the drive carrier 100.

[0108] By screwing in and tightening the screws 417, the motor housing 413 is fixed to the drive carrier 100 and clamped to it. In doing so, the bearing cover 414, together with the motor housing 413, is axially pressed against the support surface 104 (in Figure 7 (to the left as indicated by the arrow) clamped, and at the same time axially into the motor housing 413 (in Figure 7 (to the right) pressed in and fixed. The screws 417 thus have a dual function: to fix the motor 41 to the drive carrier 100 and to connect the bearing cover 414 to the motor housing 413. The support surface 104 and the inner surface 105 ensure a defined alignment of the motor 41 relative to the drive carrier 100.

[0109] An O-ring 106 made of an elastic elastomer or rubber material can also be arranged between the motor housing 413 and the inner surface 105, for example, as shown, in a groove running circumferentially around the inside of the inner surface 105. This O-ring is elastically clamped there in the radial direction and ensures that the motor housing 413 can be opened by simply inserting it axially. Figure 7 in the direction of the arrow to the left - it is held in the recess 101 by friction. Furthermore, this allows the motor 41 to be sealed against the drive carrier 100.

[0110] The bearing cap 414 has a receiving opening 418 through which the motor shaft 411 extends. A bearing 419 is arranged between the motor shaft 411 and the receiving opening 418 for the rotatable mounting of the motor shaft 411 relative to the bearing cap 414. This bearing is designed as a rolling bearing, more precisely as a radial groove ball bearing. The receiving opening 418 has an inwardly projecting shoulder section 418a against which the bearing 419 rests in the direction of the motor axis M and is supported.

[0111] The in Figure 7 The motor 41 shown has a stator 1100 fixedly arranged in the motor housing 413, which comprises a laminated core through which a radially outer, substantially hollow cylindrical stator body 1101 is formed, from which several stator teeth 1102 project radially inwards. Stator windings 1103 are arranged from the stator teeth 1102.

[0112] A rotor 1200 is fixedly mounted on the motor shaft 411. This rotor has a total of 10 (ten) rotor magnets 1201 evenly spaced around its circumference. Accordingly, the number of poles is 10 (ten).

[0113] A circumferential air gap 1104 is formed between the outer circumference of the rotor 1200 and the inner circumference of the stator 1100.

[0114] Figure 8 shows a cross-section through the stator 1100 and the rotor 1200 of the motor 41.

[0115] The stator 1100 has a stator outer diameter D1 and a stator inner diameter D2.

[0116] In the example shown, the stator 1100 has nine stator teeth 1102. These extend radially inwards from the rotor body 1101 to the rotor inner diameter D2. On the outside of the rotor body 1102, the rotor teeth 1102 each have a tooth width TW measured in the circumferential direction. Inside, in the region of the rotor inner diameter D2, a stator gap 1105 is formed between adjacent stator teeth 1102, which has a gap width SO measured in the circumferential direction.

[0117] The rotor 1200 has a rotor diameter D3, to which the rotor magnets 1201 are adjacent from the inside.

[0118] The difference between the stator inner diameter D2 and the rotor diameter D3 gives the air gap width AG = D2 - D3 of the air gap 1104.

[0119] Figure 9 shows an enlarged view of the Rotor 1200 Figure 8 .

[0120] The rotor magnets 1201 are arranged uniformly offset by a pole angle BetaP, which in the example with the 10 (ten) rotor magnets 1201 has a value of 360° / 10 = 36°.

[0121] With respect to the motor axis M, the rotor magnets 1201 extend over a magnetic angle BetaM at the rotor diameter D3. The difference between the angles BetaP and BetaM determines the magnet spacing MD measured circumferentially at the rotor diameter D3.

[0122] In the Figures 10a, b, c are alternative cross-sections according to the invention of a single rotor magnet 1201 in the view of Figure 8 shown.

[0123] According to Figure 10b The cross-section is rectangular with flat sides, wherein the magnet width is MW and the magnet thickness is MT. These are in the ratio to each other according to the invention.

[0124] The execution according to Figure 10cIt has a flat inner surface – shown below in the drawing – and a cylindrical outer surface which, relative to the motor axis M, is rounded with a radius Rmo corresponding to half the rotor diameter D3, as shown in Figure 9 This is recognizable. The magnet width MW corresponds to the dimension of the flat inner surface in the circumferential direction, and the magnet thickness MT is measured perpendicular to this inner surface.

[0125] The execution according to Figure 10a indicates how the execution according to Figure 10c The outer surface is cylindrical with a radius Rmo. Additionally, the inner surface is cylindrically rounded, coaxial with the outer surface, and has a correspondingly smaller inner radius Rmi. The side surfaces are parallel to each other and to a diameter and extend radially over the magnet thickness MT, such that the smaller inner radius Rmi is approximately equal to the radius Rmo minus the magnet thickness MT. Reference symbol list

[0126] 1 Brake device 100 Drive carrier 101 Recess 102 Opening 103 Projection 104 Support surface 105 Inner surface 106 O-ring 1100 Stator 1101 Stator body 1102 Stator tooth 1103 Stator winding 1104 Air gap 1105 Stator gap 1200 Rotor 1201 Rotor magnet 2 Brake disc 3 Brake caliper 31, 32 Brake pad 33 Mounting bolt 4 Brake actuator 41, 42 Motor (actuator) 411, 421 Motor shaft 412, 422 Gear 413, 423 Motor housing 414, 424 Bearing cover 415, 425 Extension 416, 426 Flange element 417,427 Screw 418 Mounting opening 418a Shoulder section 419 Bearing 43 Thrust bearing 44 Pressure piece 5 Actuating device 6 First actuating drive 61 Cam disc 62 Cam disc (integrated with spindle nut 72) 63 Ball 64 Raceway 65 Gear 66 Ball cage 67 Recess 7 Second actuating drive 71 Threaded spindle 72 Spindle nut (integrated with cam disc 62) 73 Driver 74 Hub part 75 Gear 76 Slot 8 Friction element 81 Friction surface 9 Counter friction element 91 Counter friction surface 92 Driver 93 Spring element A Axis R Wheel axle V Direction of adjustment L Air gap M Motor axle D1 Stator outer diameter D2 Stator inner diameter D3 Rotor diameter TW Tooth width AG Air gap width SO Gap width BetaP Pole angle BetaM Magnet angle MWMagnet width MTMagnet thickness MDMagnet spacing,

Claims

1. An electromechanical braking device (1) for a motor vehicle, comprising an actuating device (5) which is driveable by at least one electric motor (41, 42) and by which a brake part (22) is adjustable, wherein the motor (41, 42) has a stator (1100) and a rotor (1200), which is rotatably mounted therein about a motor axis (M), wherein the stator (1100) has a stator outer diameter (D1) and a stator inner diameter (D2) and comprises a number of radially inwardly projecting stator teeth (1102) which are distributed over the circumference, have a tooth width (TW) in the circumferential direction on the stator outer diameter (D1), which corresponds to a tooth root width in a radially outer external region of a stator tooth facing away from the air gap, and between which a stator gap (1105) having a gap width (SO) is in each case located on the stator inner diameter (D2), and which support at least one stator winding (1103), wherein the rotor (1200) comprises a rotor diameter (D3) and a number of poles, which deviate from the number of stator teeth (1102), of rotor magnets (1201) which are distributed over the outer circumference, are in the form of permanent magnets and extend in a rod-shaped manner axially parallel to the rotor axis (M), and which each have a radial magnet thickness (MT) and a magnet width (MW) in the circumferential direction, and the rotor magnets (1201) are each spaced apart at a magnet distance (MD) on the rotor diameter (D3), wherein the radial air gap (1104) having a radial air gap width (AG) is formed between the stator teeth (1102) and the rotor magnets (1201), wherein the ratio of the magnet thickness (MT) to the magnet width (MW) is between 0.4 and 0.55, wherein the ratio of the air gap width (AG) to the tooth width (TW) is between 0.45 and 0.65.

2. The braking device as claimed in claim 1, characterized in that the rotor magnets (1201) have a rectangular cross section.

3. The braking device as claimed in claim 1, characterized in that the rotor magnets (1201) have a cross section in the shape of a hollow cylinder segment.

4. The braking device as claimed in claim 1, characterized in that the rotor magnets (1201) have a segment-like cross section with a cylindrical outer surface and a flat inner surface.

5. The braking device as claimed in any one of the preceding claims, characterized in that a rotor magnet (1201) extends on the rotor diameter (D3) with respect to the motor axis (M) over a magnetic angle section (BetaM), and the rotor magnets (1201) are offset relative to one another by a pole angle (BetaP), the ratio of the magnetic angle section (BetaM) to the pole angle (BetaP) being between 0.65 and 0.75.

6. The braking device as claimed in any one of the preceding claims, characterized in that the number of poles is 8 or 10.

7. The braking device as claimed in any one of the preceding claims, characterized in that the number of stator teeth (1102) is 6 or 9.

8. The braking device as claimed in any one of the preceding claims, characterized in that the ratio of the air gap width (AG) to the tooth width (TW) is between 0.5 and 0.6.

9. The braking device as claimed in any one of the preceding claims, characterized in that the ratio between the stator outer diameter (D2) and the stator inner diameter (D1) is between 0.55 and 0.65.

10. The braking device as claimed in any one of the preceding claims, characterized in that the ratio between the air gap width (AG) and the stator inner diameter (D2) is between 0.015 and 0.03.

11. The braking device as claimed in any one of the preceding claims, characterized in that the braking device has at least two motors (41, 42).