EP4200536A1

EP4200536A1 - Brake device

Info

Publication number: EP4200536A1
Application number: EP21762589.6A
Authority: EP
Inventors: Michael Putz; Thomas Zipper
Original assignee: Stop In Time GmbH
Current assignee: Stop In Time GmbH
Priority date: 2020-08-24
Filing date: 2021-08-24
Publication date: 2023-06-28
Also published as: CA3190939A1; US20230392659A1; CA3190936A1; WO2022040713A1; EP4200537A1; WO2022040712A1; MX2023002310A; MX2023002309A; CN116420034A

Abstract

The invention relates to a brake device and a machine which comprises said brake device, wherein: the brake device comprises an actuator (04), a transmission, an expander device, a brake lining (063) and a frictional surface; the actuator (04) moves in a limited actuator actuation region; in at least part of its actuator actuation region, the actuator (04) turns the expander device about at least one fulcrum via the transmission; in at least part of its actuator actuation region, for the purpose of braking, the actuator (04) presses the brake lining (063), via the expander device, in the direction of and onto the frictional surface to generate a pressing force and a braking torque resulting therefrom; the transmission has a non-linearity (03), i.e. a non-consistent transmission ratio, over at least part of the actuator actuation region; and the transmission turns the expander device in accordance with the non-linearity (03).

Description

braking device

The invention relates to a braking device and a machine according to the preambles of the independent patent claims.

Different brakes with spreading devices are known from the prior art. For example, brakes are known in which the pressed-on parts, in particular the brake lining, are guided along a straight line and in which the spreading device has a special geometry, whereby the spreading device rolls off the pressed-on parts when it rotates. What is disadvantageous about such brakes, however, is that the required geometry of the spreading device has disadvantages both mechanically and in terms of production engineering and therefore cannot be produced efficiently and cost-effectively. Furthermore, the durability of such spreading devices is limited due to the special geometry.

The object of the invention is to overcome the disadvantages of the prior art. In particular, the object of the invention is to create a braking device with a spreading device that enables efficient operation of the braking device, has a long service life and can be produced easily and efficiently. Furthermore, it may be the object of the invention to provide a braking device which enables the use of a spreading device with a conventional geometry. The object according to the invention is achieved in particular by the features of the independent patent claims.

The invention relates in particular to a braking device, the braking device comprising an actuator, in particular an electric actuator, a gear, a spreading device, a brake pad and a friction surface.

Provision is preferably made for the actuator to move within a limited range of actuator activation.

It is preferably provided that the actuator rotates and/or moves the spreading device about at least one pivot point in at least a part of its actuator actuation region via the transmission.

Provision is preferably made for the actuator to press the brake lining in the direction of and/or against the friction surface at least in part of its actuator actuation region via the spreading device.

Provision is preferably made for the actuator to press the brake lining in the direction of and/or against the friction surface at least in part of its actuator actuation area via the spreading device for braking in order to generate a contact pressure force and a braking torque resulting therefrom.

In other words, the spreading device can be moved or rotated by the actuator in such a way that the spreading device presses the brake pad in at least part of the actuator actuation area for braking in order to generate a contact pressure force and a braking torque resulting therefrom in the direction of and against the friction surface.

This rotation and/or movement of the spreading device can result in a covering lift. In the context of the present invention, a pad stroke can be understood to mean that the brake pad is targeted, particularly in the direction of the Friction surface is moved. In other words, a pad lift can also be relevant to the braking effect.

Within the scope of the present invention, a lining stroke relevant to the braking effect can be understood to mean a lining stroke through which the brake lining is moved, in particular in the direction of the friction surface, in particular the friction surface.

If necessary, it is provided that the actuator brings about a lining travel, in particular a braking effect-relevant one, at least in a part of its actuator actuation range via the transmission.

Provision is preferably made for the transmission to have a non-linearity, ie a ratio that is not constant over at least part of the actuator actuation range.

It is preferably provided that the transmission rotates and/or moves the spreading device in accordance with the non-linearity.

The spreading device can be rotated and/or moved by the actuator relative to the brake pad, the parts of the braking device pressing the brake pad, the actuator and/or the, in particular stationary, transmission parts.

The braking device can be designed electromechanically.

If necessary, it is provided that when the actuator is moved, the gearing and, if necessary, the spreading device are actuated. As a further consequence, it can be provided that the actuation of the transmission and possibly the spreading device causes a lining stroke to be carried out and in particular the brake lining to carry out a movement.

If necessary, the transmission or at least a part of the transmission is designed or configured non-linearly. In particular, the transmission includes at least one non-linearity. The transmission can include several transmission parts. In particular, the transmission can include at least one gear transmission and/or at least one transmission, which in particular has at least one non-linear transmission ratio that can be changed over the actuation path. Furthermore, the transmission can include at least one gear ratio for driving or not driving different parts.

The movement of the actuator can optionally be non-linearly related to the resulting movement of the brake pad, in particular the pad stroke. It is also possible that a movement of the actuator in some areas does not produce any pad travel.

In the context of the present invention, the terms “no lining travel” and/or “free lining travel” can be understood to mean that there is no significant change in the braking effect and/or the overcoming of the air gap, but movements within the scope of manufacturing tolerances or mechanical characteristics, for example, may not be excluded are. In particular, it can be provided that at the beginning and at the end of the limited actuator actuation range, ie in particular at the beginning and at the end of the range of actuator movement, the movement of the actuator causes no pad lift and/or is pad lift-free.

If necessary, it is provided that the transmission is adapted in some areas on the basis of different requirements for the braking device, such as moderate deceleration, full braking, continuous braking and/or the like, as well as internal functionalities. In other words, the transmission, in particular the non-linearities, can be optimized for the operating states that occur during the actuation of an electromechanical braking device.

If necessary, it is provided that this adjustment and/or optimization of the transmission takes place with the overriding goal of the highest possible functional safety of the braking device and the entire braking system. In other words, this adjustment and/or optimization of the transmission should not be based on individual components, such as the electric actuator. If necessary, it is provided that at least two areas of the transmission are differently optimized and/or adapted with pad travel, in particular one that is relevant to the braking effect.

If necessary, it is provided that at least two areas of the transmission with lining travel, in particular one that is relevant to the braking effect, have two different non-linearities.

Within the scope of the present invention, a transport device can be understood to mean any device and/or machine that can be used to drive and/or that can be used to transport people and/or loads while driving.

If necessary, it is provided that the gear ratio of the transmission is selected and/or configured in such a way that at least one partial section with a non-linearity is formed, provided and/or arranged along the actuator actuation area.

If necessary, it is provided that the gear ratio of the transmission is selected and/or designed in such a way that two, three, four, five, six, seven, eight, nine, ten or more sections with differently acting non-linearities are formed, provided and/or along the actuator actuation area. or arranged.

Within the scope of the present invention, brake can be understood to mean the braking device.

Within the scope of the present invention, a rotated contact surface can be understood to mean a contact surface of the spreading device, with the spreading device and the rotated contact surface possibly rotating. Furthermore, in the context of the present invention, the contact surface can also be understood as the rotated contact surface.

In the context of the present invention, a non-rotated contact surface can be understood to mean a contact surface of a component of the braking device, which is different from the expanding device. Furthermore, within the scope of the present Invention, under abutment surface, the non-rotated contact surface can also be understood.

In the context of the present invention, the spreading part can be understood to mean the spreading device, in particular also together with the at least one rotated contact surface and/or with the at least one non-rotated contact surface.

Within the scope of the present invention, the actuator rotation range can be understood to mean the actuator actuation range.

In the context of the present invention, EMB can be understood to mean the brake device, in particular an electromechanical one, and/or the brake, in particular an electromechanical brake.

If necessary, it is provided that the spreading device is at least partially surrounded by the braking device, in particular the transmission, so that the spreading device cannot fall out of the braking device.

If necessary, it is provided that the spreading device is arranged loosely in the braking device.

If necessary, it is provided that the spreading device is arranged in the braking device.

If necessary, it is provided that in at least one part of the actuator actuation area, in particular in a first actuation point of the actuator or first actuation area of the actuator, the spreading device relative to the brake pad, the parts of the braking device pressing the brake pad, the actuator and/or the, in particular stationary, transmission parts, performs a relative movement. If necessary, it is provided that the relative movement of the spreading device takes place, if necessary, in particular exclusively, along or in the plane of rotation of the spreading device.

If necessary, it is provided that the relative movement of the spreading device occurs if necessary, in particular exclusively, essentially normal to the direction of rotation, in particular the pressing direction, of the spreading device.

If necessary, it is provided that the relative movement of the spreading device occurs if necessary, in particular exclusively, in at least one direction of extension, preferably in the longitudinal direction and/or transverse direction, of the spreading device.

If necessary, it is provided that the relative movement of the spreading device takes place in all directions, in particular in all directions of extension of the spreading device.

The at least one rotated and the at least one non-rotated rolling surface, in particular one rotated and the at least one non-rotated contact surface, can have any initial position, e.g. due to weight or also randomly due to e.g. vibrations. They can also be frictionally engaged, or largely frictionally engaged, and in doing so make no appreciable or an appreciable relative movement in the transverse direction. The frictional connection can also be overstrained and thus a sliding compensating movement between the at least one rotated and the at least one non-rotated rolling surface can occur, and a mixed form between sliding and rolling can also occur. An additional relative movement in the transverse direction can also occur and vibrations can also be superimposed on the movements and/or the relative movement in the transverse direction can be used up in the freedom of movement and thus cause the at least one rotated rolling surface to slide on the non-rotated rolling surface.

In addition, the movement of rotated and non-rotated rolling surfaces can also follow geometric changes or deformations. If necessary, it is provided that the spreading device has at least one, in particular rotated, contact surface.

If necessary, it is provided that the braking device, in particular the transmission and/or the parts of the braking device that press on the brake lining, comprise at least one abutment surface, in particular a non-rotated pressing surface.

If necessary, it is provided that the at least one contact surface is pressed against the at least one abutment surface in at least one part of the actuator actuation area, as a result of which the spreading device rotates and/or moves if necessary.

If necessary, it is provided that the at least one rotated contact surface, in particular the contact surface, is pressed in at least one part of the actuator actuation area by rotating the spreading device against the at least one non-rotated contact surface, in particular the abutment surface, and that a contact force arises between the contact surface pairs that are present.

If necessary, it is provided that the contact surface, in particular the rotated contact surface, and/or the abutment surface, in particular the non-rotated contact surface, are designed in such a way that these surfaces, in particular during the rotation and/or movement of the spreading device, have a, in particular sliding, and/or rolling relative movement.

If necessary, it is provided that the braking device is designed in such a way that the brake pad travels a path of movement that deviates from a straight line when it is pressed on.

If necessary, it is provided that the contact surface and the abutment surface are designed in such a way that the brake pad travels a path of movement that deviates from a straight line when it is pressed on. This movement path is optionally defined by the interaction of the gear and/or the spreading device with the brake lining.

In the context of the present invention, a rolling relative movement can be understood to mean that the rotated contact surface executes a rolling movement on the non-rotated contact surface like a wheel on a substrate. As a result of the frictional connection and/or the static friction, the surfaces may have essentially the same surface speeds, as a result of which the rolling takes place with particularly little slippage. If the friction capacity and/or the static friction capacity is exceeded, rolling can change to a gliding movement with increased slip, possibly even to the point of the behavior of a locked wheel on a surface, which is referred to as gliding. In between, transition areas are also possible. The theoretical ideal goal would be to achieve a geometry that essentially, especially exclusively, allows a rolling movement, especially if, as with the braking device, there can be high forces on small parts and thus high surface pressure. In other words, the spreader device can be designed in such a way that the geometry of the spreader device provides rolling movement wherever possible, even if a linear guide guides the movement of the pressed part.

In the case of the braking device, this geometry, which enables the so-called ideal rolling behavior, can only be pursued to a limited extent or not at all in favor of other advantages, such as the cheapest possible manufacturability, the use of well-rounded parts with suitable surface hardness and quality, the avoidance of unfavorable production methods such as milling curves, etc will. A straight-line or other guidance can also be dispensed with in the braking device, and instead a compensating movement transverse to the pressing direction can be permitted, with which the rolling state can be promoted due to the lack of forced guidance.

The movements brought about by the spreading device are, on the one hand, the desired ones in the pressing direction and, on the other hand, those with a different movement component, which can also be essentially normal (here also referred to as transverse) to the pressing direction, but spatially preferably lies in the plane of the spreader mechanism rotation. In the context of the present invention, transverse can sometimes also be referred to as height, according to “above” in figures and a frequent installation position of brakes. A deviation from the desired pressing direction is also referred to as a height error. The transverse movement can be prevented by a guide, for example by a straight guide. However, it can also be made possible, for example by playing with the leadership or by forgoing effective leadership. The lateral movement can also occur as a smoothing movement rather than a rolling movement, especially when a guide forces it to do so.

These movements can be caused by the movement of the spreading device, but they can also occur independently, for example if they are triggered by vibrations. Even with a rolling movement, the point of contact (point, line, surface) between the rotated contact surface and non-rotated contact surface can migrate transversely to the contact direction during the contact pressure process.

If necessary, it is provided that the actuator rotates the spreading device about a first pivot point via the gearing in at least part of its actuator operating range, in particular in a second operating point of the actuator or second operating range of the actuator.

If necessary, it is provided that the actuator rotates the spreading device about a further pivot point via the gearing in at least a part of its actuator operating range, in particular in a further operating point of the actuator or further operating range of the actuator.

If necessary, it is provided that the position of at least two pivot points deviates from one another and/or differs.

If necessary, it is provided that the position of the pivot points is limited by the design of the braking device. If necessary, it is provided that the braking device is designed in such a way that the displacement of the pivot point of at least two pivot points of the spreading device is opposed by an elastic resistance, in particular a resistance device.

If necessary, it is provided that at least one pivot point is mounted and/or freely movable, in particular not mounted.

In the context of the invention, a mounted pivot point can be understood to mean that the mounted pivot point is arranged in a fixed manner, in particular with no degree of freedom of movement, with respect to the brake pad, the parts of the braking device pressing the brake pad, the actuator and/or the, in particular stationary, transmission parts.

In the context of the invention, an unsupported pivot point can be understood to mean that the unsupported pivot point is freely movable relative to the brake pad, the parts of the braking device pressing the brake pad, the actuator and/or the, in particular stationary, transmission parts, and in particular at least one freedom of movement relative to these parts having.

If necessary, it is provided that the actuator in at least a part of its actuator actuation area, in particular in a third actuation point of the actuator or third actuation area of the actuator, the spreading device via the transmission relative to the brake lining, the brake lining pressing parts of the braking device, the actuator and / or the, in particular fixed gear parts, is fixed and/or free of relative movement.

If necessary, it is provided that the spreading device comprises at least two spreading device parts, with at least one spreading device part possibly being a pin, a spigot and/or a prefabricated part.

If necessary, it is provided that the at least one contact surface of the spreading device is formed at least partially from a spreading device part. If necessary, it is provided that the at least one contact surface of the spreading device is at least partially arranged on a spreading device part.

If necessary, it is provided that the spreading device parts are connected to one another, in particular in a non-positive, materially bonded, pressed and/or welded manner.

The spreading device can comprise at least two spreading device parts, in particular at least one spreading device holder and at least one spreading device roller arranged thereon. One part of the spreading device, in particular the spreading device roller, can be a pin, in particular a cylindrical one, or a pin, in particular a cylindrical pin.

One part of the spreading device, in particular the spreading device roller, can be connected in a force-fitting and/or materially bonded manner, in particular pressed and/or welded, to the other spreading device part, in particular the spreading device holder.

At least one spreader part, in particular the spreader roller, can be a cylindrical pin with a diameter of 6 to 10 mm inclusive, in particular 8 mm.

The spreading device can be designed as a cam or lever.

If necessary, it is provided that the spreading device is non-linear.

If necessary, it is provided that the spreading device is rotated by the actuator via the gear through a limited range of rotation.

In particular, it can be provided that the spreading device is rotated by the actuator via the gearing in a limited range of rotation. Under In the context of the present invention, the range of rotation can be understood as the range of angles by which the spreading device is rotated.

The cam or lever of the spreader may be non-linear. At least one non-linearity can be arranged on the cam or the lever of the spreader.

If necessary, it is provided that the spreading device has at least one non-linearity, that is to say a translation that is not constant over at least part of the rotation range.

In the context of the present invention, a non-linearity can be understood to mean a non-linear translation.

If necessary, it is provided that the at least one non-linearity of the spreading device is matched to the at least one non-linearity of the transmission.

If necessary, it is provided that the at least one non-linearity, in particular the non-linear translation effect, of the spreading device is taken into account in the configuration of at least one non-linearity, in particular the non-linear translation, of the transmission.

If necessary, provision is made for the actuator to be operated in at least a partial area of its actuator activation area at an operating point that deviates from the optimum operating point of the actuator.

If necessary, provision is made for the actuator to be operated in at least a sub-area of its actuator actuation area at an operating point that deviates from an operating point of maximum power of the actuator.

If necessary, it is provided that the gear, in particular the spreading device, starting from a first position, in particular a zero position, of the gear carries out or converts a movement of the actuator in a first direction for braking. If necessary, it is provided that the transmission, in particular the spreading device, starting from a first position, in particular a zero position, for adjusting an air gap, in particular for actuating a wear adjustment device and/or a wear adjustment device, a movement of the actuator in one direction, in particular the opposite direction to the first, executes or implements the second direction.

If necessary, it is provided that at least part of the actuator rotates once in a first direction of rotation and once in a second direction of rotation. The second direction of rotation can be opposite to the first direction.

The transmission, in particular the spreading device, can optionally convert the first direction of rotation of the actuator into a movement in the first direction. The transmission, in particular the spreading device, can optionally convert the second direction of rotation of the actuator into a movement in the second direction.

The zero position of the gear can be defined geometrically and/or mechanically by the gear, in particular the non-linearities. Within the scope of the invention, the zero position of the transmission can therefore be understood to be that position from which actuation of the actuator in a first direction causes a lining lift. The zero position of the transmission can also be determined, among other things, by the geometry of the transmission, in particular the beginning of the gradient.

If necessary, the actuator can be brought into a rest position, in particular starting from the zero position of the transmission with pad travel and without any braking effect. If necessary, the actuator can be moved from the rest position in the direction of a first direction to overcome the air gap and/or to increase the braking effect and/or in the direction of a second actuation direction in order to fulfill other tasks.

The rest position of the transmission can be a position of the transmission in which the air gap has a defined size. If necessary, the rest position can be identical to the zero position. If necessary, it is provided that a wear adjustment device is provided in the pivot point of the spreading device.

If necessary, it is provided that the spreading device includes a drive.

If necessary, it is provided that a wear adjustment device is provided in the drive of the spreading device.

In particular, it is optionally provided for wear adjustment to change and/or adjust the angle between the spreading device and the transmission, in particular at least one non-linearity of the transmission.

If necessary, this change and/or adjustment takes place via an adjustment device such as, in particular, a gearing. In particular, the spreading device can be changed and/or adjusted in relation to the transmission, in particular in relation to at least one non-linearity of the transmission, by means of the adjustment device.

If necessary, it is provided that a wear adjustment device is provided between the actuator and the transmission or between the transmission and the spreading device.

In particular, a holder for holding the actuator can be provided. If necessary, a wear adjustment device is arranged between the holder of the actuator and the actuator.

If necessary, it is provided that the transmission includes a wear adjustment device for wear adjustment.

If necessary, it is provided that the braking device comprises a wear adjustment device, which is actuated, in particular exclusively, by the actuator, the transmission and/or the spreading device. If necessary, it is provided that the braking device is set up for manual wear adjustment

The wear adjustment device may be a ratchet device and/or a worm gear device.

If necessary, it is provided that the actuator, the gear and/or the spreading device is set up for braking and wear adjustment, in particular for actuating a wear adjustment device.

If necessary, it is provided that the braking device comprises only a single actuator for braking and for wear adjustment, in particular for actuating a wear adjustment device.

If necessary, it is provided that the actuator comprises several parts.

If necessary, it is provided that the actuator comprises a spring and an electric motor, with the spring and the electric motor possibly being independent of one another in terms of components and/or direction of action.

If necessary, it is provided that the spring interacts with the electric motor via at least one further component and/or via the gear mechanism.

If necessary, it is provided that the actuator comprises two electric motors.

If necessary, it is provided that the braking device interacts with at least one electrical machine, in particular at least one electromagnetically excited electrical machine.

If necessary, it is provided that the transmission includes kinematic devices. If necessary, it is provided that the gearing comprises a cam, a ball ramp and/or a lever.

If necessary, it is provided that the translation of the transmission, in particular in braking operation, is variable.

If necessary, it is provided that the transmission ratio of the transmission can be changed, in particular actively, preferably by turning a ratchet.

If necessary, it is provided that the transmission ratio of the transmission can be changed, in particular passively, preferably by spring-loaded retraction of components or the elastic deformation of components.

In the context of the present invention, braking operation can be understood to mean the period of time between putting the braking device into operation and switching it off, during which the braking device is ready to accept and implement braking commands. In other words, the braking device can be ready to brake in braking mode.

If necessary, it is provided that the transmission is selected and/or designed in such a way that at least one partial section with a non-linearity is formed and/or arranged along the actuator actuation area.

If necessary, it is provided that the transmission is selected and/or designed in such a way that at least two subsections with differently acting non-linearities are formed and/or arranged along the actuator actuation area.

If necessary, it is provided that the at least one non-linearity is selected from the following non-linearities: non-linearity to overcome an air gap between the brake pad and friction surface, non-linearity to determine the point of contact of the friction surface and the brake pad, non-linearity to achieve a minimum braking effect, non-linearity to generate an increasing braking torque, non-linearity for operation with lowered electric Power requirement, non-linearity to quickly achieve high braking effects, non-linearity to measure and/or set parameters, non-linearity to reduce electrical and mechanical loads when starting the lining stroke, non-linearity to compensate for brake fading, non-linearity to adjust wear.

In particular, the invention relates to a transportation device, a machine, a vehicle, an elevator and/or a bicycle, which comprises an electromechanical brake according to the invention.

Optionally, the invention relates to a part of a transport device or a part of a machine, such as in particular a propeller shaft, which comprises an electromechanical brake according to the invention or is formed from an electromechanical brake according to the invention.

If necessary, it is provided that the machine, in particular the conveying device, comprises a further, in particular electronic, braking device.

If necessary, it is provided that the further braking device is designed as a parking brake, in particular as a spring-loaded parking brake.

In particular, the invention relates to a method for operating a braking device according to the invention.

If necessary, it is provided that the gear and/or the spreading device converts only a part of the movement of the actuator, in particular only a part of the actuator actuation area, into a covering stroke.

If necessary, it is provided that the actuator is moved in the first and second direction before and/or after the part of the actuator actuation area relevant to the lining stroke via the transmission and/or the spreading device, without generating a lining stroke relevant to the braking effect. If necessary, it is provided that the translation of the transmission is selected and/or designed such that, starting from a first position, in particular the zero position, of the transmission, the non-linearities are arranged along the movement of the actuator, in particular the pad stroke, in the first direction.

If necessary, it is provided that at least two non-linearities are arranged along the first direction according to the order given below: non-linearity to reduce electrical and mechanical loads when the lining starts to move, non-linearity to overcome the air gap between the brake lining and the friction surface, non-linearity to determine the contact point of the friction surface and the brake lining, non-linearity to achieve a minimum braking effect, non-linearity to operate with reduced electrical power requirements, non-linearity to quickly achieve high braking effects, non-linearity to generate an increasing braking torque, with the braking torque being adapted to the respective braking dynamics, non-linearity to compensate for brake fading.

If necessary, it is provided that the above non-linearities are arranged one after the other along the first direction on the transmission. In particular, the above non-linearities can be step through and/or run through one after the other during the movement of the actuator.

If necessary, it is provided that the non-linearities are arranged in any order along the first direction.

If necessary, it is provided that the above non-linearities are arranged in any desired order on the transmission along the first direction.

If necessary, it is provided that the gear ratio of the transmission is selected and/or designed in such a way that, starting from the first position, in particular the zero position, of the transmission along the movement of the actuator in the second direction the non-linearity for measuring and/or setting parameters and/or the non-linearity for wear adjustment are arranged.

If necessary, it is provided that the non-linearity for measuring and/or adjusting parameters and/or the non-linearity for wear adjustment are arranged on the transmission along the second direction one after the other. In particular, the non-linearity for measuring and/or setting parameters and/or the non-linearity for wear adjustment during the movement of the actuator can be successively passed through and/or passed through.

If necessary, it is provided that the non-linearity for measuring and/or adjusting parameters is configured for measuring mechanical losses, the zero position of the gear, the zero position of the actuator position and/or at least one spring action.

If necessary, it is provided that the non-linearity for measuring and/or setting parameters is designed in such a way that the actuator is moved in its first direction, starting from the zero position of the transmission.

If necessary, it is provided that at least one parameter of the braking device, in particular engine losses, transmission losses, mechanical losses and/or the effect of any existing springs, is measured by the movement of the actuator in its first direction.

If necessary, it is provided that the moment of the actuator that arises and/or results from the movement is recorded.

If necessary, it is provided that the assessment as to whether an adjustment of the braking device is necessary, based on a comparison of at least one parameter of the braking device, in particular the torque of the actuator, with expected values and/or with measured values of the torque of the actuator at other operating points and/or in other operating states. If necessary, it is provided that the non-linearity for measuring and/or setting parameters is designed in such a way that the actuator is moved in its second direction, starting from the zero position of the transmission.

If necessary, it is provided that a force measuring device, in particular a spring and/or a stop, is provided in the second direction, against which at least part of the transmission, in particular the actuator, is in contact, whereby the zero position of the actuator position can be measured and/or set, if necessary .

If necessary, it is provided that the at least one parameter of the braking device takes place by comparing the torque, the motor current and/or the motor voltage in normal operation and the torque, the motor current and/or the motor voltage in measurement operation.

If necessary, it is provided that the non-linearity is designed to reduce electrical and mechanical loads when the pad lift starts in such a way that the transmission ratio of the transmission of this non-linearity in the first half of the air gap is more than twice as large as the speed transmission in the second half of the air gap.

If necessary, it is provided that the non-linearity to reduce electrical and mechanical loads at the start of the lining stroke is designed in such a way that the transmission ratio, in particular the speed transmission, of this non-linearity, preferably the ratio between the speed of the actuator and the speed of the lining stroke, in the first half of the Air gap, especially in the first half of the way to overcome the air gap, is more than twice as large as the speed translation in the second half of the air gap.

If necessary, it is provided that the non-linearity for overcoming the air gap between the brake pad and the friction surface is designed such that the transmission ratio of the transmission of this non-linearity is less than half the maximum over more than half of the air gap Speed ratio in the lining stroke area following the air gap, so that the air gap may be overcome more quickly than in normal operation.

If necessary, it is provided that the non-linearity for overcoming the air gap between the brake lining and the friction surface is designed in such a way that the transmission ratio of the transmission, in particular the speed transmission of this non-linearity, preferably the ratio between the speed of the actuator and the speed of the lining stroke, is more than half of the air gap, in particular more than half of the way to overcome the air gap, is less than half as large as the maximum speed ratio in the lining stroke area adjoining the air gap, so that the air gap may be overcome more quickly than in normal operation.

If necessary, it is provided that the non-linearity for overcoming the air gap between the brake lining and the friction surface is designed in such a way that the actuator is operated with the maximum actuator power, as a result of which the air gap is overcome as quickly as possible.

If necessary, it is provided that the non-linearity for overcoming the air gap between the brake lining and the friction surface is designed in such a way that the air gap is overcome as quickly as possible, in that a device, in particular a cam or a ramp, has a gradient which is designed in such a way that, if necessary, at the beginning of the lining stroke, starting current peaks and starting current loads can be avoided and/or reduced.

If necessary, it is provided that the non-linearity for determining the point of contact of the friction surface and the brake pad is designed in such a way that the point of contact of the brake pad and the friction surface, in particular from the energy, current and/or power consumption of the actuator and/or from the course of the Actuator load, especially the moment, is recognizable. If necessary, it is provided that non-linearity for determining the point of contact of the friction surface and the brake pad can be used to check whether an adjustment of the braking device, in particular an adjustment of the brake pad and/or an adjustment of the air gap, is necessary.

If necessary, it is provided that the translation of the transmission of the non-linearity for determining the point of contact of the friction surface and the brake lining, in the possible area of the point of contact of the brake lining and the friction surface, an evaluable combination of transmission ratio and actuator torque, in particular an interpretable curve from the energy, current - And / or power consumption of the actuator generated.

If necessary, it is provided that the combination of transmission ratio and actuator torque that can be evaluated is an interpretable curve from the energy, current and/or power consumption of the actuator, the actuator load and/or the actuator torque via the actuation, in particular taking into account the respective transmission ratio.

If necessary, it is provided that in the area of non-linearity for determining the point of contact of the friction surface and the brake pad, there is a significant difference to the behavior in the air gap from the point of contact between the friction surface and the brake pad.

If necessary, it is provided that the non-linearity to achieve a minimum braking effect is designed in such a way that a specific required minimum braking effect, in particular in the event of full braking, is achieved within a minimum time of action, with the minimum time of action only being a maximum of 20% above the time which, in particular, to achieve the Minimum braking effect that is technically possible with the braking device.

If necessary, it is provided that the non-linearity for generating an increasing braking torque, with the braking torque possibly being adapted to the braking dynamics, is designed in such a way that the speed of the Braking torque build-up is adapted to the resulting dynamic weight shift of the vehicle, so that any locking of the wheels of the vehicle is counteracted.

If necessary, it is provided that the non-linearity for operation with a reduced electrical power requirement is designed in such a way that the power consumption of the actuator when the transmission is operated at low speeds and/or when the actuator is at a standstill is at least 20% lower than in comparison to non-linearity, which is designed in particular according to the criterion of the maximum achievable engine output power, for the same or a similar operation and/or operating point, in particular for operation at low speeds and/or when the actuator is at a standstill, so that the power consumption of the actuator is reduced, in particular during longer continuous braking is.

If necessary, it is provided that the translation of the transmission is selected and/or designed in such a way that, starting from the first position, in particular the zero position, of the transmission along the movement of the actuator, in particular the lining stroke, in the first direction, the non-linearity for operation with lowered electrical power requirement is arranged so that in operating states that have a long holding time and/or a high temperature load, there is a low consumption of electrical energy and/or low heat loss of the actuator, in particular the electrical actuator.

If necessary, it is provided that the non-linearity to compensate for Bremsfadinq is designed in such a way that the actuator is operated with an engine torque that is higher under the same operating conditions, in particular the operating temperature, in particular higher than the maximum permissible engine torque and / or higher than the maximum permissible shaft power than that in the case of non-linearity, which is designed according to the criterion of the maximum achievable engine output power, so that a braking effect is also achieved in the event of brake fading. If necessary, it is provided that at least one non-linearity, in particular over the lining travel, is designed to compensate for air gap errors in such a way that an air gap error, in particular a deviation in the size of the air gap from the assumed size, is compensated for, with the air gap error preferably being caused by wear.

If necessary, provision is made for the braking device to be operated up to a certain deviation in the size of the air gap error, in particular by adapting the movement of the actuator, preferably without wear adjustment and/or without a wear adjustment device.

If necessary, it is provided that the non-linearity for wear adjustment is designed in such a way that the actuator, starting in particular from the zero position of the transmission, carries out a movement counter to the direction of movement or direction of rotation used for braking, in particular a movement in the second direction, and that this movement of the actuator, in particular without a braking effect, the wear adjustment device is actuated.

If necessary, it is provided that the non-linearity for the wear adjustment is designed in such a way that the actuator performs a movement in the direction of braking, in particular a movement in the first direction, so that the wear adjustment device is actuated by this movement of the actuator, by possibly after reaching a for the braking, in particular for the parking brake, required maximum position of the actuator leads to a further movement of the actuator, in particular without a functional lining stroke, for the actuation of the wear adjustment device or prepares it.

If necessary, it is provided that the non-linearity for quickly achieving high braking effects is designed such that the actuator is operated with an engine torque that is the same as the maximum permissible engine torque and/or is the same as the maximum permissible shaft power. If necessary, it is provided that at least one actuator position of the actuator is maintained with a reduced, in particular very low, electrical power requirement or without current by appropriate design of at least one non-linearity and optionally by the interaction of this at least one non-linearity with a spring, in particular a spring effect.

If necessary, it is provided that the effective range of at least one non-linearity and/or one non-linearly acting component is distributed over several, in particular non-linearly designed and/or non-linearly acting, parts of the transmission, in particular several transmission components, preferably cams and/or ball ramps twisted against one another .

The effective range of at least one non-linearity and/or a non-linear component, in particular the effective range and/or the design of the transmission components, can each be assigned to a specific actuator actuation range.

By using additional non-linearly acting components, it may be possible to enlarge and/or increase the actuator actuation range that is predetermined and/or limited by the non-linearity of the individual components. In particular, the effective range of the existing non-linearities, preferably the actuator actuation range limited by the actuation range and/or range of motion of the transmission components, can thereby be enlarged and/or increased.

If necessary, it is provided that a first transmission component, in particular a first non-linearity of the first transmission component, is assigned to a first actuator actuation region. In order to be able to increase the range of motion and/or the range of actuation, a second transmission component can be provided, which is assigned to a second actuator range of actuation. This second transmission component can have a further part of the first non-linearity and/or a second non-linearity. The second actuator actuation area can connect to the first actuator actuation area. If necessary, it is provided that the gear ratio of the transmission is selected and/or designed in such a way that an actuator movement without braking effect causes a movement of brake components, such as in particular the brake pad carrier.

If necessary, it is provided that this movement causes no and/or only a minimized residual grinding torque.

If necessary, it is provided that a movement of the brake components, such as in particular the brake pad carrier, is caused by an actuator movement without braking effect, i.e. without braking effect, in such a way that no and/or only a minimized residual grinding torque remains, which is known under the term “zero drag”. .

Statements by the inventor are given below, which are intended to enable a better understanding of the invention. The features described below can but do not have to be features of the braking device according to the invention. The braking device according to the invention can include and/or have the features listed individually or in combination, ie in any combination.

"Actuate" can be understood as the process for more braking effect and "release" as the process for less. An actuating mechanism can fulfill both tasks.

A “ratchet” can be understood as any device or effect that specifies one direction, eg direction of rotation, or prefers or creates one of two directions. This can be achieved positively (e.g. gear teeth), frictionally (e.g. wrap springs) or geometrically via constrictions or contact pressures and, if necessary, also be translated so that, for example, a worm or screw continues to turn a worm wheel-like part with fine resolution, but the "ratchet effect" is achieved by turning the screw ratchet-like he follows. All of the ratchet functions described here can of course also be carried out with such "translating ratchets", however the translation is carried out exactly. There are very many "ratcheting" parts with certain advantages, such as fine resolution, are known. Hydraulic solutions can also be used here, which can be changed or direction-dependent, for example via a slot, a valve, viscosity or whatever. These "ratchets" can also be combined here with at least one additional function, so that they limit the moment, limit the stroke or allow the stroke from a certain state (such as moment).

In the present case, any behavior that is not based on a constant transmission ratio, such as a conventional gearbox, can be understood as "non-linear". This non-linear behavior can be defined in very different ways.

Examples:

• Curve between input and output force over actuation travel

• Limitation to only one direction of movement

• Limitation to a specific moment or force

• Allowing one part to move when another is stationary.

In the following, use is also made of the wording "via the actuation of a variable transmission ratio", which is used analogously to "non-linear", whereby here "analogous" is not necessarily to be understood as "exactly the same", but as "resulting in the same meaning “.

There are many ways to indicate the severity of braking, from perceived to physical quantities. Therefore, the term "braking effect" is used here, which includes all variants and can be expressed as braking torque, braking force, braking deceleration, etc. These effects are not mentioned individually below, but are understood to have similar effects.

"Pad position" or "pad stroke" can describe the position of a brake pad or values derived from it, such as the actuator angle. These values apply from a defined starting value, preferably the maximum distance from the friction surface (brake disc or drum or similar). After an air gap has been overcome, ie after the lining has come into contact with the friction surface (“contact point”) if necessary, the term “deformation” may be used, since contact pressure occurs from this point onwards, which leads to deformation or an overall deformation. The point of contact is not understood to be a geometric point, but the concern that is just beginning. All of this also applies analogously when several coverings are involved.

In the case of straight movements (such as with the brake lining), it makes sense to speak of force and displacement (or stroke) in connection with transmission ratios. In the case of rotating parts (such as contact pressure cams or actuator motors), the terms torque and angle are most commonly used, but one could of course also use e.g. circumferential force and e.g. displacement at the circumference. A position can be thought of as an angle, and thus of course also as a measurable variable, such as steps, etc., or as a linear measure. In the following, the expressions are used with a similar effect or with a similar meaning, i.e. "high force" also means, for example, a high actuator torque and only one term is listed, for example, but all with a similar effect are included. Since both rotary and linear movements occur in EMBs, force and moment or displacement and angle are usually used in the same sense, i.e. both versions are not named, although both usually occur, such as the angle of the actuator shaft or the stroke of the lining. This of course also includes the fact that an actuator torque can develop different contact pressures at different points of the non-linearity or that, for example, the lining position and the actuator angle are not directly related, but possibly via, for example, the non-linearity and the resulting overall transmission, connections. The terms "control" and "regulation" are also used interchangeably, unless the difference is expressly pointed out.

Terms such as "and", "or", "or" are generally not intended to be exclusive. In principle, features can also be present several times, ie, for example, several springs instead of one mentioned or several brake actuators instead of one mentioned actuator. Arrangement representations are a representation of several possibilities: if, for example, compression springs are shown, this could also be implemented with tension springs or combinations or other pushing or pulling forces. There are Modifications with the same or better effect are also possible, for example if a spring is supported somewhere else than shown.

Actuator - Configurations:

A wear adjuster is advantageously actuated with the brake actuator, however, but of course you could also use your own wear adjuster actuator.

Several electric motors can also be used, e.g. for safety reasons or for other purposes. For example, one of its own could perform the service brake function and another a parking brake function (which, for example, remains constant in a de-energized state) and the parking brake drive could also, for example, take over or support the service brake function in an emergency, for example.

brake actuator torque:

In all of the methods mentioned above, which use the brake actuator torque that occurs, self-energization should of course also be taken into account if necessary. Other actuation energies, such as springs or energy from thermal expansion (e.g. a brake disc would expand when heated, which corresponds to applied contact pressure, or a brake drum could expand, for example, which corresponds to contact pressure energy removed) must be taken into account.

For example, there could be a single optimal gear ratio curve with actuation variable gear if only one goal were optimized. For example, the shortest possible actuation time could be a single goal and one would receive the physically correct answer that the translation at every point must be such that the brake actuator would run with maximum shaft power. This would mean that the transmission ratio would have to change by several powers of ten, because the contact pressure is zero at the beginning and only very small displacement losses have to be covered and in the end, for example, 30 kN are necessary for an emergency braking of the front wheel of a car. It is recommended here not to realize such "optimal" transmission ratio curves, but to respond to requirements that are directly related to the sensible and favorable implementation under real conditions. Furthermore, it is recommended here not to strive for a single optimum, but to consider the cases that are essential in use as the “optimal target course”. In contrast to the above requirement, for example, a state with, for example, zero actuator shaft power by definition occurs very frequently, for example if a specific actuator position is not changed, for example in order to maintain the resulting braking effect. Here, for example, the thermal load on a stationary actuator with simultaneous heat development in the EMI would be included as an additional requirement with regard to actuator torque and the prevailing transmission ratio, with the actuator shaft power being zero, but not the electrical power. Here you would include the electrical power loss at the actuator, which can be small when the actuator is stationary because the holding current flows, but the small copper resistance causes little voltage drop and the current squared times the actuator resistance causes a low thermal output.

There can be many states in an EMB where it is suggested here that one does not strive for an optimal course, but rather considers the essential states. For example, spring-actuated brakes are also covered, in which a spring force takes over the actuation and an actuator force the release. When holding the released state, it is suggested here that one does not hold released with e.g. the "optimal maximum engine power", but quite the opposite, for example, with that minimum actuator torque that just allows safe actuation under all given conditions.

With the interpretations proposed here, it is not so important how the optimal target course of the non-linearity came about; the proposals here are mainly about realizing an actual course of the non-linearity that fulfills the conditions, where the resulting disadvantages (such as the fact that the theoretically shortest possible actuation time can no longer be achieved) are of course kept to a minimum. Since the task does not have a single possible solution, solution variants are compared with regard to their advantages, where one can of course also be satisfied with a single or first solution out of several theoretically possible ones, especially if one already has an overview of similar solutions. Brakes as proposed here will also often combine multiple non-linearities, such as a cam operating a lever. In this case, the mechanically and geometrically favorable solutions will be used, for example both, if necessary, and an advantageous overall actual non-linearity will be striven for. However, several non-linearities in an EMB can also be designed and interact differently. For example, a spring force can act like a crank on a cam that actuates a pressure lever, in which case three non-linearities carry out “optimal” pressure. As described above, usually not a single optimum is taken into account, but a target curve that results, for example, from the fact that the relaxing spring can always exert enough effect to press the lining over all states.

Adaptation to framework conditions:

The shape of the cam, in particular the maximum twisting angle and the leverage used, expressed by the minimum and maximum cam radius, are crucial for the size of the brake that can be achieved. Size naturally means space requirements, but also weight and price. However, the available installation space in particular can be severely limited in the area of a brake due to other components such as the rim, wheel suspension or drive shaft, but also e.g. due to spring and steering movements. It is therefore of little practical relevance to achieve a non-linearity that is recognized as theoretically optimal with a very unfavorable or even impossible size.

So here it is proposed to design the cam track according to the geometric and mechanical improvements. In this regard, it can be interesting, for example, to keep the cam rotation angle well below 180° if collisions could otherwise occur during cam rotation.

There can be completely different tasks and states for different cam positions. For example, one position in a spring-actuated parking brake can be designed for the lowest possible released holding torque, while the subsequent area should allow the contact pressure force to be applied quickly. in the This is shown below for a service brake, where a high pad movement speed is required in the air gap and the resulting contact pressure is intended to cause a significant change in the actuator torque, e.g. to be able to clearly identify contact with the brake pad from the actuator torque curve. A small radius for the roller running on the cam is recommended for this strong change in the initial behavior, because the cam track can be designed more easily for small roller radii (especially without points that are not possible in practice, see above).

Design process:

From the comparison of the suggestions, see if necessary the Fig.11, 1201 - 1202, 1301 -1302, (rounding radius with wrong slope, shifting of the radius for correct slope, reduction of the total angle of twist) one can see the interesting effect that not all "compromises" have a similar effect. The mere application of a fillet radius can cause non-actuable braking conditions, the shifting of the radius only causes a minimally larger necessary twist angle, which in turn could be reduced (whereby of course the minimum radius would have to be controlled again) and one would come up with a combination (reduction of the larger twist angle). with subsequent verification of the minimum radius) to a solution that can be very close to a target course. It is interesting, then, how easily both non-functional solutions and those close to the target requirement can arise.

Another tried-and-tested method can be to work on the non-linearity in the abstract and, if necessary, to check the effect of the change, for example which actuation time behavior arises. If you make the conversion of the non-linearity into the cam track easy to handle (mathematically speaking "only" a rolling curve), you can also quickly observe the developing track for the non-linearity that has just been changed and in turn make local changes to the non-linearity, e.g. changing the transmission ratio via a cam Expand a somewhat larger area, especially if you recognize the cam or non-linearity area in which the need for improvement lies. To do this, however, it is helpful to ensure that the conversion from non-linearity to cam surface can be carried out quickly, or vice versa, if you want to represent a change in geometry as non-linearity, e.g. transmission ratio via actuation.

There are helpful approaches for such a conversion. One can, for example, assume a force or torque transmission ratio of non-linearity via, for example, the torsion angle. For example, this could be considered the "first derivative" because it relates to geometric slope. So it is suggested that one needs an integral to get from the slope to an absolute value. It can be suggested as helpful to first determine the center path of the unwinding roller as a "cam track" that is easier to determine with an imaginary roller radius of zero. Now it is proposed to project the center points over the radius onto the cam surface. Of course, the steps do not have to be carried out exactly as suggested. Some things can also be simplified, summarized or solved in a similar way. Above all, it is important to present a path, however similar, from non-linearity to career. This can or should be automated, e.g. in the usual way with Matlab-Simulink or a similar language. To what extent the fact that it is simply a mathematical "roll-off function" helps here can be taken into account by an implementer of this proposed procedure.

It is also proposed to represent an "inverse function" to the above, i.e. to project e.g. from the cam surface onto e.g. the roller center track and then to "differentiate" the radii of the center track in gradients and from this to win the torque transmission ratio via the angle, with this reversal path somewhat may seem easier. You only need to solve one of the two paths, for example only that from the surface track to the transmission ratio curve. The inverse function can then be obtained via iterations, for example, i.e. via one of the suitable iterative solution methods, also known as "root finding". You can solve these tasks point by point, which is more in line with human understanding, because you can think about what exactly needs to be done for a point. It is proposed to adopt such a "point-by-point solution" as a general solution function, because a solution that can be shown for a point can also be formulated in general as a function. Instead of cams, ball ramps, for example, can of course also be used with a non-constant pitch or non-constant radius to the ramp pivot point, or other non-linearities such as levers, cranks, pairs of wheels with a non-constant radius, etc. can be used. In general, this proposed conversion of non-linearity into geometry and vice versa can also be referred to as transformation.

Mathematical inaccuracies can also be compensated for. The mathematical generation of the cam surface from the roller center curve can lead to a slightly different transmission ratio when the unrolling actually takes place or when the reverse mathematics from the unrolling produces the roller center curve again, especially in the area of the strongly and locally rapidly changing transmission ratio. This can be compensated for by superimposing this found deviation on the desired roller center curve, which is subsequently transformed back from the real unrolling, to the desired roller center curve with the correct sign as advance compensation, and from this the cam surface is determined.

This can all be applied similarly to other unrolls, such as ball ramps.

This interpretation of the non-linearity is not limited to the actuator torque, the actuator torque was only used as an example above. Likewise, there can be a non-linearity in a spring actuation, for example, or in the remaining moment between the spring moment and the actuator moment, or in any non-linearity, however used, in which the target behavior can be expressed via the actuation. The favorable influencing of unfavorable gradients on a non-linearity can also be favorably influenced with a further non-linearity, for example by a non-linearity being able to be designed only for a geometrically and mechanically advantageous course and a further non-linearity further improving the course to achieve the overall target behavior. For example, it can be very advantageous for a spring-actuated EMB if you combine a very strong non-linearity of a spring linkage with a cam non-linearity: The spring would be maximally tensioned when fully released and maximally relaxed when fully braked. With the cam one could aim, for example, that the relaxed spring effect results in the highest contact pressure and the fully tensioned spring effect in the same way the contact pressure acts so that the brake can be kept released with minimal torque. This can mean an extreme change in the cam ratio in the transition area within the air gap to the beginning of contact pressure. If the spring, for example, acts on a crank-like drive of the cam in the fully tensioned state, the tensioned spring can be started almost in the dead center close to the spring and thus obtains a spring momentum on the cam that increases sharply in this area and can thus be achieved by combining these two non-linearities change the cam ratio less quickly or strongly. Of course, the same could also be achieved with other combinations, eg including a ball ramp or variable radii.

The proposed method and the proposed cam can now be formulated in general as follows:

There is a target curve of whatever kind for the non-linearity over the actuation. This can lead to a possible cam track.

However, it can also lead to an impossible or undesired cam path, especially if geometric and mechanical restrictions are applied, such as cam radii, cam twist angles or mechanical loads. From this, an "improved" cam track can be proposed and it can be determined whether the resulting non-linearity curve is tolerated or whether it is further improved.

Or, for example, a more practical target course for the non-linearity can be specified, the associated cam track can be determined and this can be checked again to ensure that the restrictions are met.

It may be necessary to run through these iterations several times until a compromise is reached between a desired course of the nonlinearity and the fulfillment of the constraints.

From a mathematical point of view, these iterations can also be avoided if a mathematical connection between the course of the non-linearity, the cam track and specify the restrictions. However, this is not easy insofar as the cam tracks are "roll-off curves", which, however, does not lead to a simple mathematical representation in the general case.

Of course, all this can also be applied to other unrolling, such as ball ramps, and it can also be applied when there is no unrolling, but any nonlinear translation. There is always the desired course of the non-linearity and the course that is possible under restrictions, and despite the restrictions, one will strive for a non-linearity that is as close as possible to the target non-linearity by means of mathematical and/or iterative solutions.

"Coming as close as possible" will be evaluated in several ways, e.g. how big the time disadvantage of the brake actuation is, how high the actuator torque increases from the desired value, which rounding radii are allowed or which geometric disadvantages are accepted.

Advantageous parts and finishes

Loss-reduced expansion parts:

It is proposed that a rotational movement for actuating the brake is advantageously also generated in the brake. For example, rotatable expansion parts in drum brakes can be used, and in general, for example, cams, eccentrics, levers, ball ramps, whereby these parts can also be non-linear.

wear adjustment:

Furthermore, particularly in FIGS. 20-2302, advantageous examples of wear adjusters are shown, with two functions being derived from the movement of the brake actuator, namely normal brake actuation and wear adjuster. In the case of mechanical, hydraulic or pneumatic brakes, for example drum brakes, there are various known readjustment methods, for example if there is too much travel or if there is still too little contact pressure after a certain actuation. All of these methods are of course possible here. Parts can be used particularly advantageously here whose behavior changes under the influence of force or moment, e.g. bending, evading against a spring or not yet evading, so that, for example, a change in a specific actuating position (or an area, e.g. when the covering is straight begins to build up contact pressure) is expected, and that, for example, if this change does not occur, it is concluded, for example, that there is too much air gap. If the change is not carried out, a function is then triggered, eg actuation of a wear adjuster. For example, there could be a leading springy part, e.g. on the lever or on the cam, which is normally pushed away when the contact pressure begins, but is not yet pushed away in this actuation state without the beginning of the contact pressure force and therefore carries out a wear adjustment or notifies it to be carried out later. It is also possible, for example, after a certain angle has been exceeded, to carry out an adjustment movement via a limit such as a slipping clutch, so that the adjustment is not carried out when the contact pressure is applied if the slipping clutch slips from a certain position.

In particular for FIGS. 20-2302, a rotational movement of the brake actuation is assumed. It is assumed that the wear adjustment is added to this rotary movement, i.e. more twisting has to be done with wear. Of course, disc or drum brakes or others can be used, advantageously the same type on one axle. In all versions, other movements, such as pulling or pushing movements, can also be used instead of rotary movements. Individual brakes can be actuated as described or e.g. the brakes of an axle or a group of axles together and the wear adjustment can also be carried out separately or together for a brake, an axle or a group of axles.

However, the wear adjustment does not have to be included in the actuation movement, but can also be fed separately to the brakes, similar to what is shown. A complete EMB with actuation actuator and wear adjuster can also be used on one side and only the brake mechanism alone on the other side, which is also actuated by the complete EMB, or any number of EMBs can be actuated by any few complete ones. In all of the following versions, at least one spring can also be involved, for example to hold a parking or service brake position or to assist in releasing or actuating the brake. In these cases, the behavior of spring(s) and brake actuators must be summarized with the correct sign and related to the joint effect (torques, forces).

In the case of brakes that are operated together by just one actuator, the adjustment (e.g. via a ratchet effect) can also be carried out separately for the brakes. For example, the adjuster parts for each brake can be available separately and operated separately by two adjuster ratchets via an elevation (e.g. pin) and a compensating part (e.g. spring, torque, force, travel limiter) to make a wheel-specific adjustment by e.g Brake with more air gap gets a longer stroke due to smaller force on a spring. Compensation similar to that of a balance beam can also be advantageously suggested, e.g. by e.g. the side of the beam that comes into contact with the surface finishes earlier and adjusts the other side more. For example, a roller on a lever could, in abstract terms, act as a roller between two levers, so that both levers can find a position for similar force build-up. Then the roller could have a convex rolling surface, for example. Such rotating or other position-changing, balance-like compensating parts are of course proposed as being practicable, so that the above solution can be thought of as a principle, for example. "Arrangements one behind the other" can also be recommended as equivalent, so that e.g. one brake first builds up actuating force and thus causes force to also build up on the other, e.g. one part is brought up to another and then both build up force.

Such compensation parts, which in principle are similar to a balance beam, but can also be solved differently, such as a differential, can also be referred to as "kinematic chains" and have, for example, one input and two outputs and can be used here in any compensation function, e.g particularly advantageous, for example, to compensate for small differences in, for example, the actuation path when the brakes are actuated jointly. This can be thought of similar to a hydraulic Imagine compensation, which is of course also possible here and sets the same pressure on two outputs, for example.

The adjustments or individual operations can also be combined, for example, one of many solutions proposed as particularly advantageous is to provide a separate wear adjustment (e.g. ratchet) for each brake or, for example, each side, so that the brakes can respond to similar lining contact behavior ( to the e.g. drum or e.g. disc). Differences can then be compensated for using a balance beam-like behavior, so that e.g. if the ratchets were set "one tooth different", the balance beam-like behavior can compensate for the contact pressure.

For the special requirements of the EMB (e.g. position control instead of the usual force control), the above statements are of particular importance, so that the completely different controls (position or force) cannot simply be equated. Position-controlled brakes are practically completely new territory, because position measurements on brakes have so far only been used for laboratory or test purposes.

If, for example, more than one brake is operated by only one actuator, the brakes can preferably also be adjustable in terms of their similar behavior to the lining contact with the drum or disk, for example, so that, for example, adjustment options (which could, for example, safely maintain the state via friction) can be set uniform application on all brakes is adjustable. Pairings of the brakes and brake parts for low overall tolerance can also be recommended, such as packaging similar brakes or a combination of pads and drums, for example, so that similar overall properties result and it can also be recommended, for example, to process them before delivery, for example, such as grinding the pads ( also in the state already installed in the brake, for example). The pads can also be shaped such that when new they preferably fit in the middle of the long side of the brake shoe, for example, in order to reduce tolerances from initial contact points, e.g. whether the pad initially contacts the actuated shoe side first or the non-actuated side. Particularly in the case of “servo drum brakes”, it can be advantageous to mount the actuation of a brake shoe together with the support for the brake shoe on one component, for example on a plate that can be rotated around the wheel hub, for example. This creates a stabilizing effect on this brake shoe, because this shoe can be seen as a simplex shoe from the point of view of its operation and this "stabilized" effect can be passed on to the second shoe. Otherwise, with the servo drum brake, the support point of the first shoe would move away from the actuation. If, as suggested, this wandering away is suppressed, there is also an actuation that is more favorable in terms of the overall transmission ratio. With a normal servo drum brake, the movement of the first shoe would result in a longer actuation path at the actuation point of the first shoe. If, as suggested, the actuation point and the support point are in one piece, the relative actuation travel for the first shoe remains smaller (can be as small as for "Simplex"), although a servo effect (to actuate the second shoe) arises from co-rotation . With this installation, the strong dependence of the self-energization on the coefficient of friction can also be reduced, because after this installation, a first simplex brake presses on a second simplex brake. This assembly and the overall support or mounting of the joint support of the first shoe can be designed in such a way that there is a dependency on the direction of rotation or this dependency on the direction of rotation occurs as little as possible or not at all.

These supports can also be used to record forces, e.g. by measuring or switching. For this purpose, additional resilient parts can be used or, for example, the stiffness characteristic of the "second simplex brake" can be used to convert a force measurement into a distance measurement. If the second brake shoe is considered here as a simplex brake, its stiffness characteristic indicates the force-displacement relationship, i.e. one can deduce the braking force arising from the first shoe from the entrainment movement of the joint assembly and, in particular, recognize whether the first shoe is already developing braking force or is still in the air gap.

Influence on control: In all the versions mentioned, position measurements, for example of the actuator shaft angle or the cam angle or the lever angle, etc., are recommended, with a simple version also using, for example, stops and the resulting reaction detected to find the position or recognizable areas. For example, the area between the parking brake side and the service brake side of a cam can be recognized by the two increasing motor currents. For brake control, for example, analog electronics are recommended if, for example, simple position control with a potentiometer, for example in the area of the cam, is possible without the expense of software (and possibly its safety problems), for example the actuator position via a target/actual comparison. The motor can, for example, be a DC motor (also a geared motor at low cost) and be supplied via an analog circuit. To avoid losses with analog motor control, the motor (also DC motor) can be operated with pulse width modulation, which is controlled analogously, for example. The comparison of target and actual braking effect (eg deceleration, positions, overrun strength) can also be carried out in the same way. Digital controls are of course also possible, as well as mixed controls, for example analog comparison with digital ABS or ESC, but neural networks or fuzzy logic are also possible and separate structures, such as that one part is in brake electronics and another in one other device.

All of these versions are not tied to parking brakes and service brakes. Many other requirements can also be solved in the same way as those described, only one of the functions can be used or new ones can be added, e.g. a service brake that releases heavy braking without current and maintains or starts weak braking without current. Influences can also have an effect, so that, for example, too high a braking effect reduces the actuation position and self-reinforcing effects can also play a part, which can be taken into account in the design. All moments and forces can also be correctly related to one another, such as self-reinforcement or mechanical losses, and various states can preferably be included, such as changes with lining wear, temperature or aging.

Possible advantageous properties and designs of the braking device are listed below. However, the features described below can do not have to be features of the braking device according to the invention. The braking device according to the invention can include and/or have the features listed individually or in combination, ie in any combination.

That non-linearity and brake control can be designed in such a way that any part of the lining wear or a wear adjustment that has not yet been carried out or has not been carried out correctly can be compensated for with the brake actuator or that the brake actuator assumes such positions that cause a correction or these non-adjusted lining position deviations correct. That between the scanning of a cam (e.g. by a roller) and the generation of a rotary movement (which e.g. rotates on an expanding part) apart from a lever there are no other transmission parts influencing the movement process, i.e. that the roller rolling on the cam is mounted directly on the lever is without, for example, a push rod, train transmission, etc. being interposed. Of course, this does not apply to parts that are not necessary for cohesion, such as a bolt in the middle of the roller, rolling elements for roller bearing, rings of bearings, etc.

In the case of a spring-actuated brake, in particular a parking brake, the variable transmission ratio via the actuation runs in such a way that even if the air gap is incorrectly set, the brake can be released with the brake actuator against the spring action, or that even in the case of an extreme incorrect position of the air gap, e.g the friction surface (e.g. brake disc, drum, rail) it is possible to release the brake against the spring action, which may be necessary, for example, when the brake is removed or during assembly.

In the case of a spring-actuated brake, in particular a parking brake, the transmission ratio that can be varied via the actuation is such that even if the air gap is incorrectly set, the brake can be released with a device counteracting the spring action, or that even in the case of extreme misalignment of the air gap, such as the absence of the Friction surface (e.g. brake disc, drum, rail) it is possible to release the brake against the spring action, which may be necessary, for example, when dismantled or during assembly, the device being a screw, a wrench attachment on a moving part, such as a transmission shaft or the cam etc. can be. With a spring-actuated brake, in particular a parking brake, there are several positions in which the brake remains without a moment generated electrically by the actuator, ie both in the released state and in the braked state, and that an additional moment is introduced for changing the states must be, for example, via the brake actuator or, for example, via a part that is accessible from outside the brake. This can be used e.g. as a "bistable parking brake" that remains in the parking brake state without power supply, can be switched to the braked or unbraked state with power supply and safely remains in the released state by switching off the power supply, with the power supply being switched off e.g. outside the brake can take place or, for example, within the brake and, for example, only parts of the power supply can switch off, such as for the brake actuator.

That with a spring-actuated brake, in particular a parking brake, the spring actuation without a moment electrically generated by the actuator only achieves a braking effect below the full braking effect and with a moment electrically generated by the actuator, a higher braking effect is generated.

With a brake that is essentially actuated by spring force and essentially released by the brake actuator, the non-linearity with the mechanical and geometric restrictions can be designed in such a way that in the released state only the maximum holding torque required for safe spring actuation is required up to to none at all and that possibly also a release movement with the brake actuator is possible if the linings are completely worn up to possibly no linings at all or disc, drum or rail.

That the friction surfaces can have any shape, such as disks, drums, rails, or that the relative movements to be braked can be rotating, linear, or whatever.

That a spreader including any storage and one or more primary brake shoes are mounted on a moving part that a Self-reinforcement caused movement of the brake shoes leads to no relative movement between the expansion part and the brake shoe.

That the brake shoes of a drum brake are spread apart with a spreading part, in which the contact point of the pressing part on the shoe follows the movement of the shoe as well as possible.

That at least one own wear adjuster is available or the adjuster is actuated with the brake actuator.

The fact that a part is moved for readjustment, e.g. a lever is pivoted, and this pivoting can also affect the actuating cam or e.g. the entire actuating assembly with the motor.

That the adjustment can also be carried out with liquids, for example, or that the lining pressure is carried out via an intermediate element with liquid.

That any vehicles and devices are equipped with this brake, such as cars, commercial vehicles, buses, airplanes, trailers, elevators, machines, position holding devices, emergency stop and safety devices, device shafts such as propeller shafts on wind turbines, ships and others.

That, after applying different approaches to quality assurance, concrete correction values for individual parameters that describe the behavior of the brake, such as the size of the air gap or stiffness parameters, are finally specified and taken into account in calculations from this point in time and until more recent values are available.

That the brake control electronics take into account that recorded actuator data, such as motor current, are subject to fluctuations due to geometric irregularities of the friction surface, which show a speed-dependent pattern when the friction surface rotates, and this is reflected in the data interpretation. That this pattern is used to detect contact between the friction surface and the brake pad.

That friction surfaces are equipped with geometric irregularities in order to be able to detect contact with a brake pad.

That the mechanical losses in the brake actuation (in particular, e.g. also the static friction) in the evaluation of the actuator torque (e.g. to determine the pad contact pressure force) are occasionally or permanently reduced by vibrations or the like, e.g. vibrations from the operation of the brake or the object to be braked help , to reduce the friction in the brake actuation by "shaking" or to overcome the static friction or that such vibrations or oscillations are brought about intentionally, e.g. with the brake actuator, whereby statistical methods can also support those caused by vibration or "shaking". Calculate or suppress deviations in measurements. Other well-known effects can also be calculated out, such as the power consumption caused by acceleration or deceleration in the mechanics and/or the actuator, in order to obtain overall measured values that are free of mechanical losses as well as possible on the one hand, but have as few influences as possible on the other due to vibration or oscillation. These values can be processed arbitrarily, e.g. statistically, e.g. as angle-torque pairs or just as many measured values to determine or subtract the mechanical losses, e.g. to use a certain type of averaging or e.g. low-pass filtering over all measurements. Vibrations of different strengths, for example, can also be used or brought about in order, for example, to determine different contributions to the mechanical losses, in that different parts are affected differently, for example, or in order to increase the accuracy. The use of vibrations to overcome friction, in particular static friction, is known in actuators in order to also carry out small adjustments. It is therefore proposed here to apply this principle to measurements in order to determine values that are as free as possible of friction, in particular static friction in the brake actuation. These measured values can, for example, also be compared with stored ones in order to to obtain, for example, one or more values from the large number of values or comparisons, such as mechanical losses or actuator torque.

That a force control or a displacement control or a combination of both or a change between these controls is used and e.g. an instantaneous force-displacement characteristic of the brake is assumed and e.g. in the event of changes to the brake actuator setting it is switched to a position control using this instantaneous force-displacement characteristic and then if necessary, for example, a change is made to a force control, or that, for example, both types are operated simultaneously and a specific respective proportion is used via a (also variable) weighting of both.

The fact that various parameters that can be used as a complement to brake control are used in combination in such a way that one parameter represents the actual control variable for which a specified target value corresponding to the current braking effect requirement is achieved as precisely as possible by the electronics in order to, for quality assurance, additionally for derive one or more parameters from the current braking effect requirement value ranges that must not be left during the setting process for the control parameters. For example, a force control based on the effective motor current and the local transmission ratio can represent the actual control and a range for the permissible motor position can also be defined, which avoids serious incorrect settings.

The statuses or measurements on the actuator used for detection can also be used for purposes other than those directly related to braking, such as a stop that can be used to find a starting position when it is reached or, for example, a wear adjustment that is determined, for example, by measurements on the actuator different actuator torque can be distinguished from a position for determining the initial position and can therefore, for example, fulfill two functions, namely, when the actuator torque starts to be smaller, for example, it can first be used to determine an initial position, and further actuation in this direction can cause wear adjustment, for example, or also the extent of the can influence wear adjustment. That the electrical or electronic brake control or regulation supplies the electrical energy and/or the electrical current (or an effect-like quantity such as power, torque, thermal effect, etc.) required to hold a position (or, for example, an actuator angle) or a position range below the Achievement of the position or the position range required value lowers, which lowers the current required to keep released in the case of a spring-actuated parking brake, for example, or that, for example, an actuation range for prolonged braking is operated with a lowered current. This can also be caused by the property of the actuator control without consciously causing it, if, for example, a proportional controller sets only a small actuator current for an exact position or a small deviation and sets more current for a larger deviation, for example. Of course, additional uses can be helpful, such as using the area of static friction in such a way that the static friction allows a position to be held even with a smaller current, or making changes, for example, minimally abruptly, i.e. for example with small changes in position in the direction of higher actuator torque, not increasing more and more sets the current, but makes a small jump (which, for example, could not be felt at all or hardly at all for the braking effect, at least is accepted here) in the position or the actuator angle and then uses the current-reducing advantage of static friction again. Any method that uses the current reduction option by using a certain state in the area in which the mechanical losses make it easier to hold a position is recommended here. For example, you can also insert current reduction attempts to observe whether the position (or a position range) is retained, or you could, for example, deliberately approach a slightly incorrect position in order to then reach the target position (or a position close to it) by reducing the current. A current value (or, for example, a value of an actuator torque), which just barely allows a position to be held, can also be included, for example, in the determination of the mechanical losses. Of course, foresighted or knowledge-based methods can also be used, such as setting a point with lower power consumption on the non-linearity where, for example, the braking effect is little or no different.

That an input current reduction (eg DC power supply) in the electronics of the actuator control or regulation is achieved in that the actuator with operates at a lower speed than would be the case without the intended reduction in input current, in order to be able to lower the average voltage applied to the motor by the electronics due to the lower voltage generated by the running motor, while the input voltage into the electronics continues to correspond to the approximately constant supply voltage (where "current" also includes effect-like quantities).

That this is used, for example, to keep an overload or an avoidably high load away from a power supply and, for example, can also affect several EMBs or this can also be communicated to or between the EMBs, for example.

That short-term peak values of the actuator power supply, such as those caused by highly dynamic motor controls, especially in the case of sudden and strong changes in the motor position command, are prevented by limiting the rate of change of the default value for the torque-generating motor current, without leading to a significant slowdown in the entire motor actuation.

That measurements such as brake actuator torque, position, speed with sign, temperatures are recorded several times, treated with statistical and mathematical methods (e.g. averaging, grouping according to various criteria), compared with stored values and with each other and that statements about the current state of the brake, such as wear to be adjusted, air gap size, brake rigidity, pad material thickness or error messages, error entries, warnings, data to the environment and the driver.

That the brake can receive signals from the outside (e.g. brake control, sensor data, parameters, software) via wire, wireless, radio, internet, telephone, infrared, etc. and data via wire, wireless, radio, internet, telephone, infrared, etc. to the outside can give. That information about the current braking effect, such as measured deceleration, overrun effects or current consumption of the brake actuator, is converted into signals that give the person controlling the braking feedback on the braking effect achieved and to transmit these signals to the person simply by sensors, such as eg as dynamic resistance directly on the brake lever or pedal, eg via electric motors or magnets, or also via other signal forms that can be modulated, such as vibrations or noises.

That there are sensors that indirectly detect contact between the friction surface and the brake pad, e.g. via vibration or sound waves.

Within the scope of the present invention, it can be assumed that lateral compensatory movements should not be minimized in principle, but that they can either take place harmlessly in intentional lateral play, or can even be intentional in order to follow changes in geometry. The compensating movements that may be permitted in the braking device on the one hand convert actuating energy into unwanted friction and on the other hand they can be a wear problem, depending on how often, with which contact pressure forces and with which materials they occur. If, for example, few full stops are assumed, the wear caused by compensating movements can be insignificant. If the air gap is passed through very often before the lining comes into contact, the wear caused by compensating movement can still be insignificant if hardly any contact pressure is required, e.g. only against a spring.

If there is small lateral play (as suggested here as a possibility), e.g. the lateral compensating movement can be absorbed by the play and thus wear caused by a scratching movement can be avoided, which can be used e.g. in the area of very frequent normal braking.

The loss of actuation energy due to a lateral scraping compensatory movement can be estimated, for example, if it is assumed that, for example, a pad contact pressure stroke of, for example, 2 mm is covered and, for example, an unwanted 0.2 mm frictional compensatory movement with a metal-to-metal coefficient of friction of, for example, 0.1 takes place: Then the lateral force would only be 1/10 of the pad contact pressure force and the lateral movement would only be 1/10 of the pad contact pressure movement and the energy loss would therefore only be rough. 1% of the actuation energy.

Control, mechanical losses:

The process described in FIG. 30, for example, can of course also be modified with the aim of determining the status, for example by omitting or changing the sequence. The processes can be erratic or arbitrary, such as sinusoidal or S-shaped (e.g. speed or movement profile). , but they can also be superimposed on the course of movement (e.g. due to a change in speed, a change in current, even up to a brief shutdown and/or even a reversal of the current direction). The processes do not have to be triggered by this method either, but other processes can also be used. For example, the driver can use a “brake release” to observe the actuator acceleration. In particular, it applies here that, as is well known, “no energy can disappear or be gained in total”, for example based on the fact that the sum of moment from mass inertia plus moment from brake actuation plus moment from losses plus moment from the actuator plus moments from other things ( eg springs) must be zero. In particular, it is proposed that intentional or unintentional changes (e.g. the actuation) are also examined for the conversion of the energy form: For example, intentional accelerations (or decelerations) could be inserted in an actuation speed in order to determine the reaction or the accelerations ( or delays) are not inserted intentionally, but can also be done “of their own accord” or, for example, by the driver. This brings us to the general formulation of the method: Each actuator movement or change in it can (should) be examined for conversion of the energy form, including conversion into losses if necessary, in order to find parameters of the process such as total losses, partial losses, expected actuator values for certain Braking, etc. In particular, one can, for example, know the motor torque (or, for example, the torque-generating current) with the known mass inertia, the presumed clamping force from the brake or the clamping force derived from the measurement Investigate spring effects and possibly other known influences to determine how the sought-after influencing variables (e.g. losses) must be (or are suspected) in order to explain the actuator torque curve, possibly taking into account the conversion of the energy forms. Of course, this can be done to determine a wide variety of results, for example to explain the engine torque curve for certain actuator observations. In general it can be seen as finding an explanation for an observation. It could also be called a transformation: With a Fourier transformation, for example, a temporal amplitude curve is transformed into the strength of frequencies. Here, for example, a temporal curve of an actuator moment, for example, is transformed into parameters (e.g. losses), which are seen as contributing to the curve.

For control or regulation (both terms are used here equivalently, unless the difference is pointed out), sensors were used in the past, mainly for the contact pressure force, for example. This is of course also possible here, but it is also recommended - if sensors are required - to use them for the right purpose, namely the braking torque. Patents for "sensorless" control (without force or moment sensors) also exist, of course, whereby the contact pressure force is mainly inferred from the actuator motor current. It is also recommended here that the known acceleration of the mass inertia is deducted. Then there are the unwanted mechanical losses (since they make the connection between the motor current and the contact pressure inaccurate) which, as far as they are known, should of course also be factored out, which is also recommended here. Of course, deviations in the control behavior of the real brake from the planned or theoretical one should also be recognized. Of course, there are also patents for this in which measured values are compared with stored values and these obvious methods are of course also recommended here. Since the losses in the actuation direction increase the actuator torque and in the release direction less torque acts on the actuator by the losses, it is recommended to use this difference as a measure of the losses (actually double the losses when reversing the direction), but in a different way to what is already known . It is known that a real operating behavior is compared with a stored one, which of course is also fundamentally possible here. Here, however, additional or alternative suggested that NO comparison with stored data is particularly interesting, because a comparison with stored data is always associated with the problem of whether the stored data came about under the same conditions as the currently measured behavior. Of course, you could save many behaviors and select the most appropriate one, but the problem still arises as to whether the same thing was really saved under ALL conditions, there can also be many factors that have a greater or lesser influence and whose influence is not or not fully taken into account when saving became.

Therefore, it is also proposed here that the difference between actuation and release is used as a measure for the losses, but without reference to what is stored. However, this can add a new task: Activation and release can be so far delayed that the brake has changed (e.g. due to thermal expansion) and one would possibly form a difference from more or less incoherent things. On the other hand, it is firstly suggested that the changes are kept small and, for example, a difference is only formed in the air gap, in which no heat input is yet formed. Secondly, it is recommended to keep the time and thus the change short, with which you can, for example, follow the actuation with a minimal release, which can also be so small that it is imperceptible, because it's only about the difference between actuations and loosening is possible, or you overshoot the actuation minimally and loosen easily, or you can also install minimal direction reversals during the actuation, even in such a way that the reversal can be imperceptible. Thirdly, it is suggested that "a not particularly accurate determination is still better than none", which means here that braking processes are used in which, for example, no particular change occurs in the brake, that would be, for example, the many light brakings in which, for example no intense heat occurs. Fourth, it is proposed that braking can also be compensated in a similar third way, e.g.

Warming and thermal expansion are known or can be modeled and, for example, the influence of thermal expansion is calculated. It is also proposed here that it is particularly interesting to make actuator movements for difference formation that are not intended for a covering movement and/or do not cause any significant movement. Of course, it would be helpful if one could expect a known moment or a known course from the actuator movement. Regarding the known It is suggested here that the course of the actuator torque is recognized from the point of contact with the lining, and the actuator angle can thus be deduced upon contact. An obvious method is also already known, with a behavior being determined during a first movement of the pad carrier against a spring. In fact, springs are often found in such brakes, which press back the lining or hold the mechanics together and their use for calibration seems obvious if the spring effect is known. For cars, for example, the clamping force of a front wheel disc brake is roughly 35 kN. Statistically, the vast majority of braking takes place at roughly 1/3 to 14 of these, i.e. at roughly 9 kN. Up to and roughly within this range, for example, you want to control the brakes relatively precisely; ABS or ESC would help against emergency braking. Particularly weak braking (e.g. black ice) would be effected with roughly 3 kN clamping force. If you were to install a spring with a few kN in the pad actuation, you could actually generate the lowest real pad forces (translated to the actuator) at which you could calibrate against the really weakest braking. Such a spring would, of course, be further tensioned during further braking and would cost additional actuation energy and increase the size of the actuator. In addition, however, such springs would require considerable installation space and costs, and weaker ones will be tried. However, it should be taken into account that the floating caliper can jam slightly to severely and is exposed to additional forces such as cornering force or vibrations. With roughly 10 kg of floating caliper mass, rust, dirt, cornering, vibrations, these forces can easily reach hundreds of Newtons and a spring of this magnitude can, in the worst case, cause even worse than no meaningfulness, namely the "measured" could be interpreted as spring force and following this assumption, one could trigger a significant malfunction of the brake.

Therefore, another method is proposed here with regard to the calibration of the actuator torque measurement, which does not have the above problems:

First, at least one measurement combination of actuator angle (or a propositional measure such as position on a part that is motionally coupled to the actuator) and moment (or a propositional measure such as current, Power, force, etc. on the actuator or a part that is coupled to the actuator in terms of movement) made up of at least one movement and secondly provided that this movement is free or low in disturbing influences and thirdly the measurement can be interpreted conclusively, e.g. to the accuracy to improve the actuator torque measurement or to determine losses. A calibration spring has already been proposed above, which can be, for example, in an actuator rotation range which, for example, makes no or no appreciable pad travel. In this way, disruptive influences (see, for example, above) from the pad lift, such as forces, are avoided. Losses can be measured along the path to spring contact, see Fig.30. As shown in FIG. 30, the actuator would overcome losses when covering a negative angle, which would also be negative because of the negative direction of rotation. If no force is taken or added for other purposes, the actuator torque now corresponds to the losses and can be recognized immediately, even with no difference to a different direction of rotation. These are "no-load losses", e.g. of a motor gearbox. These can be different, for example due to the different position or viscosity of the fat, so it is good to know the current value. Loss fluctuations can also be detected to a limited extent in the course of rotation. From spring contact, the spring characteristic can be recorded and also compared with the spring characteristic of the actually installed spring or, for example, angular points on the spring characteristic are connected to a resulting torque from the spring. For example, if this spring is in the rotational movement of the non-linearity, the spring can be relatively small in contrast to the spring discussed above in the pad travel and still generate significant actuator torque, because a further translation between rotation of the non-linearity and pad travel greatly increases the contact pressure. "Notable" can mean that, for example, roughly that actuator torque is generated that later corresponds to a normal or light or defined brake actuation and one already knows what torque can then be expected when actuation, also with the problem of losses ( which were already included here). This spring also requires no useless tensioning energy in the brake actuation. It doesn't have to be a spring either, it can also be a rubber or a stop, for example. A stop would be very high

Cause deceleration forces when driving into the stop (e.g. to find it), which a spring or rubber with lower deceleration forces can do. It does not have to be an explicit part either, an existing part or any part can also be used, "nothing" would also be possible in the sense that the actuator does not continue to move in this direction. Torques that occur when a function (eg a wear adjuster) is actuated can also be used. Something that can be found by the actuator moment (e.g. stop, spring, rubber, etc.) is also recommended here in the sense that a starting position can be found or defined at the same time.

If the actuator now turns back towards the starting position, the losses suddenly appear in the other torque direction and when the direction of rotation changes, the losses are in principle twice as high. This process can take place, for example, when a brake is switched on and, for example, make the following statements: how large are the no-load losses, also with possible fluctuations, also possibly dependent on the direction of rotation, where is a starting position or, for example, an angular reference point (whatever it is called), how large will the actuator torque be , if a certain, eg weak braking takes place? However, since it does not trigger braking, the process can also be carried out at will, except possibly during braking.

Of course, it is also possible or useful to record an actuation characteristic of the brake (e.g. actuator angle and actuator torque, also with the difference between actuation and release) up to the area of lining contact pressure forces, e.g. when the vehicle is stationary or a normal braking process as a characteristic curve recording use.

To determine losses, it is also recommended that, alternatively or additionally, another known force can be used in addition to the spring: The mass inertia is determined to a large or predominant part by the motor because the proportion of fast-rotating parts is higher with the square of the ratio ( the slower parts can of course also be taken into account). In this way, for example, in a range without any significant lining lift (others are of course not excluded), for example, a certain speed change can be applied over time, the actual behavior can be measured and thus the inertia Torque can be measured, which, however, still contains the mechanical losses in the measured value. If the theoretically necessary moment is subtracted, the losses remain. Of course, this calculation can also be made in any other way that describes the same physics, e.g. time for a specific movement, movement in time, moment and time, etc. Of course, all other physical quantities involved can also be used for inertia-based loss detection , such as the energies (rotation, losses, etc.).

With what has been explained up to here, the distribution of the losses (up to now also referred to as mechanical losses) from the electrical input to the contact pressure on the lining would not be (easily) distinguishable, although the method described with the current-torque relationship is of course very large helps. Therefore, another method is proposed here that can also determine the distribution of the losses between mechanical and electrical: two forces are already shown above that act purely mechanically (other ones could of course also be imagined): the spring and the mass inertia. If now only these act, for example in the de-energized state, then the electrical losses are switched off and one can distinguish between a system with electrical losses and without and distinguish between these two losses. Of course, the question remains as to whether an unpowered motor has no electrical losses, but this does not have to be clarified scientifically, it only needs to be applied in practice. Other "currentless states" can also be used to measure the reaction, such as reversing direction or releasing the brake. Instead of "currentless" states of different currents can also be compared and thus a "currentless" can be calculated. "De-energized" does not have to be exactly 0 either, but can be any suitable value. If the same force times the same distance is covered in a shorter time, proportionally more power is required. It is hereby recommended that something similar is also used to determine the electrical losses (or to determine the division between mechanical and electrical): If a movement with the same energy takes place in a different time, there is a correspondingly different performance and one can choose from at least two such Processes that determine or estimate the losses in various services. Mathematically, this can of course be expanded in such a way that processes with different energies can also be compared can become. "Energy" is only a physically meaningful expression here, other values can also be used with which this principle can be achieved.

If the brakes are now actuated, you will find, for example, an increasing actuator angle with an actuator torque curve and you can always compare how the respective actuator torque (including instantaneous losses) behaves with regard to the spring characteristic, with the spring characteristic in the figure having the opposite sign (the Signs only have to be taken into account correctly or, for example, calculated without a sign in this case). The known non-linear transmission ratio can also be used to draw very precise conclusions about the pad contact pressure, since the losses are also well known. In addition to the "idling losses" there may be further losses up to the contact pressure of the lining, but these can, for example, depend more heavily on contact pressure than are subject to fluctuations (e.g. due to grease viscosity). In this way, they can be easily calculated or calculated out or recognized depending on the influencing variables, as shown below. Of course, the actuator torque curve does not have to correspond exactly to the planned curve, the measurements can also show the broken curve. Then you can see, for example, that the contact point (at which actuator angle the lining comes into contact with the friction surface) is different than planned, e.g. due to lining wear, and a wear adjustment can be requested, for example. If the brake is released, the curve jumps down again by twice the losses, at least assuming that nothing has changed affecting the relevant conditions in the brake, which could actually be the case, for example, when braking without, for example, significant heat or thermal expansion and/or wear and tear has taken place. The losses visible here are not only the no-load losses, but also include all others. What is referred to here as "jumping losses" when the direction of rotation is reversed, in reality takes place within a relatively small change in the actuator angle, especially if the play is "pressed out" of the mechanisms due to the constant load direction (e.g. pad contact pressure) and the play is essentially on the same side issue.

A non-linear brake, i.e. with a gear ratio that varies over the pad stroke, is recommended as advantageous if it has the Lining contact pressure works with an actuator torque that does not vary greatly, because then the torque range in which the spring characteristic is compared is relatively limited. In contrast, the actuator torque with a linear drive (e.g. ball screw) would change extremely strongly from an air gap to full braking. A non-linearity divided into areas is also particularly recommended, as this makes it easier to implement an area without any significant pad lift, for example.

Brakes with several, very different springs are shown in the figures. You can now do everything recommended regarding the calibration spring and the

Loss detection (e.g. in the area without significant pad travel) can of course also be used with any number of springs, because it is always a matter of the correct sum of the moments (at the same point) having the correct sign. Such brakes always have at least the moment that the brake needs to apply, the moment of the electric motor (which actuates) and the moment of inertia. With springs or other energy stores or sources, new moments are simply added with the correct sign and everything mentioned above applies analogously with more moments. For the sum of the moments that an actuating motor has to generate, it is irrelevant how many moments are in the sum.

A rotary position sensor can also be located directly on the actuating motor, e.g. for a brushless DC motor (BLDC). It is also recommended that this sensor can also be used advantageously in such a way that if this sensor fails, operation of the BLDC motor is no longer possible and the brake therefore goes into a safe or desired state, for example.

Finding the position (eg angle) of the actuation snail can still be inaccurate if the snail angle varies depending on the snail moment, which would be the case when using a spring, for example. In this case one would find the position eg at a certain moment or moment range. Alternatively or in addition, it is also proposed to use the fact that the known gear ratio (e.g. gear train of the motor) specifies a connection between motor angle and snail angle and therefore only possible positions and not all others for determining the exact snail angle when finding the Starting position of the snail can be used. For example, the knowledge can be used that the correct starting position of the snail must be at a specific motor angle, but the motor angle may be unknown by integer revolutions, for example, but then the snail angle is very precise when the integer ratio is known related to the engine angle.

For e.g. safety reasons or e.g. for reasons of time (if the above finding of a starting position would take too long) at least one further position sensor can be recommended, e.g. an angle sensor on the actuation-snail.

To further increase the accuracy of the brakes, the following is suggested: Absolute accuracy, especially in the range of weak to normal braking, is necessary above all for the so-called blending, if, for example, a total braking torque has to be composed of regenerative and friction braking and a certain adjustment accuracy is therefore required from the friction brake will. In addition, it is recommended that in the case of quickly observable reactions, e.g. when the blending composition is changed (e.g. when regenerative braking becomes weaker with decreasing speed), one reacts to unexpected deviations, e.g. when the wheel slip changes, although the total braking torque on the wheel remains the same was intended or whether the wheel slip changes differently than expected, for example if the total wheel braking torque is changed. Comparisons between such reactions are also recommended, e.g. wheel slip on at least two wheels. You can of course also use statistics so that you don't change the brake parameters immediately for every difference in wheel slip, because different road conditions could, for example, lead to different slip or reactions for a short time.

For longer periods of braking, it is recommended to also use the following simple physical facts: A friction brake must dissipate almost all of the mechanical power into heat, where the mechanical power is braking torque times angular velocity. This means that you can compare two brakes (e.g. left and right opposite ones) by simply measuring the temperature for the same braking performance suitable installation location, at least somewhere in or on the brake. If the temperatures differ accordingly despite the same or similar assumed braking performance, you can change the settings of the brakes and also use a correction for the future. In principle, any sensible change to the brake setting is conceivable, for example, you can reduce the warmer setting a little in the braking moment and/or increase the colder setting a little, you can also use physical or other values (e.g. empirical values) to determine "somewhat" more precisely or you can also use any kind of determinations (e.g. models) to decide which ones to increase or decrease. Reactions capable of learning can also be favourable, which, for example, learn from success (eg approaching temperatures) which procedure is assessed as favourable.

The actuation snail can also be used in both of its directions of rotation for e.g. different service braking operations: e.g. one direction could come to full braking faster (e.g. emergency braking), but the other direction could need less current for longer, weaker braking or it could e.g stroke (for unworn linings) and the other direction more stroke, for example to be used from a certain lining wear.

A calibration spring placed anywhere (or e.g. the lining release springs) can be used to calibrate the engine torque as above, e.g. in the air gap area.

Different beginnings and courses of the two actuation snails (service brake, parking brake) can also be used to increase the accuracy in an evaluable manner. Another drive can also play a role here, e.g. a cable pull for safety reasons, which only becomes effective if, in the event of a failure, the driver e.g. continues to pull the lever or press the pedal. A cable can also cause or release a parking brake.

It can be advantageous to move the two brake pads with different strokes, so it is proposed that the stroke can also be designed favorably via the movement per pad. This can be advantageous, for example, if the brake shoes develop different braking effects, such as with self-reinforcing drum brakes such as Simplex. The motor of the brake actuator can, for example, be attached to a drum or disk brake, the braked movement does not have to be circular, but can also run in a straight line or otherwise and brake an elevator car, for example.

That the linings are lifted off the friction surface in such a way that an air gap is created.

That an adjustment option for the correct position of the brake pads (e.g. air gap on both sides) is provided or that this adjustment takes place automatically.

Further features according to the invention can be found in the claims, the description of the exemplary embodiments and the figures.

The invention will now be explained further using the example of exemplary, non-exclusive and/or non-limiting embodiments.

Unless otherwise specified, the reference numbers correspond to the following components:

Brake 01, brake disc 011, brake drum 012, losses 016, 1g braking 017 (e.g. g/3 = 017/3), target braking effect 018, wear adjustment 02, spring for wear adjustment 021, slipping clutch 023, driver 025, toothing 026, adjustment lever 027, friction in wear adjustment 028, non-linearity 03, actuating cam 032, roller for it 033, cam axis of rotation 034, spring-loaded support 039, indentation for ratchet advance 0311, cam track round 0321, cam track pointed 0322, cam lift 0323, cam radius 0324, cam radius shifted 03241, flat cam track 0325, allowed gradient 032221, small roller 0331, actuator 04, motor 041, actuating spring 042, motor electronics 043, calibration spring 046, parking brake drive 047, parking brake position 0471, parking brake spring 048, calibration spring characteristic 049, rotatable bracket 0411, measured data from the actuator 0431, pressure 05, expansion part drive 05 052, unbraked position 053, braked position 054, S-Cam 056, expansion pivot point 057 , expansion part rotation axis 0571, connection to the actuation 058, pressing movement 059, expansion part lever radius 0511, rotated Contact surface 0591, non-turned contact surface 0592, friction pairing 06, brake lining 063, driving force measurement 064, brake shoe 067, air gap 068, brake shoe support 069, spring(s) for creating an air gap 07, wear adjustment actuation 08, area used for braking 081, area not used for braking 082 , Fixed part (e.g. wheel bearing part) 09, vehicle stability function 106, position of non-linearity without lining travel 111, wheel suspension 13, contact point with increased air gap 1502, contact point with reduced air gap 1503, increased constant losses 1504, actuator torque in the air gap 1505, increased percentage losses 1506, lining displacement force 1507, stability influencing variables 1603, model input variables 1604, calculation model 1605, manipulated variables 1606, function of time 16051, friction coefficient model 16052, air gap model 16053, stiffness model 16055, other models 16056, service brake 16061, parking brake 16062, wear adjustment 16060, initial position .

Fig. 1 shows a brake 01 in which a friction pairing 06 is pressed by an expanding part 051, with the expanding part 051 being rotated about an expanding part pivot point 057 with an expanding part lever radius 0511 and a pressing movement 059 (right) via the rotated pressing surface 0591 on the non -turned contact surface 0592 causes. The rotated contact surface 0591 is preferably a circular or

be a cylinder segment, the non-rotated contact surface 0592 is preferably, for example, an imaginary surface, but can also use a reduction in friction by rotating, ie, for example, be designed as a rotating roller surface. The pressing movement 059 does not have to run in a straight line, but can more or less well follow an already existing movement, which can result, for example, from the rotation of a brake shoe around a support point or, for example, from the deformation of parts such as a brake caliper. Strictly speaking, the pressing movement 059 describes a curve (or straight line) on which the contact point (the contact line) of the rotated contact surface 0591 moves onto the non-rotated contact surface 0592. In addition, “lateral play” can allow a lateral movement, which in FIG. 1 essentially runs in the drawing area normal to the pressing movement 059 (ie essentially up or down in FIG. 1). The pressing movement 059 will advantageously lie in a plane approximately normal to the axis of rotation 0571 of the expansion part, but can also have a different effect, for example approximately parallel to the axis of rotation 0571 of the expansion part. The rotational movement of the expansion part 051 is supplied by a non-linearity 03 (transmission with a transmission ratio variable over the actuation path), for example a roller 033 can follow an actuating cam 032 and the expansion part rotation axis 0571 can rotate via a lever, for example. For how the lever movement is taken from the cam curve, there are many possibilities in addition to a roller 033, e.g. instead of the roller 033, a part on the lever can slide on the cam or make a rolling movement, so that e.g. a lever surface interacts with the cam curve that they roll off one another (“rolling lever”). There is preferably no further part between the scanning (eg roller 033) and the lever that influences the course of movement, ie the scanning part (eg roller 033) is preferably attached to the lever, mounted or rolling, among other things to reduce costs, installation space, complexity, additional bearing points to save. Parts influencing the course of movement include, for example, rams, pulling or pushing devices. Fastening parts such as bearing bolts in the roller 033, rolling elements, bearing rings are of course not affected.

The non-linearity 03, e.g. the cam axis of rotation 034 (or e.g. a toothing 026 on the cam or e.g. a driver 025) is driven by an actuator 04, which in turn consists of an electric drive and other components, such as other non-linearities 03, and energy stores such as springs can exist, which can also be arranged structurally separate from the electric drive. The electric drive is preferably operated by motor electronics 043, which can also measure motor data (e.g. current, torque, position, etc.). In an extreme simplification, the actuating cam 032 can also be the same component as the expanding part 051 and thus the roller 033 can also have the same component as the non-rotated contact surface 0592, which in this case has the same component as the rotating roller surface 033 and, through the rotation of the roller, a compensating movement between the rotated contact surface 0591 and the non-turned contact surface 0592 is particularly low-loss and low-wear.

Fig. 1001 shows the effect of Fig. 1 in a greatly simplified manner, with, as is often the case, two expansion parts 051 forming the expansion part 051, which actually acts as a whole (the entire expanding part is always referred to as the expansion part 051): one as "fixed" Assuming a fixed part 09, the spreader part 051 ultimately presses the brake pad 063, if necessary after overcoming an air gap 068, onto a brake disc 011, brake drum 012 or whatever type of friction surface (e.g. rail), with arrangements on both sides naturally affecting the action and reaction force use, are more advantageous, for example instead of acting on part 09 assumed to be “fixed”, one could also directly or indirectly act on another friction pairing 06, which is indicated by the lower arrow on friction pairing 06.

Fig.2 shows a brake 01 similar to Fig.1, but here, for example, with a double-acting expanding part 051 (a single-acting one would also be possible), which here also has different expanding part lever radii 0511 (top, bottom), but above all a wear adjustment 02 is supplemented: A non-linear brake 01 can do without a wear adjuster if the wear can be covered with the range of motion of the non-linearity 03 or a wear adjuster can also act differently than in FIG. Fig.2 suggests that a wear adjustment 02 (bottom) can be, for example, in the rotating drive of the expanding part 051 (which could in principle adjust the greatest wear), but also, for example, a wear adjustment 02 (middle) between actuator 04 and non-linearity 03 ( which could fit together, for example, with a brake 01 that becomes stiffer as it wears), but also, for example, the entire actuator 04, possibly also with the non-linearity 03, can be changed in position for a wear adjustment 02 (above), e.g. rotated, whereby of course preferably only one of the three wear adjustments 02 shown will be present. The actuation of wear adjustment 02 is preferably derived from a brake actuator movement, with the division of the actuator movement into pad contact pressure stroke or wear adjustment 02 also being referred to here as non-linearity 03, so that this brake 01 preferably has an additional non-linearity 03 for wear adjustment 02. The pivot point of the expansion part 057 can be unsupported (caused by the rotation of the expansion part as an apparent point around which the apparent radii rotate) or the axis of rotation of the expansion part 0571 can be “fixed” or “floating”, with the bearing forces preferably being smaller than the contact pressure. A simple, cost-effective approach is shown in FIG. 3, the figure shows a schematic of possible drive methods of a corresponding braking system, eg for trailers for bicycles or agriculture with a wheel suspension 13, which can also be designed as an axle suspension and can also have a spring deflection determination.

Here, both brakes 01 (e.g. for brake discs 011 or for brake drums 012, preferably both the same) are actuated by a common brake actuator via a mechanical connection. The actuator 04 can be designed, for example, as an electromagnet or linear drive (top), rigid electric motor 041 (middle) or electric motor 041 (bottom) with a rotatable mount 0411.

The brakes 01 are mechanically manufactured and adjusted in the same way so that the connection to the actuator 058 ensures the same braking effect on both sides. A brake 01 that acts more strongly would again become similar to the other brake 01 due to increased lining wear. Of course, an entire group of axles could also be actuated in this way by, for example, just one brake actuator, preferably with axles lying close to one another being synchronized in this way, and for example two of the upper axle assemblies being given mechanically linked actuation.

An electric motor, electric linear actuator or actuating magnet can force-control the contact pressure 05, i.e. even if wear is not adjusted (e.g. without an additional wear adjuster), the actuating force would bring the brake 01 into the correct braking force position. A parking brake position 0471 could occur e.g. after exceeding a lever dead center or a spring action, or both.

A drive with a configurable non-linearity 03 can be seen in the middle, here an actuating cam 032, which acts as a self-resetting service brake in one direction, for example, and has a stable parking brake position 0471 in the other direction, for example a depression or flat area. Of course, the parking brake position 0471 could also be omitted or, for example, follow the end of the service brake positions. On the one hand, this cam can be shaped in such a way that it covers the expected wear caused by the lift and rolling process. However, the brake 01 can also be designed to be particularly stiff, ie it requires a relatively small actuation stroke up to full braking in comparison to the wear. A common wear adjuster can be on the connection to actuation 058, for example. In this way, the course of the cam can be optimized or designed in any way, since the cam always works together with the correctly adjusted brake 01, at least within the tolerance range of the wear adjustment 02. By reversing the direction of rotation of the motor 041 (e.g. DC motor), it can be decided whether the service brake range or the Parking brake area is to be actuated. With all these simple motor or electromagnet controls one can take advantage of the property that the motor torque or an electromagnet force is approximately proportional to the current and therefore the "controller" described above can operate this motor 041 or electromagnet directly with its current control or PWM. It doesn't matter whether the "controller" is in the towing vehicle or trailer, because both are coupled together.

The variant with the rotatable motor mount (or another geometrically variable component in the drive of the actuating cam 032), together with the spring-loaded support 039 and the cam profile, means that a favorably designed brake actuation is possible despite wear (possibly without an additional wear adjuster). The actuating cam 032 can, for example, be designed in such a way that a required braking effect is still possible in a required time with the maximum permissible wear. This means that the actuating cam 032 will initially, with a large air gap 068 due to wear and tear, run steeply in order to make a quick stroke in this low-power operation. However, it would be too steep to build up higher power right from the start with a much smaller air gap 068 (fresh pads). The resilient support 039 can ensure that the actuating cam 032 can deviate from the steep start and continue rotating to a less steep area. Unfortunately, this means that the driving torque of the actuating cam 032 cannot be kept constant, because the supporting spring specifies how far the evasive movement will be, but the supporting spring can at least ensure that there is no high driving torque that cannot be actuated. A wheel or axle load averaging, which detects a position, a path, an angle or a force, for example, can be used to support the braking effect control. Since the brake requires a supporting torque against the braking torque, a supporting torque, a supporting force or a position can be determined or the change in the above wheel or axle load averaging can be used to determine the supporting torque or braking torque. A motor or generator torque of the vehicle drive motor can also be used together with the friction braking effect to determine the friction braking torque: if, for example, falling generator torque is to be compensated for with increasing friction braking torque, one can observe the possible reactions to determine whether the two torques are how behave as desired, i.e. whether the wheel or vehicle deceleration reacts as desired, the deflection of the wheel or axle, in principle any expected change can be used as a comparison for the correctness of the friction braking torque and used to correct the friction braking torque or to correct a wear adjustment.

In a more complex variant, each brake 01 in the above vehicle can have its own actuator actuation, e.g. by assigning a common actuation variant from the above figure to each brake 01. How to achieve uniform braking despite individual brake actuation is described below.

An easily accessible adjustment option is also proposed in such a way that the motor 041 of the actuator or the actuator itself can be adjusted on its holder in such a way that the wear is adjusted manually, from the actuator itself or in some other way. For example, the actuator or motor could have a pivot point and a slot and screws could be loosened for adjustment and then screwed back on to secure the position of the actuator.

In Fig. 4 it is proposed how the current air gap can be advantageously determined in order to derive the need for wear adjustment 02, e.g. after comparison with a target value of the air gap.

The wear adjustment has completely different requirements for a non-linear brake than for current, power- or pressure-actuated brakes. In Existing designs practically always use a linear or almost linear contact pressure, which means that errors in the wear adjustment do not cause errors in the contact pressure as long as this can still be generated due to the possible stroke. In the case of non-linear contact pressure, the brake must always be operated in a selected part of the non-linearity and the behavior between the actuator and the contact pressure nevertheless changes at every point, which must be taken into account. Special requirements such as accuracy and reproducibility are therefore placed on the wear adjustment of the non-linear EMB, which also affect the precise design of the non-linearity in order to be able to operate the EMB with the desired properties.

For the readjustment, it is suggested that this be carried out, for example, with an electric motor, which can be a separate motor or an existing one (e.g. brake actuator) or a manually operated readjustment or even the omission of an readjustment. For the implementation of the wear adjustment, there are many suggestions for mechanical design variants, such as screws.

The proposed air gap detection could thus identify a need for readjustment and cause it to be carried out immediately or with a time delay. In the case of a manual adjustment, for example, a corresponding message could be generated. Or, in the event that no adjustment is made, the linear pad movement required to overcome the measured air gap can be included in the calculation of the brake actuator movement. Mixed variants are also advantageous. For example, small readjustment movements can be taken into account by means of an appropriately adapted actuator movement and only a larger readjustment requirement can actually be readjusted (e.g. to increase the service life of the adjuster). Changes due to temperature fluctuations, for example, can be prevented by wear adjustment, however carried out.

The need for adjustment can be determined in many ways, e.g. by decreasing braking effect or contact pressure and automatic or manual adjustment, by determining the force or moment, which can function as desired, e.g. mechanically or by electrical determination.

A sensory detection of the contact between the brake lining and the friction surface is also possible and is also known, for example, in the truck sector, but is expensive and potentially vulnerable. Here it is advantageously proposed to use sensors that touch detect indirectly and can be arranged in areas of the brake where they are protected from environmental influences. Examples of corresponding measured variables are vibration or sound waves. The use of electrical conductivity can also be proposed, for example by means of conductive material introduced into the lining material, which causes a current when the lining comes into contact with the friction partner.

In FIG. 4, a particularly advantageous determination of an adjustment requirement by means of a moment or current measurement on the brake actuator is proposed. 4 shows that the force for shifting the lining (lining displacement force 1507, left y-axis) in the area of the possible air gap 068 (the lining movement is on the x-axis), which can come from mechanical losses or a spring, for example , is in any case very small and, also because of the flat course, not very meaningful with regard to the beginning of the contact pressure, especially if the measuring device is designed for the maximum (full braking) contact pressure.

Due to the non-linear translation to the brake actuator, the actuator torque (right y-axis) (or the torque-generating current) shows a much more meaningful course. For an advantageous use of the data that can be recorded, it is hereby suggested that mass inertia effects, friction losses and e.g. influences such as temperature, speed or age be taken into account in such a way that the most precise possible connection between the measured current and the effective torque is established.

Assuming the air gap 068 in Fig. 4 would be a correct air gap and was recorded, for example, when touching the lining or when braking lightly. With increasing wear, the contact point would migrate to the contact point with an enlarged air gap 1502 because the lining only touches the friction surface with a longer stroke and the brake actuator torque is smaller here due to the other non-linearity. The contact point with a reduced air gap 1503 would indicate that the air gap is too small (e.g. due to temperature or due to an excessive previous wear adjustment process) and the brake actuator torque can be higher due to "faster pressing" non-linearity.

The actuator torque characteristic can now also change due to other effects.

It can shift upwards, for example because of viscous, cold grease in the engine transmission, which is shown in curve 1504. There may also be losses increase by percentage, further increasing curve 1504 to 1506. The expected air gap 068 would therefore require a higher actuator torque 1505. Now, given all these possible influences, it is no longer possible to work out in advance why an observed shift to the actuator torque 1505 has occurred, whether due to changes in the air gap or for other reasons, because there are far too many variables.

As a first solution, it is suggested that the course of the torque-displacement (or - angle) curve be determined (measured) at several points and calculated whether a shift in the x-axis provides a good explanation, which would correspond to a wear adjustment suggestion.

Constant changes in the losses (e.g. viscous grease) have a particular effect on small actuator torques and the following estimate is suggested here: in an actuator area in which there is not yet any contact pressure, the brake actuator torque just determined is compared with an expected one. Of course, this can also take place several times and also in different directions of rotation, and a known temperature variation can also be taken into account. Now, for the first correction method, this determined basic displacement of the actuator torque is taken into account and, according to the above procedure, that the x-displacement is assumed to be the cause, one comes to a good conclusion. In addition or alone, one can take into account how quickly the brake actuator torque curve increases, which is caused by various, location-specific, non-linearities and indicates at which point of the non-linearity one is and can therefore also be interpreted as an x-shift.

In the case of the above-described method of the type in which the engine mount turns away in the event of excessive loading or some other compensating movement takes place, this movement can also be used or included for wear detection. Measured or determined movements, forces or moments can also be included, such as a pad gripping force or a gripping effect when weak braking begins.

A wear model (based, for example, on temperature, braking torque, speed, braking work, process such as emergency braking or landing, etc.) can also be carried along in order to be taken into account in the wear adjustment. The wear adjustment can also take into account values of other brakes, such as a temperature of a brake located on the other side of the vehicle, and the brakes can be set or actuated, for example, in such a way that the same or similar values are set on both sides. A guide can also be used for this purpose, so that the measures to increase accuracy do not leave the permissible ranges, or the wear adjustment can be carried out in such a way that the measured values (eg temperature) on both sides approach a model value. Of course, fundamental inequalities between two brakes, such as reduced braking on one side due to ABS, would be factored out.

FIG. 5 shows an example of how a brake control recommended here as advantageous can be constructed, with functions being able to be added or omitted and the order in which they are run through can also be different. It is therefore a fundamentally possible functional description.

A target braking effect 018 is assumed, which can come, for example, from the driver, pilot or, for example, from an automatic device. It is recommended that the target braking effect can (but does not have to) be pre-treated, e.g.

To determine target braking effects of individual wheels, which can be done in a vehicle stability function 106 with e.g. characteristic curves and where other influences such as "blending" can also be treated and measurements of e.g. wheel speeds, steering angle, yaw rate, etc. can be included as stability influencing variables 1603.

The large block for the calculation model 1605 shows how the actual brake control generates the manipulated variables 1606 for the brake actuator from a setpoint braking effect or the result of a vehicle stability function 106, with 16061 for example being the control of service braking, 16062 for example a parking brake function control, 16063 for example a wear adjustment control, 16064 moving to a starting position, etc. The function is shown here using a single EMB, but of course a system as shown above could also serve several EMBs.

The special feature of this advantageous model is that it is assumed that it is impossible to store characteristic curves (eg stiffness characteristic curve) and values (such as the instantaneous coefficient of friction) beforehand because both From the start of braking as well as from the end of braking and also for all subsequent braking operations, the states in the EMB result as a function of time 16051, braking power (braking torque * angular velocity), thermal cooling resistances and heat capacities. Without heat capacities, the problem of previous storage would be "only" multi-dimensional, because each input variable in the storage creates a new dimension for all stored values, which, for example, with a fifth input variable instead of only four, causes a huge increase in storage space. If, however, a development over time occurs due to heat capacities, an additional storage would now have to take place in addition to the multidimensional storage for every other possibility of the development over time, and not just for braking, but for all subsequent cooling phases and new brakings again as additional storage. The function of time 16051 shows that, for example, temperatures in a temperature model develop over time (depending on braking power and e.g. speed-dependent air cooling as well as possible radiation cooling, "black body radiation") and this model feeds a (also) temperature-dependent coefficient of friction model 16052 as well as e.g calculates the calculated air gap 16053 with regard to temperature (the air gap can also be calculated alternatively or additionally using a wear model, for example), can use the current (e.g. estimated) air gap 068, can, for example, take into account thermal stiffness changes 16055 and of course can operate other models 16056. Measurement data from the actuator 0431 can of course flow into the calculations 1605, e.g.

It is therefore proposed here to build up the brake control or regulation advantageously from models in which the development as a function of time 16051 is determined from time and input variables.

In Figure 5 it is assumed that the brake actuator is viewed as a single, purpose-fulfilling actuator, which consists of at least one actuating component in the structural implementation, but can also be made up of several, such as double windings for safety reasons, several motors , Also for different functions such as parking brake or service brake or common functions, such as that the parking brake motor also the service braking could take over, but can also use stored energies such as eg from at least one spring, also via other non-linear gears.

Since the physical property of the electric actuator is at stake here, the manipulated variable for an actuator can basically be position (e.g. motor shaft angle) or moment or force and, of course, composite values such as angle and moment. In the case of the composite, it is recommended that, for example, the control or regulation 1605 shown above adjust the torque via the current of the motor 041 and at the same time ensure that the angle of rotation of the motor 041 remains within an allowable range, both from the above models (or with a similar effect) is determined, whereby this is of course only one possibility among many to control or regulate the actuator, since measurement data from the actuator 0431 (such as actual values of e.g. current, torque, angle, voltage, temperature, etc .) go into the large block 1605, i.e. into an electronic system.

Figures 601 - 603 show a floating caliper disc brake (unbraked in Figure 601), in which the inboard lining is pressed via a cam-like spreader part 051, for example, as is also known, for example, as a spreader part in mechanically actuated drum brakes. The EM expands when tightened and bends, as shown in an exaggerated manner in FIG. The cam-like expansion part would possibly perform a "scratching" movement on its two contact surfaces, because its rotation results in a height difference (between the unbraked position 053 and the braked position 054) and also a rolling movement on its surfaces. On the one hand, this expansion part can be designed and installed in such a way that its “scratching” incorrect movements correspond as far as possible with the incorrect positions caused by deformation of the brake parts, thereby compensating for them. Remaining errors in the heights can be caught in play and displacement, as indicated, for example, by the tilting of the wear adjuster. Since high surface pressures occur on the expansion part, hardened surfaces are desirable, as shown, for example, in the variant in Fig. 603 with the pressed-in, hard pins with any desired cross-section. Of course, all other methods of spreading can also be used, such as, for example, ball ramps, also with a variable pitch or variable, for example spiral, track and multiple ball ramps. In FIGS. 701 and 702 are different rolling bodies, which mostly use a segment of a circle as a rolling surface, but could of course be arbitrary or, given small dimensions, could also have imprecise small contours due to the production process. It would be advantageous to use needles or rollers (e.g. press-fitting into bores) from roller bearings, for example, in order to achieve hardness, good circularity and cost-effectiveness. The other rolling surface will usually be a straight line (Figures 701 and 702 above), but could also be different (Figures 701 and 702 below) and will deviate minimally from the original (eg straight) due to the effects of use. If this expansion part is turned from the left position (Fig. 701) into a surface-pressing one with the expansion part pivot point 057 (Fig. 702), there are several processes: An xy sine-cosine movement describes the circular path of an initial contact point, whereby a lot of x (in the direction of contact pressure ) and little y (height deviation). In addition, rolling around the circumference of a circle creates a path proportional to the rolling angle. With a 360° roll rotation, the entire circumference would be rolled off, here only a rolling segment proportional to the angle. This unrolling causes more y than x movement in the drawing. These movements can never be height-compensating because one height difference goes with an angle function and another angle-proportional. If rollers do not roll in a circle and/or rolling surfaces are not level, this could bring advantages in terms of a height error, but disadvantages in terms of price. In addition, there can be an error that, by definition, a point of contact must always have the same tangents on both curves that are in contact, and this would therefore also have to be taken into account with regard to a height error.

If, for example, 6 mm needles are 15 mm apart, a lever length of 45 mm would have a transmission ratio of 1:3 and would transform a 2 mm stroke into a 6 mm stroke and make a pivoting angle of approx 0.19 mm with a 19 mm roll circumference and ±3.6° and 0.03 mm height error from the circular movement.

You can now operate such a pressure lever with regard to its rolling geometry in the area of minimum height error, which would mathematically be a specific area of a cycloid. But you can also see the acting forces, movements and the Emphasize production possibilities: front wheel disc brakes in cars act up to 35 kN, for example, and up to 240 kN in trucks, for example, with contact pressure strokes of 1.8 mm (cars) being made. If, for example, you choose a roller diameter of approx. 6-8mm (car) because of bending and surface pressure, the rollers could be ground down to bring them closer together, but you will not always easily reach the mathematically optimal range of the height-optimal cycloid path. Practically, an approximation to the minimum mathematical height error results in a geometry that is difficult to manufacture with small rolling radii that are close to each other and where a force-transmitting connection of both rolling radii is geometrically difficult because the connection can be thin to connect through the middle between both rolling radii.

In Fig. 705 you can see the expansion part with expansion part pivot point 057 and the thick circular parts (which represent the pressing of the expansion part). The thick circular parts press on the two thick rectangles, which are not rotated with the spreader. The expansion part pivot point 057 could be mounted, but in Fig. 705 it can also be rotated without a bearing, since the expansion part between the thick pressure surfaces, which are shown here e.g. rectangular, essentially cannot leave the position.

In FIG. 705 there is a rolling pairing operated mathematically close to the optimum of the cycloid, with the thick circular arcs rolling on the thick corners. When turning clockwise, a point of support would move further up due to the angle function. The rolling circumference at the circular arc would also roll up. The support point does not remain at the same height, but both movements are similar, which means that little or no relative movement (“scratching”) is necessary. The two arcs could be connected between the rolling corners, which already gives little material in the area connecting through the middle. These roll-off curves with a radius of 4 mm, for example, have to be manufactured with uncomfortably precise results. If holes are now drilled to insert pins (dashed circles), the connecting material is largely drilled away and the rolling surfaces must be left free for the pins. These are good reasons to refrain from being close to the mathematical optimum. In the case of this opposite interpretation, one will select a location of rollers of suitable diameter that is favorable in terms of manufacturing technology and forces. You can either accept the height error and, if necessary, assume that unwanted movements or deformations are taking place, e.g Braking takes place, for example, at 14 to 1/3 full braking deceleration). Or you can use movements or deformations that unavoidably occur when the brake is actuated by allowing height errors and other movements to act at least in the same, compensating direction, or preferably by designing them in such a way that height errors and other movements compensate as well as possible. This “other” movement occurs with drum brakes, for example, when the brake pad carrier that is pressed on moves (e.g. around its bearing point) or when the calipers of disc brakes deform under contact pressure, ie widen and bend, for example.

Incidentally, scratching movements during braking can even make less of a difference than, for example, constant frictional movements caused by vibrations, e.g. from an unbalanced wheel or diesel engine, and thus (e.g. partially) allowing height errors that cause scratching movements is quite possible and can be significant for production and costs bring benefits.

FIG. 8 shows how a contact pressure force can be generated as close as possible to the lining contact pressure or the wear adjuster placed in between. Inserted or otherwise attached or secured (clamped, welded, screwed) parts are dashed as non-turned contact surfaces 0592 (also with special properties such as hardness, wear resistance) and black are inserted needles or otherwise attached or secured (clamped, welded, screwed ) Parts of the turned contact surfaces 0591 of an expanding part (also with special properties such as hardness, wear resistance). The geometry of the rolling of the black needles on the gray areas is preferably designed in such a way that the parts can be manufactured sensibly, but that errors in the rolling movement are, for example, small or such that they can be compensated for or tolerated by play, deformations, displacements but also preferably act so that deformations when pressed have the same effect as the errors and therefore compensate each other as far as possible. Here, for example, the length of the arcs unrolled by actuation could be chosen in comparison to the angular function movement of a point at a needle in such a way that the elevation of the dashed unrolling surface (right) can be compensated. Residual errors are caught here, for example, by tilting the part that presses the lining.

801 is a possible embodiment with a lever with a roller 033 for the cam 032 in FIG. 8 and two ends for two pressings, e.g. as an expanding part 051, which can be on both sides of the wear adjuster, so that the wear adjuster has space in between . Each of the two pressing ends can, for example, use the needles, rollers or other pressing parts on both sides, so that four synchronized pressings are created here, for example. Of course, the counter surfaces for the contact pressures must also be positioned accordingly and often present. This lever can also be assembled, for example, from parts such as strip steel, sheet metal, etc., for example welded (indicated as a spot weld in Fig.801 in the corner in the text "Fig.801"), spot welded, riveted, screwed, glued, folded and use bending joints etc.

9 shows an actuator torque displacement behavior of a realistic EMB with the pad travel on the x-axis and the actuator torque on the y-axis. which, as described in this work, was designed in such a way that the smallest possible starting radius of a cam was combined with a roller diameter that can withstand the pad contact pressure and the torsion angle of the cam fits into the geometric conditions in the EMB. Under these conditions, the actuator torque over the brake actuation is by no means even approximately constant.

The two bold curves (dashed and solid) are for a correctly set air gap, the dashed line is for a fully worn brake pad, all others for full pads. It is proposed here not to save force-deformation curves, but to generate them dynamically from a model in the brake control, because these curves would be output by the model for certain temperatures, which in turn depend on the time history of the thermal power of the braking to which the model reacts .

It is now also proposed to output the force-deformation curves as force-coverage lift characteristics. Shown above (thin solid lines) is that the Air gap, for example, 0.1 mm smaller (above the bold line) than could be provided or 0.1 mm larger (below the bold line). Of course, one could try to set the air gap as precisely as possible in order to eliminate this influence. Here, however, it is recommended to also determine inaccuracies in the actual air gap size, since an air gap setting (or wear adjustment) can, for example, be subject to tolerances, the contact point can only be determined within the scope of possible accuracy, adjustment can only be made in certain steps (e.g. ratchet progress) or other effects can lead to deception. This includes, for example, abrasion that accumulates on the friction surface of the linings and remains on the friction surface to an unknown extent or is transported out again. It is therefore proposed to allow such an almost abrupt change in the air gap to a certain extent, even if a pad wear model would not show such abrupt wear.

Such accumulating abrasion can also change the stiffness characteristic of the brake if, for example, only parts of the pad surfaces are affected. The rigidity can also be subject to larger manufacturing tolerances (e.g. casting material, geometric casting tolerances), change over a longer period of time (e.g. material thickness reduction due to e.g. corrosion) and change thermally, e.g. if stresses develop in the material due to uneven temperature distribution. These influencing variables are preferably included in the stiffness model here, which also speaks against mere storage.

One of the many options suggested here (whereby, for example, parts can also be used), for example first in a non-braking area, which can also be an area 082 not used for braking, for example, is to determine the actuator torque, for example in order to determine the instantaneous mechanical losses (e.g. due to gear grease temperature). Then it is proposed to determine the point of contact by increasing the actuator torque (in relation to the instantaneous mechanical losses and the local non-linearity) before a noticeable braking torque occurs. For this purpose, actuator angle and torque measurements can be made and these can also be statistically evaluated over the large number of measurements, eg averaged. Even with the still weak, increasing braking, it is proposed to determine the gradient or the behavior of the braking stiffness. This could also have happened, for example, during a preceding braking process, but it will It is recommended to determine the slope or behavior of the braking stiffness even without prior braking. The stronger the braking increases, the more statistical evaluations and the better measurable actuator torque can be used to determine the gradient or the behavior of the stiffness characteristic more and more effectively.

Additionally or alternatively, the brake can be controlled or regulated via the pad contact pressure force, which is calculated from the measurable engine torque and the non-linearity, preferably taking into account the mechanical losses and inertial effects. Overall, it is therefore also possible to improve the model of the instantaneous stiffness, whereby other measured values or calculated values can also be included, such as the mechanical work used for the actuation (or released during the release). If springs are involved in the brake, they must be included in the calculation with the correct sign and according to their current effect, e.g. expressed as spring torque.

It can be seen in FIG. 9 that the actuator torque changes significantly. Here it is suggested to make use of the motor characteristic of the actuator. It is recommended here that the increase in speed when the actuator torque decreases is used to shorten the actuation time. The curve above shows that the actuator torque is significantly below the maximum over larger areas and this behavior is used here (or brought about in the design of the non-linearity) so that the actuator shortens the actuation time at higher speeds, although it is not in the Point of maximum shaft power is operated.

An advantageous method is shown in Fig. 10, as can be obtained from measurement data from the actuator 0431 (which can be recorded as an angle and moment on the brake actuator, for example, but any similar representation would also be possible, since there are mathematical relationships between values at different points) statements can be made about the brake, e.g. to determine the current state of wear, the need for wear adjustment or a more precise estimate of the contact pressure. In FIG. 10, the pad travel is on the x-axis and the actuator torque on the y-axis and, for example, full braking with 1 g is achieved in 017 and, for example, “usual” braking with g/3 in 017/3. The thin line is the expected behavior of the brake, which can be stored in the EMB-ECU, for example. However, it can also be determined depending on the situation, for example for "actuating" when the air gap 068 has been assumed to be correct, for example for mechanical losses determined in the past. However, as already shown, the expected behavior can also prove to be non-storable because it can depend on the development of temperatures that cannot be saved in advance, i.e. the development of temperatures depends on current conditions such as current braking power, cooling conditions, etc. and these must be measured in this procedure and/or continuously modeled as a function of time.

The bold measurement data from the actuator 0431 are initially marked as "conspicuous" because they are first above the expected behavior (full curve with air gap 068) and then (with more stroke) below it. With this assumed result of many possible measurements, it is shown that it is possible to identify individual "faults" (detected in a comparison or an evaluation in the comparison or evaluation block 1608) individually and also, for example, in different time horizons, for example so quickly that "errors" can be recognized (or already reduced or compensated) before braking or before an unfavorably incorrect braking effect , whereby the quickly recognizable evaluations can be described as "quick acceptance" 16091 . With more (e.g. statistical) effort, a more precise analysis of the brake properties is also possible, which of course requires more data and time and is therefore presented here as a slow assumption 16092 and of course also has the purpose of improvement.

In a "quick assumption" 16091 it is suggested here, for example, that if more than one measuring point in the area of the air gap is too high, one assumes that the instantaneous mechanical losses are higher than assumed in the target curve. This can also be expressed, for example, in an absolute or percentage correction number. With more actuation than, for example, up to the expected start of contact pressure (e.g. end of the air gap), the points are then below the target curve and, supported by the later increase, one can assume, for example, in a "quick assumption" that the air gap is larger than expected, for example . The assumption is also secured, for example, by the fact that the points are largely below remain the desired course, which can be attributed to the flatter course of a cam, for example. As a reaction to various findings, corresponding correction values for individual parameters of the calculation algorithms (e.g. size of the air gap) can be taken into account in all subsequent calculations in the brake control electronics.

With this method, you can make a "quick assumption" 16091 and also secure it if necessary, which of course is based on the fact that the engine torque develops differently if you use other areas of non-linearity than intended. If, for example, weaker than usual braking below g/3 is desired, this method can be used to make "quick assumptions" beforehand that prevent or reduce an unintentionally incorrect braking effect. The more different states of the EMB are available for measuring point determination, the better analyzes of the states and causes of deviations in the EMB can be made. For example, with different actuator loads, angles or speeds (incl. sign) you will compare with the corresponding target curves, because e.g. the mechanical losses can be different depending on the actuator speed and direction of rotation, for example, or can have different effects. A particularly advantageous situation for the collection of high-quality measuring points arises when the service brake actuator is also used for the parking brake function. Approaching the parking brake position involves an actuation path that is significantly greater than in the majority of service braking operations. In addition, there are significantly lower requirements for the actuation speed, which means that, for example, the influence of mass inertia can be minimized.

Based on the need for wear readjustment recognized above, for example, readjustment can either be carried out at a favorable point in time or, for example, you can continue to operate the EMI for the time being with this not (completely) correct setting. As a result, you can carry out a "slow evaluation", for example, which uses better statistical methods (e.g. averaging) to determine the actual deviation status of the EMI or, advantageously, can also distinguish between several causes of the deviation. For example, one could tell apart that the mechanical losses in the EMB are statistically higher than expected or that the wear adjuster, for example, is statistically set a little too far away and these results can of course be taken into account or saved in the brake control or output, e.g. as a warning. Influences due to, for example, plausibility or, for example, impossibility can also flow into the above method, such as the fact that at a similar temperature of a gear grease is not to be expected, that the mechanical losses have changed significantly from one operation to the next, or that, for example, a "quick assumption" is impossible for an incorrect air gap, because due to a wear model, for example, not that much wear is possible. Of course, these are just examples of many useful options.

It is particularly advantageous if, in addition to the actual data of the measuring points, i.e. cause/effect pairs, e.g. motor position and current, additional information (e.g. current temperature) is recorded and saved as metadata. As a result, it is advantageous for "slow" evaluations to categorize all of the recorded measurement points according to various criteria. Examples of this would be low/high temperature or low/high modulation. If the analysis of deviations of the measured values from the expected values reveals differences for different categories, more detailed interpretations are possible. If, for example, in the example above, there is a horizontal shift in the curve, particularly at high temperatures, an incorrect assessment of the thermal expansion can be assumed. If there were differences between low and high motor positions, on the other hand, an error in the stiffness curve depicting the behavior of the brake would be suspected.

Figure 11 shows the design using a roller and cam as an example, in which non-linearity for “largely constant actuator torque” is deliberately not used and instead the change in the transmission ratio is greatly restricted in favor of other advantages. While other designs require a mathematically justified optimum, the aim here is mechanical optimization in order to be able to use ratios with a given behavior (e.g. lever combinations) or a behavior that can be designed within limits (e.g. gear pairs with a non-constant radius, ball ramps, cams) to advantage.

Due to the limitation of the ratio of the minimum to the maximum power or torque transmission, preferably below 1:20, the motor can no longer be operated in an optimum manner essentially over the entire actuating stroke. He runs rather in wide (and always traversed) actuation stroke ranges strongly deviating from the optimum and could of course assume all possible load states in areas of the actuation stroke, ie from zero to maximum shaft power. Here, among other things, it is also proposed to use its sensible speeds in a wider range, eg in ranges of efficiencies accepted as “good”.

By leaving the operating optimum, it is possible, for example, to design the cam track in a mechanically favorable manner, e.g. without sharp points, without points with a small radius and high stress, without points that are difficult or impossible to machine due to angular conditions and could tend to "self-locking". , eg when the angle of the roller lever is roughly perpendicular to the cam tangent. A roller for rolling on the cam track can thus have a larger diameter and thus carry greater forces. In addition, the use of other non-linear components as cams is possible, since e.g. gear wheel pairs with a non-constant radius or ball ramps can only be used if the variation of the transmission ratio is limited, which can also be further reduced, e.g. to below 1:10. This is based on a condition that has been determined to be favorable from a mechanical engineering point of view, e.g. a minimum roller diameter, which is derived from conditions such as roller and cam strength, width, number of operations and force spectrum. A cam shape is then determined that is also classified as permissible from a mechanical engineering point of view, i.e. does not fall below minimum radii for reasons of material strength, for example. This ultimately results in the achievable non-linear translation. The operation of the electric motor “essentially continuously over the entire actuation stroke at an optimal operating point” is not pursued in the present interpretation and can even represent a contradiction because it corresponds to an impossible requirement.

The transmission ratio of a combination of roller rocker arm or rocker arm and cam can be specified, for example, as the ratio of the rocker arm angle to the cam rotation angle. The rocker arm torsion angle is created by the center point of the roller. In FIG. 11, a desired movement of the center point of the roller 033 rolling on an actuating cam is shown in broken lines, various associated roller positions are shown in broken lines. However, the cam surface is formed on the circumference of the roller as a curve drawn in bold "with a loop". However, in the example shown, either the roll is too big or the center curve has one too small Change radius in the "kink", in any case points of the cam surface arise, which would move away from each other during production and are not possible. It is also not possible to simply "round out" this surface area, as this would produce a different transmission ratio from the one required. According to this design method, it remains to bend the dot-dash center curve in a larger radius (dot-dash center curve on the right), which according to the design results in the transmission ratio, which fundamentally differs from a largely constant actuator torque.

Even if the cam surface curve no longer contains any impossible points, it is still necessary to check whether the resulting radius of the cam surface is possible according to the requirements or whether it needs to be increased. According to this interpretation, one would arrive at the midpoint course shown on the right, for example, and from this one can determine the resulting transmission ratio, which no longer allows too large changes. The same also applies to other rolls, such as ball ramps, and similar restrictions on a maximum possible change in geometry in practice also apply, for example, to gears or friction wheels with a non-constant radius, in which, for example, manufacturable tooth geometries or possible rolls at all points on top of each other ( without positions "in each other's way") must be taken into consideration.

The following advantageous methods for achieving a favorable cam surface can also be suggested, which can also eliminate "too small radii" and "loop through impossible points":

The cam twist angle can be increased because the points on the cam surface are "pulled apart" and can find better places. Although this increases the transmission ratio, it can be compensated for by a lower transmission ratio in the upstream motor gearbox. Similarly, the inner starting radius of the cam can be increased, which also "pulls" the points apart. But you can also partially pull the points apart, for example twisting the starting points of the cam so that loops are pulled apart, i.e. eliminated, and that too small radii are enlarged. This can lead to quite good solutions, but initially changes the gear ratio and cannot be fully compensated for by changing the gear ratio of the motor gearbox. Fig. 1201 shows how an actuating cam 032 with an angle of rotation of approx. 270° (thin) changes an initially very large gradient into a flat one and how it can still keep the mechanical load low at the "round" transition point of the cam track around 0321 because the cam is still "round enough".

If, however, the angle of twist is to be reduced (shown in bold) with the same radii determining the cam lift 0323 (beginning and end radius, dashed, the lift 0323 in between), the transition point would have to be designed with a smaller rounding radius or even as a pointed cam track 0322. However, up to the "kink" of the pointed cam track 0322, almost half the covering stroke has already been covered here. In order to comply with a permitted minimum fillet radius, the non-linearity must be designed in relation to the geometry, e.g. up to roughly half the pad lift. With a reduced torsion angle, this leads either to a smaller achievable maximum lift or to a more rapid increase in the cam radius. The actual optimization goals for the non-linearity can therefore not be achieved.

1202 shows the situation in which the stroke 0323 is to be maintained with significantly reduced minimum and maximum radii (both thick dashed lines). The cam track is not simply proportionally reduced, since the lift 0323 should not be proportionally smaller, but a new, bold, pointed cam track 0322 results, which again gets a pointed point in the transition from steep to flat, which is significantly more pointed is as the rounder point of the original, thinly drawn, cam track around 0321.

Figure 1301 shows that the "too sharp" points (0322 in Figure 1201 or Figure 1202) that have arisen above cannot simply be rounded off with a cam radius 0324, as is actually obvious. Originally, a flat cam track 0325, ie a higher power transmission, was chosen, for example because the brake can only be actuated in area 0325 (or actuated as planned) if a smaller actuator torque is required as a result. However, if the slope of the fillet radius 0324 is greater at 0325, the actuator may not be able to actuate the EMB in this area, or not actuate it correctly. A possibility is therefore shown in FIG. For example, you could move the fillet radius to 03241 (so that the flatter point 0325 works properly) and then you would have a "wrong" cam track course along the circular track, but not the desired, dashed course.

Now, however, the cam track in the area of the new rounding 03241 in FIG. 1302 is less steep than necessary for the dashed actuation and can therefore be actuated, but more slowly. At the end of the shifted fillet radius 03241, the permitted (dashed) gradient 032221 can be used again. Now the input torque of the non-linearity can be in the desired range again, but the necessary torsion angle has increased somewhat. For this, too, it is proposed that the total torsion angle can be reduced again in a further iteration. Using this method, one can approach the desired course of the non-linearity, but in some cases the limitations will be considered more important than the achievement of the target course of the non-linearity.

In Fig. 1303 you can see that two different sized rollers (roller 033 (large) and small roller 0331 (for the beginning of the actuation) can be used. Due to the small radius of the small roller 0331, a steep flank of the cam track pointed 0322 If the actuating cam 032 has been turned so far that the small roller has traveled the path along the fully drawn raceway, the large roller 033 rolls behind the flank on the dashed raceway and the smaller roller has one from here The course of the career, which provides relief for the smaller role, which is shown here by the continuation of the fully drawn career to the left of F from the assumption of the career for the larger role.

The two tracks and rollers can also be spatially staggered. Furthermore, the raceways do not have to be rigidly connected, but the small raceway can be twisted first and then the large raceway, for example via a driver, which could also achieve a total twisting angle of more than 360°, for example. So it is not absolutely necessary to use different rollers or roller diameters, but this spatial arrangement can also serve to bring about a larger total angle of rotation and this can also be done, for example Drivers or translations move the individual raceways or the individual cams or cam parts from certain states or angles of rotation, so the raceways can also be driven, for example, with different gear ratios. A, for example, three-dimensional screw-like track with, for example, only one roller is also possible. However, the raceways can also be movable relative to one another in other ways (e.g. also spring-loaded) so that, for example, compensating movements are made possible and, for example, a raceway can be changed in position (or its position can be changed under actuation or loading) so that, for example, the pitch at the current cam position is changed.

FIG. 14 now shows a practical example of a cam surface that can be implemented according to the method and the resulting shaft torque on the brake actuator (y-axis) over the additional gas stroke (x-axis).

On the left is a practically possible pointed cam track 0322 shown in bold, whereby the course at the inner beginning is already problematic. The small lower circle with connecting line to the roller 033 to the actuating cam is the fulcrum of a lever on which the roller is located. The resulting brake actuator torque (i.e. engine torque) is shown in bold on the right and the deviation from a constant curve over the linear brake pad actuation stroke is obvious. Full braking would correspond to 1g braking 017, normal braking of normal drivers reach approx. g/3 at 017/3. In the area of the air gap 068, it is not possible to achieve a higher or even constant engine torque, despite the steep start of the cam. The fact that the difference between "g" and "g/3" is so small is due to the force-displacement characteristic of this EMB and this brake pad (both have a realistic background). By applying the improvement measures mentioned above, such as more cam rotation angle, larger initial radius or the smallest possible roll, one would possibly arrive at the thinly drawn course of the engine torque, which is more in the favorable range, but still does not keep the torque constant, especially in the usual braking range up to g/3. The largely optimal (constant) engine torque could only be achieved with other measures, such as sliding scanning instead of roller, very large cam radii, etc., which are not proposed in the present method. In Fig. 15 a possible brake for a front wheel of a car or similar is shown, which for example achieves a max. 40 kN lining pressure force (on the left y-axis) and is operated with an air gap of 0.4 mm 068 (total air gap) and a non-linearity with a limited Change in the transmission ratio has, as is possible with the method presented here. With a contact pressure movement of approx. 1.8 mm (on the x-axis), the resulting contact force (lower full curve in relation to the left y-axis) increases in accordance with the stiffness characteristic against full braking, whereby such stiffness characteristics are usually not straight lines as with springs, but begin softly and become hard against full braking.

The dashed horizontal line would be a design based on constant actuator torque (on the right y-axis) and would theoretically lead to a theoretically optimally short actuation time in the maximum actuator shaft power. However, the design proposed here is based on the fact that one does not want to change the transmission ratio too much or too abruptly, and thus comes to a comparatively very unfavorable course of the actuator shaft torque (upper curve using the right y-axis) in favor of the advantage of mechanically advantageous ones Interpretations (see above). In general, it is suggested to see the design as a connection between the transmission ratio (e.g. output torque to input torque) and the selected mechanical and geometric realization, i.e. the mechanical and geometric realization result in the transmission ratio. Or vice versa, the transmission ratio is selected at each actuation point in such a way that a desirable mechanical and geometric realization is found, e.g. roller diameter, (minimum, maximum) cam radius, minimum radius of curvature of the cam surface. The process can also be iterative, e.g. starting with a target gear ratio across the actuation, then adding the mechanical and geometric constraints, determining the gear ratio from that and then making e.g. mechanical or geometric changes to better achieve the target gear ratio.

From this definition of the design process, it follows that neither the engine torque nor the engine power is taken into account and that these should be largely constant. This definition can therefore be applied to the designs of all types of EMBs, e.g. also to a spring-loaded EMB, which, for example for safety reasons, automatically goes into the braked state and is released via the brake actuator, similar to a railway compressed air brake with a spring.

For this purpose, a first non-linear combination for the spring is proposed, with the relaxation of the spring being translated non-linearly in such a way that the increasing lining contact pressure force is achieved despite the decreasing spring force. To do this, you can, for example, combine non-linearities and let the spring act on a crank pin of the cam, for example, whereby you can give the most tensioned spring a small moment-generating angle, which can then lead to an increasing normal distance when relaxed.

The cam transmission ratio via the actuation can now be designed in such a way that this spring moment is translated via the actuation into slightly more than the contact pressure force required via the actuation. So there is no engine at all here. A further requirement for the transmission ratio can be, for example, that it reaches the actuating force with changes in stiffness (e.g. from full to worn linings) and with changes in the air gap to be taken into account.

When the brake actuator now rotates the cam to release, it must actuate the remaining torque between the spring torque and the counter torque from the brake. This brake actuator torque could now be demanded as being optimal for the motor, or also in places as as low as possible in order to keep the brake released with the lowest possible brake actuator torque at which safe actuation will just take place. In addition, one could also make the requirement that the brake actuator can also get the brake from the braked state to the released state if no brake disc or brake drum is exerting a counterforce, e.g. during assembly.

From all of these requirements, as described, you will get a desirable progression of the transmission ratio and then (also iteratively) check or define the mechanical and geometric properties and, if necessary, arrive at a transmission ratio that does not correspond to the desired one, but does meet the desire for feasibility Fulfills. An additional non-linearity can also be built in, via which the brake actuator has a releasing effect. It must It is also not a question of cams, but it can be any kind of non-linearity in which one is looking for a way between requirement and realization.

The resulting sub-optimal activation time or motor size can (but of course does not have to) be compensated for by higher rigidity of the brake (because energy is force times distance). In practice, an increase in the air gap to 0.4 mm also seems desirable in order to support real lifting of the pads.

16 shows a method that is generally known today for computer optimization, which cuts away parts (dashed straight line in the direction of the arrow) from a solid (here, for example, the large circle) in order to achieve the required cam lift 0323 and checks the remainder to see whether it is a results in a better or worse overall result and accepts the cut or not. A result comparable to that of the methods from FIGS. 1201-1202 and 1301-1303 can be achieved with a different procedure. Therefore, such procedures that find a comparable solution by "trying" are also recommended here, which of course includes in extreme cases that people (also with e.g. scissors and cardboard) use such procedures that can definitely be described as "trying" of whatever kind . Of course, instead of "cutting away", "add" could also be used, e.g. starting from the dashed circle, or "change" in general.

Fig. 17 shows curves for the EMB from Fig. 15 in an e.g. simplest and most cost-effective embodiment, in which no wear adjustment is made and the air gap therefore increases with increasing wear of the brake pads. The upper set of curves represents the resulting actuator torque (right y-axis), the lower set the generated normal force (left y-axis). The x-axis shows the pad lift. The full curve with an air gap of 068 (0.4 mm) is one that will occur, for example, after a long period of operation. The long-dashed one, which will occur, for example, at the end of the planned service life, although braking can of course still be done with a reduced full braking effect. For example, the one with the short dashed line can be the new condition.

The curves shift on the non-linearity according to the enlargement of the air gap and can no longer use it as with a constant air gap. The "new" curve will therefore have a contact pressure force that is too early on a non-linearity part that is still too steep and will deliver an excessive actuator torque that is still in must lie within an operable range. Since this state requires less actuation time, the brake actuator can be run more slowly in order to reduce the electrical input power again. Conversely, the worn state uses a flatter part of the non-linearity, but requires less moment and power but also additional time for further movement. Accordingly, the motor could be made faster here by field weakening or other measures such as voltage increase or winding switching.

In this version, there is no longer an optimum, but only different cases for which the non-linearity can be interpreted together in a favorable way, e.g. so that the maximum increase in power is possible without any problems when the engine is new, taking into account any slower engine running. You don't have to divide the 3 areas into three and you can, for example, design a favorable "new" condition and accept all others with, for example, field weakening or longer actuation times, especially if the wear is normally low and you are dealing with a longer time towards the end of the lifespan can cope. A manual wear adjustment can also be provided so that a "new" condition or a better condition can be brought about again. This setting range can, for example, only be possible to a limited extent or in leaps and bounds or in just one jump, so that the user cannot cause an "insufficient air gap" state. If the setting is incorrect, the control electronics can, of course, issue countermeasures or warnings.

Fig.18 is similar to Fig.17 (same axes) and shows the method of "evasive movement" via e.g. movable holder of the non-linearity shown, e.g. pivoting or movable holder of the motor-cam assembly.

In Figure 18 the non-linearity of the worn state (long dashed line) would be interpreted "favourably", showing that again it amounts to a compromise. If the new state (short-dashed, thin) is present, this non-linearity would not be operable in a permitted manner because the brake actuator torque would be too high. The brake actuator with cam can now deviate (e.g. rotate around a mounting) causing the cam to rotate into a flatter part, causing the engine torque to drop again as shown with the thinner arrow. The turning away can continue and thus already cause a strong lowering (eg see thicker arrow).

This pushing away against a spring is of course primarily “lost energy” because it goes into the spring instead of into the motor actuation. This effect can be limited if the spring presses against a stop, for example, and is only deformed when the stop spring effect is exceeded. The "lost energy" can also return if, for example, the spring can relax again with further actuation.

The control electronics can either recognize the curves that result from the movement (e.g. in the momentary angle behavior) and adjust to the state of wear. However, the displacement or rotation of the holder can also be recorded, e.g. also only point by point when (e.g. in the actuator angle) e.g. a stop is left. Since this change takes place slowly due to different wear and tear, the electronics can also record this statistically by averaging or otherwise and smoothing it out, for example.

Such twisting or movement influences on the component holder can also occur in other ways, e.g. via wheel deflection. The actuator assembly does not have to be twisted or shifted either, the roll-off lever or another part can also be influenced accordingly.

In Figures 1901-1905, for example, it is proposed how a known and frequently used expansion part 051, with an unbraked position 053, a braked position 054 and an expansion part pivot point 057, can be advantageously modified (schematic) for pressing the two brake shoes 067 with brake lining 063, for example can (whereby, of course, pressure could also be applied in a similar way in other brake designs, such as disc brakes or brakes on linear rails, and also, for example, only a single pressing movement is used). Its stroke can be so small that only the contact pressure is possible and therefore an additional wear adjuster is necessary, which, for example, pushes the other end of the brake shoes apart. Or the stroke can be so large that the wear can also be covered with the spreading mechanism.

One sees in Fig. 1901 a common, inexpensive expansion possibility, whereby a part is twisted between the brake shoes similar to a flat-head screwdriver. Included scratch the edges due to the change in height of the contact point as a function of the angle function, abrasion and relatively high mechanical losses occur, which not only increase the actuating energy, but also cause unpleasant hysteresis, so that the brake requires significantly less force when releasing than for actuating it. However, this solution is not excluded here. It is physically less advantageous, but may be cheaper and can also be replaced by improved variants, such as with rounded edges or with compensating parts that come into contact with the expansion part or with those with favorable behavior in terms of "scratching".

Here, as a modification of a conventional expansion part, no complete elimination of the change in height is proposed, but only a good reduction or it may be assumed that a change in height can even be desirable in order to follow other movements, e.g. a brake shoe movement or e.g. one caused by deformation caused movement. If the loss-generating relative movement between the expansion part and the brake shoes is reduced with simple means (as suggested above as an example), e.g. to less than 2/3 of the unfavorable situation, you are already on the right track. These mechanical losses of the known expanding part may have been accepted, e.g. with a hand-operated drum brake, because e.g. the manual force was sufficient for the brake actuation with a suitable transmission ratio and there was therefore obviously no need for improvement. With an EMB, however, the mechanical losses have to be overcome by the actuator and so it does play a role for the actuator size (installation space, weight, costs, etc.) whether it has to deliver e.g. 50% to 100% more power. In addition, the mechanical losses worsen the relationship between actuator torque and contact pressure. For these and other reasons, this improvement with the mentioned expansion part is recommended for the EMB. It should be mentioned here that there is another known variant with a so-called S-Cam for drum brakes, but in which a roller runs on the S-Cam on each brake shoe and a different approach is taken.

Two overlapping movements are used for this purpose, a rolling process on a circular rolling track, for example, and the change in height as a function of the angle function. Shown is an example of how to form an inexpensive and easily manufacturable roll-off track can generate from pens. For this purpose, for example, in Fig. 1902, two holes can be drilled into a round part that is still full, and then the disturbing material can be removed, for example by milling away. Of course, these steps could be performed in other ways, such as by stamping, pressing, forging, casting, sintering, or cutting. Now, in Fig. 1903, the pins or needles or rollers etc. can be inserted into the remaining material. These rolling surfaces can also be created differently than with pins, i.e. arbitrarily, for example by milling, sintering, and do not have to be exclusively circular or part-circular. The needles (or cylinders, pins or rollers, etc.) do not have to be pressed in either, they could also be formed, for example, by pressing or forging the entire part, but they are nevertheless described below under the designation "pins" or something similar. The position and the diameter of the rolling pin is now selected in such a way that the scratching relative movement between the brake shoe and the rolling surface is small (Fig. 1904), whereby one can never reach exactly zero because the rolling movement is proportional to the rolling angle and the change in height from an angular function originates. Advantageously, one will also include the rotary movement of the brake shoe around its (here lower) bearing point, so that one would like to follow this rotary movement as a goal.

FIG. 1905 shows that the two brake shoes 067 require fundamentally different movement sequences for the pins. Suppose the brake shoes make a circular movement (indicated by the arrow from 069 upwards, interrupted to indicate that the vertical distances are pushed together) through the lower bearing of the brake shoe support 069, which runs minimally downwards with a pressing movement at the point of contact on a pin. Now the upper pin would have a combination of its circular path around the center of rotation and its rolling on the pin circumference. This combination should make small relative errors against the shoe contact movement at the contact point of the pin, which can be supported by the pin radius and pin spacing as well as the beginning and end of the pin twisting movement. A symmetrical lower pin would naturally unfold this component exactly opposite to the upper pin due to its circular track motion. Therefore, an arrangement that is point-symmetrical to the upper pin (with regard to the pivot point of the pins) could also result in acceptable behavior under certain circumstances, but a more favorable solution would be to improve the area of the circular path of the lower pin so that more Downward motion by circular track at lower pin comes up, as shown at far right by pins not being exactly point-symmetrical to fulcrum. Of course, one could also or additionally make the two pin diameters different, for example, or also use non-circular pins, for example. An unbraked initial position is shown in dashed lines, in which the brake shoes leave an air gap to the drum. The bold circles show the pins at a maximum possible angle of rotation, which could correspond, for example, to full braking with the linings worn to the maximum, if one also wants to cover the lining wear from this rotary movement. The position of the lower pivot point of the dashed lower halves of the covering has of course been shortened and is not true to scale. The position of the bottom pin shown would perform a little less movement in the horizontal direction than the top pin, since the horizontal component of the trigonometric function acts at a slightly different angle than above. This can simply be neglected because of the lining wear that sets in, or it can be compensated for by giving the lower pin more center distance. The radius of the two brake shoes (from the pin to the pivot point) can also be slightly different, so that these differences in the position (angle, center distance) of the two pins can or should be taken into account.

Optimal designs could therefore be found for the pin movements and shoe movements, with the entire movement sequence on the pin following the shoe movement well in the sense of a local optimum, and this is also one of the design goals.

However, one does not have to focus on the remaining relative movement error, because in the area of overcoming the air gap, the contact pressure is small and so are the losses from the relative movement. The movement component for lining contact pressure for normal braking can also be small, so that small remaining relative errors do not have to be in the foreground. If a larger rotation is necessary to cover the pad wear, the pressing part of the movement will adjust itself absolutely in such a way that the height errors can be compensated.

The focus can therefore also be on converting a comparatively bad situation in terms of "scratching" and losses into a significantly better one and at the same time on good manufacturability and favorable mechanical stress (e.g. the small rolling radii and the remaining cross-section of the center piece) and to coordinate the optimization of the pins and their positions with these necessities. The main goal is to go clearly from the unfavorable condition of the screwdriver-like part to a reduction in the undesired relative movement, while striving for mechanically and geometrically sensible solutions and not just striving for an approximation to a mathematical optimum. One could also get by with just one pressure pin, for example in duplex brakes and two such spreading mechanisms, or one could arrange the central bearing of the spreading mechanism in one pad shoe and spread the second away from this shoe with just one pin. The expanding parts described will have a slight non-linearity, which can be used, for example, to compensate for different brake rigidities with different lining wear. However, it is also possible elsewhere, for example in the case of a non-linear drive for these expansion parts, to take account of their slight non-linearity. This rotatable expansion part does not necessarily have to be driven from the center of rotation, but can be rotated in any way, for example by attaching a lever to it or, for example, teeth for a gear drive. The center of rotation does not have to be stored and it does not have to be used either, for example, there could be a lever on the rotating expansion part and the center of rotation could, for example, neither be stored nor used, but simply result from the rolling movements.

In Figure 20 it is suggested that a depression 0311 can be in an area 082 not required for braking, which here, for example, pulls a lever with the lower driver 025 upwards and thus brings progress on a ratchet-like (here star-shaped) toothing 026 and so that the brake actuation shaft rotates in the direction of more braking. Such a progression of the ratchet could, for example, also be obtained intermittently from a black, rectangular lower driver 025, whereby this stop can rotate the ratchet in the direction of wear adjustment. Of course, you could also use other positions for wear adjustment, eg 025 at the top with more cam rotation than for full braking, but the actuation of the wear adjuster is usually more difficult here. But you could also, for example, during normal brake actuation in the area used for braking 081 then carry out a force- or torque-limited ratcheting progress if this does not occur from a certain expected contact pressure and use the limitation to ensure that too much is never readjusted. In general, any device that behaves in a direction-dependent or controllable manner can be used as a "ratchet" in all versions here, regardless of whether it is gearing, frictional locking, wrap springs, clutches, etc Give leg of a wrap spring. The arrows indicate different actuations for turning the wear adjustment further. The inner, adjusted gearing would, for example, turn an expansion part 051 or an S-Cam 056 to press the lining (schematically indicated).

Fig.21 shows suggestions as to how three functions can be derived from the movement of the brake actuator (e.g. normal service brake actuation, wear adjustment and a parking brake position, which can also be held permanently), whereby in turn, for example, several brakes are addressed together or, for example, only one and, for example, brake disks 011 or brake drums 012 can be used, of course the same ones are best (in contrast to FIG. 21), of course the principle can also be used with only one brake instead of two, of course only one of the levers with the toothing 026 is required and the others only show possibilities. It is also indicated that a joint pressing rotary movement via the teeth 026 also includes the wear adjustment or, for example, a separate adjustment movement is carried out on the wear adjustment actuation 08, here, as is known, e.g. between the brake shoes. The actuating cam 032 with cam axis of rotation 034 has a constant increase in height for the service brake in one direction of rotation and in the other direction of rotation, for example, a depression or a path with a constant radius to the cam pivot point as a parking brake position 0471 and thus pivoting the lever (in the illustration upwards) with the roller, the force is transmitted to the thinly drawn brake actuation shaft, for example to the right via the expanding part drive 052 to the expanding part 051 .

On the disc brake 011 in the left area, an adjustment lever 027 is connected directly to the brake actuation shaft. This lever can be pushed up in a special position, which is an advance on the "ratchet" and with it Causes wear adjustment. Here, for example, a parking brake sink is operated with so much contact pressure that enough parking brake effect is created, but the pads could be pressed even further. A special part, here for example a pin or driver 025, assumes the lifting of the adjustment lever if the cam is twisted more than is necessary for the parking brake position. If this lever now rotates the ratchet forward (more contact pressure), the brake is adjusted correctly (or better) again and, from the brake actuator's point of view, is again operated at correct (or better) operating angles, as corresponds to a correctly adjusted brake. However, since in this case the adjustment lever is directly connected to the brake actuation shaft, the adjustment lever must be pressed further and further for renewed ratchet progress, since it is connected to the pad pressure and this moves closer and closer to the brake disc with wear. Nevertheless, more cam twisting force should not necessarily be required, because more wear means less counterforce from the brake actuation and, on the other hand, the "pin" or driver 025 and the adjustment lever can be arranged and shaped in such a way that the desired behavior results. They can also have a cam-like effect and be designed, and the "pin" can be anything, eg a roller. How and at which cam position exactly the (here exemplary) upward movement is triggered can ultimately be solved in many ways. This method is therefore particularly suitable when little wear is expected, as could be the case with bicycle brakes or bicycle trailer brakes, for example, or with parking brakes, where there is actually little or no wear at all if they only hold the standstill. Of course, the above is not tied to brake discs, the friction surface could also be a drum or rail or something else.

The adjustment lever on the shaft therefore has the property that it has to be turned more and more with increasing wear adjustment. This effect was treated here with a “double ratchet” on the drum brake in the right area of the figure. The function is almost the same as described, only the adjustment lever 027 on the right can also make progress on its "ratchet" after the adjustment process in order to remain in the area of the old position, i.e. to essentially always carry out the adjustments in a similar or the same swivel range. Common parts can also be used for these two "ratchets", such as the teeth on the shaft or friction partners or the wrap spring or its legs, so that, for example, a ratchet action is actuated with one wrap spring leg and the second ratchet action with the other wrap spring leg. There may also be intentional friction between all the parts described here, for example to prevent the "ratchets" from twisting unintentionally, for example due to vibrations. Of course, friction surfaces other than drums can also be used, such as discs, rails or others.

In the lower area of the right-hand drum brake, it is shown that an adjustment movement of its own could also be applied to a wear adjustment actuation 08, e.g. as a rotary movement or, as indicated by the double arrows, as a pushing movement which, for example, can turn the adjustment screw like a ratchet.

In general, it is suggested that if there is a need for readjustment that is recognized when there is a large contact pressure, for example if there is a lot of contact movement (this can come from the contact actuation of an actuating spring or from parking brake positions or positions with more actuation movement than to the parking brake position), readjustment should usually be used can be stiff or impossible if the contact pressure or parts of it weigh on the adjuster. For this reason, the following solutions are proposed: either the readjustment actuation becomes so strong that the readjustment movement becomes possible, or the need for readjustment is “saved” and carried out when the readjustment is normal again. For example, you can intentionally move or turn a ratchet in the non-adjusting direction in such a way that the ratchet is moved, for example, a tooth against the adjustment direction and that is also possible because the adjustment shaft or the adjuster are stiff due to the pressure, but the Ratchet arm can make a tooth against the adjustment direction of rotation, for example. This movement would be made, for example, against a spring, at least against a part that can "store" this intention. When the brake is released again, this spring can then turn the adjuster shaft or the adjuster in the direction of adjustment when rebounding. Such "ratchet-like functions" can be, for example, ratchets, wrap springs, friction devices, etc. and be combined locally or, for example, on the drive of the actual adjustment device in the brake, e.g. is rotated against the screw axis, which can also be proposed, for example, with "ratchet-acting" trained wheels such as bevel gears. One possibility is shown in FIG. 22 as an example with its own wear adjustment, whereby the principle can of course also be used if the lining pressure itself also takes over the wear adjustment at the same time due to the possible stroke and again preferably two of the same brakes are used or just one.

A "possible force or moment limitation" such as with a possible slipping clutch 023 can, for example, advantageously ensure that excessively incorrect, excessive readjustment is impossible because the limitation is not able to do so. For example, a possible spring action from a spring for wear adjustment 021 cannot allow incorrect adjustment, which has a less favorable effect than slightly rubbing linings. A "possible path limitation" can also be used, e.g. instead of a slipping clutch 023, so that the adjustment process does not adjust the path or angle of the desired air gap, but already adjusts it if more path or angle arises in the lining lifting movement. The path of the air gap can, for example, be evaluated from "rotatable wear adjustment". The wear adjuster shaft or rod to the brakes or the adjuster in at least one brake can intentionally have so much friction in the wear adjuster 028 or be provided with additional friction that unintentional further turning of the wear adjuster (e.g. due to vibration) is not possible. An additional ratchet with teeth 026 for the

Readjustment movement can also be recommended for specifying the readjustment direction. A combination ratchet with friction can be recommended, for example in the form of a wrap spring. Without these additional components, the adjustment can also be carried out functionally if necessary. In the case of a drum brake with a brake drum 012, a wear adjuster 02 with an expanding effect can also be used, for example. Of course, both brakes should be the same or only one can be used. In all of the above embodiments, one can advantageously determine during the wear adjustment whether the behavior corresponds to what is expected and actions can be derived from this, such as, for example, adjust more, less or not, warn or store deviations. The adjustments can also be designed in such a way that incorrect (eg too great) adjustment runs as unfavorably as possible, ie is not possible due to the required actuator torque, for example. In FIG. 21, the parking brake position was in the cam area opposite the service brake, so active service braking would have to be released before parking braking 16062 can be carried out. Therefore, for example, only a single brake that is currently active is brought from a service brake position into a parking brake position if desired and, if possible, not all brakes at the same time. However, the parking brake positions can also take place differently, for example at the end of the service braking or by locking a brake position by means of a holding device. However, a separate actuator can also be used for the parking brake, which could also take on other functions, such as wear adjustment or emergency braking.

In Fig. 2301 -2302 advantageous exemplary embodiments for wear adjustment and braking force detection are shown using an internal shoe drum brake, with other versions being possible, of course, such as disc brakes or brakes for linear movements, e.g. on rods or rails.

2301 shows that the expanding part 051 (top) can also be movably mounted and should be in a starting position, e.g. by spring action or e.g. by a driving force measurement 064. If the driving force measurement 064 is now rotated by the braking force, a position sensor, a force sensor or a switching or whatever recognizing function can detect the braking force or at least trigger a switching function at least at one point of the braking force, for example, with the arrows in Fig. 2301 064 different possibilities are indicated. This can, for example, support a "hillholder function" in which it is recognized, for example, that a vehicle is being braked but wants to roll back and, when the driving force is applied, leaves the rolling force component, does not roll back and then possibly even pulls the brake slightly forward would. From this, a favorable point for releasing the brake in favor of beginning forward travel could be determined. Possible brake shoe supports 069 can shape the freedom of movement or force dissipation.

This driving force measurement 064 can also serve to increase the accuracy, for example by recognizing the point at which the pads are slightly driven along or even measuring the braking force and also controlling it, for example. A possible lower Brake shoe support 069 could also be mounted jointly on or with this movable catch or also freely movable relative to it. The possible brake shoe supports 069 can be used, for example, to restrict the range of movement of the entrainment, with which, for example, unpleasant noise development can also be prevented, such as squeaking or shaking. For this purpose, the stops can also be soft or rubber-like, for example. The possible lower brake shoe support 069 can also be used for a servo function, in which the bearing of the expanding part is taken along by the braking force on the "primary shoe" and thus (here below) the "primary shoe" on the "secondary shoe" exerts another pressing force. The "secondary shoe" will rotate and then be stopped by brake shoe support 069 or drag force measurement 064. This can (but does not have to) be carried out symmetrically in both directions of rotation by means of two stops, but it could also, for example, only be pressed on one shoe with the expanding part 051 or an asymmetry of the braking effect depending on the direction of travel could be brought about. Mainly with these servo drum brakes, there can be another non-linearity in the actuation of the expansion part, e.g. spring, so that when the expansion part is stationary without self-reinforcement, the resulting higher drive force of the expansion part initially results in spring deformation, for example, and then the actual expected rotation of the expansion part takes place with self-reinforcing rotary movement can. "Top" and "bottom" only have an explanatory effect and can be used in any other way.

In the lower area of Fig. 2301 it is also shown that the lower brake shoe support 069 can also be designed, for example, with a wear adjustment device 02 (e.g. adjustment screw) or with a cam or double cam (indicated by thick cam tracks at 02) and the force can be derived or applied, for example can pass on the other shoe. The support points of a cam track can still be selected in such a way that the shoe is geometrically positioned in a favorable manner according to the wear of the lining. A cam track can preferably run relatively flat in order to keep the force acting on this cam drive from the pad contact pressure low via the friction. Such an adjustment cam or adjustment screw can be driven, for example, from actuator areas that are otherwise not used for braking, or from the brake actuation. Advantageously, for example, "behind" a parking brake position, an adjustment, for example via a Spring actuation "marked" and then readjusted by this spring action after releasing the brake. Or, for example, after a certain brake actuation, an attempt can be made to actuate the adjustment via torque or force limitation, but this would be impossible if the adjustment was set correctly, since the pad contact pressure force or even driving force would require more adjustment torque than is possible via the limitation. A wrap spring ratchet, for example, can advantageously hold the position of an adjustment cam (or double cam), for example, since it also generates static friction in the holding state and only allows movement of the cam, for example, in the adjustment direction, and a second ratchet effect can serve to turn the adjustment cam, for example, in the adjustment direction and the Adjustment rotation can be torque-limited via a slipping clutch, for example. An adjustment cam does not have to have a cam track on both sides for both shoes, but could also be mounted in one shoe so that it could rotate on one side. The parts of FIGS. 2301 -2302 can be mounted in many ways, for example on rotatable plates, which also contain a driving force measurement, for example. Driving force control is also proposed by driving force measurement.

Fig. 2302 below shows a possibility of a close, e.g. concentric, drive of adjustment and contact pressure, whereby on the one hand a wear adjustment 02 (e.g. wear adjustment cam or screw) is driven and secondly a contact pressure actuation, shown here in dashed lines, here e.g. an actuation cam 032, which is a lever drives the spreading mechanism. You can also see that the expansion part 051 of the pad pressure does not necessarily need a guided center point, but can also be held in place in another way, e.g. via the shown upper and lower guide of the pins, which form the expansion part 051 here. The guide is only necessary mainly in the air gap or with small forces, since with greater contact pressure the friction of the pins on the rolling surfaces takes over the guide, so the black guide between the pins would be more advantageous here.

If the center point of the expanding part is not guided, the advantageous rolling designs of the pins described earlier become simpler, since each pin does not have to be designed favorably with regard to its guided pivot point and the movement of the brake shoe, but without a guided center point only the movement of both pins with regard to both shoes is important and the fulcrum can move freely. In such cases without a guided center point, one would, for example, press the pins into a "flat iron-like" lever part or press them in between two "flat iron-like" lever parts or fasten them in some other way, eg by soldering, welding, gluing or riveting.

In Fig. 24 a proposal for a brake actuation with a spring effect is shown, whereby at 0571 one can e.g.

For example, a service brake function is supported by the upper actuating spring 042 in such a way that the service brake function can release itself, i.e. the spring support is less than the effort required for pressing, in which the upper actuating cam 032 runs relatively steeply. This allows u.a.

Actuator actuation energy can be saved. This interpretation could, for example, not be solved completely on its own, but sufficiently extensively.

The cam side of the parking brake function (lower actuating cam 032) runs flatter so that the spring can always actuate, whereby "steep" and "flat" of course always refer to the resulting non-linearity from a mathematical point of view and the forces or moments of course always correlate correctly and consistently are to be obtained. You can now, for example, design the flatter parking brake side in such a way that the parking brake side is not spring-actuated to the end if the wear adjustment is set correctly. If there was too much wear, the parking brake side would then continue to turn and trigger or mark a wear adjustment. This adjustment position could also be approached actively with the brake actuator if actuator-controlled parking braking were necessary. The parking brake positions would remain spring-loaded in the absence of power and when the power is switched on, the brake actuator can take over the required braking and functions again. In this embodiment, the spring could, for example, act in a crank-like manner on the cam. However, this means that the non-linearity of the spring is linked to the behavior of the crank.

Of course, as shown in dashed lines, the spring could also have any other non-linearity, such as with its own cam (the dashed actuating cam 032 or double cam), which of course is much more creative freedom. Under certain circumstances, both cams could even act on the same roller.

For example, you could also change the spring preload (indicated e.g. with the arrow at the upper actuating spring 042) in order to switch the EMB from a parking brake behavior to a self-releasing service brake behavior and thus use only one cam.

Since it does not matter in principle how the actuator motor and spring interact in a precise mechanical design, it is only important here that they can interact via linear and non-linear gears, wherever and however the parts are arranged.

An advantageous lever is proposed in FIG. 25 (as is also realistically conceivable in these proportions), which executes rolling movements between the rotated contact surfaces 0591 and non-rotated contact surfaces 0592 and on the long lever arm with a non-linearity 03, e.g. an actuating cam 032 on a roller 033 is pressed. For the turned contact surfaces 0591, roll-off cylinders are favorable in terms of production technology, can be hardened and are very round, which will be important later on. Basically, however, it is about this unrolling of round parts, whereby basically every manufacturing option is open and they are therefore referred to in the following generally as "unrolling cylinders" (although other, non-circular or non-cylindrical geometries are also permissible here), both together in general hereinafter referred to as expansion part 051.

The expansion part 051 from FIG. 25 would generate a 0.6 mm y-movement with a 1 mm contact pressure stroke per roll-off cylinder due to the angle function of the rotary movement, but a y-movement of approx. 0.7 mm due to the roll-off on the circumference of a roll-off cylinder and thus with 2 times 1 mm in the event of full braking stroke cause a total y error of 0.2 mm, since the errors on both unwind cylinders add up.

Fig.26 shows the "disturbing, lateral, or scratching" y-movement (y-axis) over the contact pressure stroke (x-axis), with the solid line above showing the y-movement through the angle function and the dashed line above the y movement through the rolling circumference, ideally both should be the same, but here one error remains, the represented by the top arrow. The bottom curves are the same except they cause the y errors to go in the opposite direction. Decreasing the rolling circumference results in proportionally less y-motion and the y-error can result in smaller y-motion with smaller rolling circumference (dashed-dotted line) as shown below, which can also result in a different sign of the error. This also suggests that the total y-error can be reduced by combining different unwinding cylinder diameters, but can never be completely eliminated because an angle function and an angle-proportional unwinding circumference are never exactly the same.

Here it is proposed to dispense with the most perfect possible rolling along a straight line and to find another, more realistic approach that is also based on good manufacturability and commercially available, preferably very good circular rolling cylinders and thus also with regard to the most “linearly guided” movement possible contradictory, seemingly suboptimal solutions are allowed.

FIG. 27 shows that no "guide" is required for the lining pressure for a disc brake, as is the case in a hydraulic pressure cylinder. On the contrary, the upper and lower swung dashed lines are supposed to show that (as suggested here) there will be a freedom of movement, which doesn’t have to be sharply defined, but e.g. can also behave elastically. A contact pressure 05 (e.g. a wear adjuster) is pressed in any contact movement 059 by an expansion drive 052 and deforms the brake, e.g. from the unbraked position 053 to the e.g. braked position 054

This creates different areas during the pressing process, starting with the area of the air gap and low contact pressure: Here, the rightmost contact pressure 05 (here a part that contributes to the lining pressure, e.g. also a wear adjuster) will have any initial position that, due to the weight, e.g lower position, but can also have a different y position, for example due to vibrations. Due to the low contact pressure between the turned contact surface 0591 and the non-turned contact surface 0592, there is hardly any wear and loss of actuation energy. With further actuation (increasingly left 05), the "error" from the rolling circumference and angle function according to the above illustration, e.g.

(intentionally designed or given by geometry) such that at higher friction (between rotated contact surface 0591 and non-rotated contact surface 0592) the presser moves down. In this area, therefore, a transition occurs between the above first area and this one. In reality, in this example, a (also random) position of the area of the rightmost 05 can change in this area via realignment movements (e.g. the pusher moving downwards, relative frictional movement between rotated contact surface 0591 and non-rotated contact surface 0592). An area is indicated here for heavy braking, where significant deformations (eg bending) can occur. In this area, a higher "error" between the angular function and the roll-off is accepted or sought in order to even compensate for the change in height due to deformation with a favorable movement. It is entirely up to you how many areas are included and with what behavior. It is essential that lateral compensatory movements and/or frictional compensatory movements are allowed here and even changes in geometry (e.g. through deformation) can be compensated.

28 shows a brake in which the brake actuator consists of three drives, for example, and one part (e.g. motor 041) via a gear mechanism, for example, acts on the non-linearity 03, for example an actuating cam 032, such that the brake is released automatically if the power of this motor 041 fails. In addition, an actuation spring 042 (drive 2) can support the actuation of the brake or the release, or as shown here as a compression spring, with weak braking it can support the release and with heavier braking it can support the actuation. This actuating spring 042 does not have to act directly on the non-linearity 03 but can also be arranged and act in any other way. The third drive for the entire brake actuator is, for example, an electric parking brake drive 047, which can also be implemented in a different way, for example with a cable pull. Here, for example, a worm drive would prevent the parking brake from being released when there was no current. The parking brake spring 048 can be present as an elastic coupling element and, for example, continue to rotate the actuating cam 032 when the brake cools down and requires retensioning. For this purpose, the actuating cam 032 can also consist of two (or more, if additional functions are desired) cams, so that one can also be specially designed for this further rotation. The parking braking area of a cam can also be in a different direction of rotation, for example. Of course, this parking brake drive 047 can also be used as a safety function in the event of failure of the other motor acting on the non-linearity for service braking. But that could also be a cable without a worm drive, for example, which comes into effect in a bicycle brake if the motor acting on the non-linearity fails. The parking brake movement of the non-linearity can take place separately from the other motor, for example via a freewheel, so that the parking brake position of the non-linearity can be reached, for example, when the other motor is not rotating. A different action of the actuating spring 042 is also recommended here if, for example, both a stable "fully released" position and, for example, a stable "well braked" position with, for example, only one non-linearity are to be achieved (which can be useful for a bicycle or bicycle trailer, for example). ): Here you could, for example, attach the actuating spring 042 like a crank (e.g. to the non-linearity), so that the pressing actuating spring 042 is relaxed in the direction of released and also in the direction of being actuated and there is a dead center in between (similar to that indicated in Fig. 28) . If the power supply fails, the vehicle (e.g. bicycle trailer) could continue to drive with the brakes fully released or a parked trailer could remain in the parking brake position. The parking brake position could, for example, be brought into the "released" position manually without electricity. Since with this actuation spring action there is a spring action from the moment the dead center is exceeded, which acts to support the actuation, operation with little actuation power is possible, i.e. the electric motor power for actuation can be less than without a spring and it is even possible to hold an actuation position without electricity if there are mechanical friction losses in the brake actuation and the so-called cogging torque of the electric motor alone can hold the operating position. Of course, locking or braking devices can also be present in order to hold the actuator in a certain position. Even if this brake can be released with manual operation, electronic anti-theft protection is possible: as soon as power is available again (e.g. wheel hub dynamo), unauthorized operation can be brought back into the braked state with the electric motor of the brake.

Another safety design of the actuating spring 042 would be, for example, that it should always bring about a braked state: A bicycle or, for example, a rail vehicle could thus be used, for example, in the event of a complete power failure (or, for example, a shutdown from, for example, For safety reasons) go into a normal, for example, braked state, which can be uncomfortable with regard to driving on, but can be sought as a safety solution.

A calibration spring 046 can be present, for example in order to be able to compare a known or stored spring characteristic (or at least one value) with the engine torque determined (e.g. from the current) in a non-braking state and to be able to control the brake more precisely or the beginning To be able to better recognize the touch of the brake pad to the disc. This calibration spring 046 can act in braking, in an air gap or in an actuator movement that does not cause any significant movement of the lining, or in several such movements, also with different effects and tasks. A spring that fulfills at least one other function can also be used for calibration purposes. How the motor torque is represented here is arbitrary, it can also be shown here as "force", current or without a unit, but this calibration will advantageously also record and take into account the current friction in the drive. A spring for creating an air gap 07 can help in a known manner to push the brake or the pads apart in the unbraked state, i.e. away from the braking effect. The spring for air gap generation 07 can also be related to the engine torque for calibration purposes. The spring behavior can also be included in the determination of the mechanical losses, also in connection with the air gap, contact point and course of the non-linearity. The calibration spring can, for example, be used in an engine area without or with very little pad lift and from pad lift the additionally acting release spring can also be used for calibration purposes. This calibration can also be seen as a determination of a deviation, also as a comparison (also including the course of the non-linearity and characteristic curves of the springs) with something measured, but also as an instruction (what needs to be done to get better or to achieve something), whereby at least one value is worked out here that explains deviations in such a way that they can be compensated.

A drum brake application is shown in FIG. The expanding part 051 shown above presses on the two brake shoes 067, which can rotate around their brake shoe support 069 and have different turning radii (longer and shorter arrow pointing upwards). Such simplex drum brakes can develop self-reinforcement because the “primary shoe” gets a drag component around the pivot and the “secondary” gets a slightly weight-reducing component. To do this, however, the spreading force must allow for a slight displacement (which can also be provided here, indicated by the horizontal arrow) in order to be able to follow the differently pressed shoes. In the case of mechanically actuated drum brakes, however, the expanding parts 051 are often pivoted with little play in order to absorb leverage (e.g. from a cable pull). It is therefore proposed here, in such a case, if necessary, to design the two partial strokes from the two (upper, lower) rotated contact surfaces 0591 by positioning the rotated contact surfaces 0591 (rolling cylinder) differently to the expansion part pivot point 057 in such a way that the resulting contact pressure curve is similar to that of a displacement would be self-reinforcement. Furthermore, the roll-off cylinder diameter and position relative to the pivot point are advantageously designed in such a way that the contact point on the shoe follows the combination of circular movement of the shoe and any deformations or geometric changes with as little relative error as possible. The different lever ratio (e.g. to the imaginary center point of the rubber layer on the drum) due to the different radii (longer and shorter arrow pointing upwards) can also be taken into account in the position of the roll-off cylinder. In contrast to the disc brake from Fig. 28, the parking brake position 0471 can also be selected in the opposite direction of rotation of the non-linearity 03 (with e.g. cam axis of rotation 034) to the service brake (which would of course also be possible with the disc brake from Fig. 28), with the parking brake position 0471 can be held automatically even when there is no current, for example by means of a special geometry or spring action. Of course, the expanding part 051 with the drive of this drum brake could also be designed similarly to the disc brake from FIG. 28; there are many possible combinations. At the end of the parking brake position (or, for example, the service brake position), a wear adjustment can be carried out (also specifically if necessary) or, for example, stored in a spring for wear adjustment 021 and carried out when the brake is released. In the case of the non-linearity 03, there may be a special area in which an initial position of the non-linearity without pad travel 111 can be found, for example in that an increasing engine torque can be identified in each of the two directions of rotation. To find the initial position of the non-linearity 03, a stop or a spring can also be approached, including the calibration spring 046 mentioned, which can have the particular advantage that it can be approached, for example, before the first real braking and can be in an actuator rotation range, for example , which can have special properties such as, for example, no significant pad travel or, for example, in a direction of rotation or range of rotation that is not used for normal brake actuation (which would require a different installation, for example acting on the non-linearity). Thus, for example, before the first braking, a calibration can be carried out, which values measurable on the actuator (eg current, power, energy, etc.) correspond to which spring action and also, for example, via the (possibly also extrapolable) spring characteristic or points from it. The currently prevailing unwanted mechanical loss can also be recognized. A distinction can also be made as to whether only "idling losses" occur, as long as the actuator movement is not yet associated with any significant movement of the lining and no spring is yet acting, and from when the spring effect is recognized. When braking, it can then be concluded very precisely when the pad contact pressure force begins to increase, for which purpose the instantaneous non-linear translation between the actuator measurable value and the pad contact pressure force must of course be taken into account.

The drum brake from Fig. 29 has the advantage that it can lift the linings off the drum, for example by means of release springs. Now, however, a low-backlash rotatably mounted expansion part cannot make any compensating movements to compensate for different thicknesses of coverings with different beginnings of contact. Here it is suggested that either slight elastic mobility can provide compensation and more even touching or, on the other hand, the coverings will wear out so compensatingly through defined movement that they will put on similarly. This can be supported in particular by precisely manufactured or intentionally adjusted coverings and also by choosing the appropriate contact pressure geometry (as suggested above).

30 shows a possible recommended procedure with regard to a calibration spring 046 or a resilient effect that can be used (also to a limited extent) for the same purpose (advantageously, for example, also in a range from little to essentially none Lining stroke, i.e. also in an actuator direction of rotation not used for normal service or other braking, i.e. a range not used for braking 082): From a starting position, e.g still without spring action (which can be seen, for example, as running with covered losses without any other energy supply), tensioning the spring with the calibration spring characteristic 049 from (e.g. essentially) the mass inertia of the rotation, determining the "braking distance" until the spring brings the rotation to a standstill , Acceleration by the spring (now e.g. against the above direction of rotation), whereby this acceleration can e.g. also take place with a defined motor current (e.g. also advantageously zero), Approaching a point from which normal operating or other braking in an area used for braking 082 is started, eg the axis intersection point in Fig.30. This process could be carried out in a short time, e.g. when switching on the brake, and gives a very comprehensive picture even before the first braking and brings the brake into a defined state for the following braking(s): You can see electrical and mechanical losses when accelerating, too until the spring is reached, then when the spring is tensioned. Tensioning, for example without (or with defined, e.g. loss-covering) electrical energy, can make the mechanical losses recognizable; before reversing the direction of rotation, a measurement can show what is necessary (e.g. current, torque, etc.) to keep the spring tension at a standstill, during the subsequent acceleration after reversing the direction of rotation, for example, the mechanical effect of spring force against mass inertia can be seen show how to overcome the mechanical losses. The initial losses 016 in the area 082 not used for braking can be seen as "no-load losses". According to the calibration spring characteristic 049, the losses 016 may be higher. In principle, when the direction of rotation is reversed, they appear twice (double arrows 016 left), because they appear first in one direction and after reversal in the other. The same applies to the losses 016 on the right, which are usually even higher than 016 on the left due to the pad contact pressure. This is not tied to exactly this process.

It is recommended (but not mandatory) to attach the spring at a point in the translation where the spring can be actuated more than in the pad stroke, because then, with a smaller spring force, the above process runs closer to the range of normal braking or a smaller spring can also be used. In the above process, many measurements can be made, but this is not mandatory, you can also, for example, only measure the total energy consumption over the entire process and, since no energy would have been necessary without loss, deduce the state of loss from the energy. How exactly the process runs, whether only parts of the process take place or are used and what is measured when and how can be freely designed, as well as which areas 081 and 082 are used or unused. It is essential that the process can be used for calibration ( eg when switching on, but also otherwise). One can also see, for example, in the area 081 used for braking that the actuator torque increases later, for example due to an excessively large air gap, which leads to the dashed curve 081 . It can also make any measured values recognizable, ie also, for example, the measurable state to be expected on the actuator for a specific pressing force (braking effect). In general terms, above process is conversion of one form of energy into another (eg electrical into mechanical or eg kinetic into potential like spring tension, mechanical into electrical). Of course, the method can be related to this energy conversion in general and is not limited to specific components such as "calibration spring". A physically equivalent process (or partial process) occurs, for example, when the actuated brake (which manifests itself as a spring effect) accelerates the motor when it is released or brakes an actuation movement, for which purpose the acceleration or deceleration can also be allowed to run with zero motor current to essentially recognize the mechanical losses. The clamping force (or the moment resulting from it) acts in the brake (and possibly other forces, for example from springs) as an acceleration or deceleration force. If this is stored (e.g. as a characteristic curve) the actual state of the brake could deviate from the stored one and if the clamping force is measured or estimated (e.g. from electricity) the measurement has tolerances. In the case of a brake in which the movement of the actuator and the movement of the lining are linked via a constant transmission ratio, the actuator torque would change significantly with the contact pressure position, which of course can still be an application for the energy method described here. However, it is recommended as particularly advantageous to also use a so-called non-linear EMB, because with this the actuator torque does not change so much over the actuation changes than in the case of a linear and thus the accelerating or braking torque is better known in the event of deviations than in the case of a linear, or does not contain such strong deviations.

A motor controller (e.g. for BLDC, e.g. FOC) has a lot of the information required here, e.g. position, speed, torque (e.g. from torque-generating current) or can be supplemented with additional necessary information such as mass inertia, the expected clamping force from the brake (or the suspected or determined from measurements). It is therefore recommended that the explanations for Fig. 30 are also obtained in direct cooperation with the engine controller or the information available here for the parameters sought (e.g. losses), which of course does not have to be permanent, but can also take place on a case-by-case basis. The application of the above energy conversion or the sum of moments is of course also recommended.

Fig.3101 -3102 show suggestions for symmetrical actuation of both pads on disc brakes, similar to drum brakes (e.g. as in Fig.29): With electromechanically actuated disc brakes, a spring can push the pads apart again when they are not braked, but the lifting process of both pads ( such as with drum brakes) may not be carried out. For this reason, it is proposed on the one hand that a lift-off process is carried out in the same way as for drum brakes, even with a disc brake with, for example, two springs against a fixed part (e.g. wheel bearing part) and that the disc brake is actuated as symmetrically as the drum brake above and the wear is adjusted symmetrically (symmetrical wear adjustment 02 in Fig .3101 ), or use one-sided behavior (as shown below) for simplicity. In the case of drum brakes, a wear adjuster can move apart, e.g. at the lower pivot point of the pads, similar to a cable tensioner with left-hand and right-hand threads when the middle part is turned. This point does not exist with disc brakes. It is therefore proposed, for example, to use a double-acting wear adjuster (02 in FIG. 3101) with two spreader parts 051 or, as in FIG .

The wear adjuster 02 Fig.3101 or the expanding part 051 Fig.3102 can be more or less elastic (or in such a way that a compensating movement is possible) with a be connected to a fixed part 09 (indicated by the curved connection from 09 upwards), whereby with more elastic there is better compensation for any asymmetries (both contact pressure forces or in geometries or wear), but with more rigid attachment asymmetries are "ridden" more quickly, e.g the linings wear out faster so that better symmetry is achieved. In Fig. 3101, however, the expansion parts 051 will probably be operated together (could also be of different strengths), in Fig. 3102, however, the two wear adjustment devices 02 will probably be adjusted together (could also be of different strengths). The above has the disadvantage that the full clamping force is on all parts (spreading part, wear adjusters), even on the duplicate parts.

In Fig. 3201 -3202 it is therefore proposed that the center-related drum brake-like adjustment can also be achieved in a different way: A compensating movement is derived from just one wear adjuster 02, with which the migration in the event of wear is compensated: The arrow from the center of the disc in Fig. 3201 on the distance of the expansion part pivot point 057 to a reference, eg center of the pane (possibly also eg pane surfaces). In Fig. 3202, the arrow points to the unworn starting position as in Fig. 3201, but you can see that the pivot point of the expanding part 057 has migrated to the left due to wear, which is visible as an arrow in the "fixed part" 09 in Fig. 3202. The wear adjuster in Fig. 3202 can produce exactly this offset (arrow in 02). For this purpose, the wear adjuster can have two threads, for example: one with a larger pitch that carries the full clamping force and a second with, for example, half the pitch that only has to carry the load of the center guide (the fulcrum). How the necessary “displacement of the pivot point of the expanding part 057 towards fixed” is carried out can be chosen, so any suitable means can be used that cause the displacement. The above applies again to the elastic guidance of the fulcrum and the movement geometries. There are many possibilities (not explicitly shown here) to generate this partial movement (eg half of the wear adjustment), eg with lever reduction. One can also indicate many possibilities here, how a basic setting is brought about, on which the partial movement is then built up: e.g. exact production in cooperation with lining wear (the Residual inaccuracies compensated), adjustment options (e.g. with adjusting screw), sluggish displacement option, with which an initial state is set when the brake is applied vigorously, etc. It is also possible to only bring about the basic setting during production and then to use precise brake pads, whereby it is not difficult, for example is to grind the linings together with the carrier plate to a precise thickness during production, for example, at least to the same pair in pairs.

In Fig.3301, for example, the possibility is shown that an adjustment with a clamping screw (indicated by the fixed part 09 going) is possible, which can be adjusted ex works or when changing the lining for the correct air gap on both sides or, for example, if the brake tries to find the right position by applying it.

Fig. 3302 shows that there is no need for an operation (e.g. clamping), but that the adjustment can also be made automatically (e.g when tightening) can take place. It is proposed here that this "take place automatically" must of course not only take place on defined occasions (e.g. changing the pads), but of course also more often or with every brake actuation. This means that the above "moving the fulcrum towards fixed" can also be combined or combined with the adjustment option and an additional device for "moving the fulcrum towards fixed" becomes unnecessary or combined with the automatic adjustment option (e.g. pressure spring) and it remains with a comparable effect only one setting that does not require operation (e.g. pressure spring). Of course, this can be applied to many brakes, for example drum brakes, where this possibility can also be between the brake shoes or at the non-actuated end of the brake shoes or with floating caliper brakes. The large upper dot in Fig.3301 -3302 simply represents a general part of the brake, which is adjusted by the automatic or non-automatic adjustment. In Fig. 34 it is shown that the transmission ratio of the expansion part should be defined and not subject to unexpected changes, expressed by the dashed curve of a desired transmission ratio with linear stroke on the x-axis versus the angle on the y-axis. The actual contact point of the rotated contact surface 0591 on a non-rotated contact surface 0592 should be well defined, which can be achieved with precisely manufacturable round parts (e.g. cylindrical pins), but only worse if curves are created by milling, for example. On the left in FIG. 34 it is shown that a rolling circle, which rotates around the pivot point 057 of the expanding part, results in such a defined relationship between the angle and the linear stroke in the case of a circular pressing movement, for example. If, however, it is not a circle that rolls off, but something else, such as the stairs shown, which could have been created by milling, the connection between angle and path is superimposed by a disturbance. This disturbance naturally influences the contact pressure, because the force transmission ratio (due to the change in lever length) is disturbed, and the contact pressure resulting from the contact pressure travel and the elasticity, because the contact pressure travel (due to the change in lever length) is disturbed. This disturbs the control of the brake. For this reason, it is proposed here to use rolling parts that can be produced well, accurately and inexpensively, such as cylindrical pins or parts for this purpose, which of course make geometric specifications due to their load-bearing dimensioning. This geometry requirement can be consciously accepted here, even if no minimum of the lateral compensatory movement is achieved. The proportion of the disturbance naturally depends on the extent of the geometric inaccuracy in relation to the total stroke of the pressing movement. Therefore, with a small stroke, for example, a geometry that is as precise as possible (eg circular shape) can be advantageous, for example using ground cylinders, but with a larger stroke, for example, a forged, pressed, cast, etc. contour can be sufficient.

In an embodiment that is not shown, the braking device comprises an actuator 04, in particular an electric actuator 04, a transmission, a spreading device, a brake pad 063 and a friction surface. The actuator 04 moves within a limited range of actuator actuation. In at least part of its actuator actuation range, the actuator 04 rotates the spreading device about at least one pivot point via the transmission.

According to this embodiment, the actuator 04 presses the brake pad 063 in at least part of its actuator actuation area via the spreading device for braking in order to generate a contact pressure force and a braking torque resulting therefrom in the direction of and against the friction surface.

Furthermore, the transmission has a non-linearity 03, ie a non-constant transmission ratio over at least part of the actuator actuation range, and rotates the spreading device in accordance with the non-linearity.

The invention is not limited to the embodiments shown, but includes any braking device and any machine according to the following patent claims.

Claims

patent claims

1. braking device,

- wherein the braking device comprises an actuator (04), in particular an electric actuator (04), a gear, a spreading device, a brake pad (063) and a friction surface,

- wherein the actuator (04) moves in a limited actuator actuation range,

- wherein the actuator (04) rotates the spreading device about at least one pivot point in at least part of its actuator operating range via the transmission,

- and wherein the actuator (04) presses at least in part of its actuator actuation area via the expanding device, in particular for braking, the brake pad (063), in particular to generate a contact pressure and a braking torque resulting therefrom, in the direction of and/or on the friction surface, characterized,

- that the transmission has a non-linearity (03), i.e. a ratio that is not constant over at least part of the actuator actuation range,

- and that the gearbox rotates the spreader according to the non-linearity.

2. Braking device according to claim 1, characterized in that

- that the spreading device is at least partially surrounded by the braking device, in particular the transmission,

- and/or that the spreading device is arranged loosely in the braking device,

- And/or that the spreading device is arranged in the braking device.

3. Braking device according to claim 1 or 2, characterized in that

- that in at least part of the actuator actuation area, in particular in a first actuation point of the actuator (04) or first actuation area of the actuator (04), the spreading device relative to the brake pad (063), the brake pad (063) pressing parts of the braking device, the actuator (04) and/or the, in particular stationary, transmission parts, performs a relative movement,

- the relative movement of the spreading device possibly, in particular exclusively, taking place along and/or in the plane of rotation of the spreading device,

- where the relative movement of the spreading device optionally, in particular exclusively, takes place essentially normal to the direction of rotation, in particular the pressing direction, of the spreading device,

- where the relative movement of the spreading device takes place, if necessary, in particular exclusively, in the longitudinal direction and/or transverse direction of the spreading device,

- Wherein the relative movement of the spreading device optionally takes place in all directions of extension of the spreading device. Braking device according to one of the preceding claims, characterized in that

- that the spreading device has at least one contact surface (0591),

- that the braking device, in particular the transmission and/or parts of the braking device that press against the brake lining (063), comprises at least one abutment surface,

- that the at least one contact surface (0591, 0592) presses against the at least one abutment surface in at least part of the actuator actuation area, as a result of which the spreading device rotates and/or moves,

- and that the pressing surface (0591, 0592) and the abutment surface are designed in such a way that these surfaces perform a relative movement, in particular a sliding and/or rolling movement, in particular during the rotation and/or movement of the spreading device. Braking device according to one of the preceding claims, characterized in that

- that the actuator (04) rotates the spreading device about a first pivot point via the gearing in at least part of its actuator operating range, in particular in a second operating point of the actuator (04) or second operating range of the actuator (04), - and/or that the actuator (04) rotates the spreading device about a further pivot point via the gearing in at least a part of its actuator actuation area, in particular in a further actuation point of the actuator (04) or a further actuation area of the actuator (04),

- and/or that the position of at least two pivot points deviates from each other,

- and/or that the position of the pivot points is limited by the design of the braking device,

- and/or that the braking device is designed in such a way that the displacement of the pivot point of at least two pivot points of the spreading device is opposed by an elastic resistance, in particular a resistance device,

- And/or that at least one pivot point is mounted and/or freely movable, in particular not mounted. Braking device according to one of the preceding claims, characterized in that

- that the spreading device comprises at least two spreading device parts, with at least one spreading device part optionally being a pin, a spigot and/or a prefabricated part,

- and/or that the at least one contact surface (0591, 0592) of the spreading device is formed at least partially from a spreading device part,

- and/or that the at least one contact surface (0591, 0592) of the spreading device is at least partially arranged on a spreading device part,

- And/or that the spreading device parts are connected to one another, in particular in a non-positive, materially bonded, pressed and/or welded manner. Braking device according to one of the preceding claims, characterized in that

- that the spreading device is non-linear,

- and/or that the spreading device is rotated by the actuator (04) via the gearbox in a limited range of rotation, where appropriate, the expanding device has at least one non-linearity (03), i.e. a non-constant translation over at least part of the rotation range. Braking device according to one of the preceding claims, characterized in that

- that the gear ratio of the transmission is selected and/or designed in such a way that the actuator (04) is operated in at least a partial area of its actuator operating range at an operating point that deviates from the optimum operating point of the actuator (04),

- and that the actuator (04) is operated in at least a partial area of its actuator activation area at an operating point that deviates from an operating point of maximum power of the actuator (04). Braking device according to one of the preceding claims, characterized in that

- that the transmission, in particular the spreading device, starting from a first position, in particular a zero position, of the transmission executes or implements a movement of the actuator (04) in a first direction for braking,

- and/or that the gearing, in particular the spreading device, starting from a first position, in particular a zero position, to adjust an air gap (068), in particular to actuate a wear adjustment device (02) and/or a wear adjustment device, a movement of the actuator (04) in executes or implements a second direction, in particular one that is opposite to the first direction. Braking device according to one of the preceding claims, characterized in that

- that a wear adjustment device is provided in the pivot point of the spreading device,

- and/or that the wear adjustment device comprises a drive, - and/or that the spreading device comprises a drive, and where the wear adjustment device may be provided in the drive of the spreading device,

- and/or that a wear adjustment device is provided between the actuator (04) and the transmission or between the transmission and the spreading device,

- and/or that the braking device comprises a wear adjustment device which is actuated, in particular exclusively, by the actuator (04), the transmission and/or the spreading device. Braking device according to one of the preceding claims, characterized in that

- that the actuator (04), the gear and/or the spreading device is set up for braking and wear adjustment (02), in particular for actuating a wear adjustment device,

- and/or that the braking device comprises only a single actuator (04) for braking and for wear adjustment (02), in particular for actuating a wear adjustment device. Braking device according to one of the preceding claims, characterized in that

- that the actuator (04) comprises several parts,

- and/or that the actuator (04) comprises a spring, in particular an actuating spring (042), and an electric motor (041), the spring and the electric motor (041) possibly being independent of one another in terms of components and/or direction of action, and/or where appropriate, the spring interacts with the electric motor (041) via at least one further component and/or via the transmission,

- and/or that the actuator (04) comprises two electric motors (041),

- And/or that the braking device interacts with at least one electrical machine, in particular at least one electromagnetically excited electrical machine. Braking device according to one of the preceding claims, characterized in that

- that the gearbox includes kinematic devices,

- and/or that the gearing comprises a cam, a ball ramp and/or a lever. Braking device according to one of the preceding claims, characterized in that

- that the transmission ratio, especially when braking, can be changed,

- and/or that the translation of the transmission can be changed, in particular actively, preferably by turning a ratchet,

- And/or that the transmission ratio of the transmission can be changed, in particular passively, preferably by spring-loaded retraction of components or the elastic deformation of components. Braking device according to one of the preceding claims, characterized in that the transmission is selected and/or designed in such a way that at least one partial section with a non-linearity (03) is formed and/or arranged along the actuator actuation area, and/or that the transmission is selected and /or is designed such that at least two sections with differently acting non-linearities (03) are formed and/or arranged along the actuator actuation area and/or that the at least one non-linearity (03) is selected from the following non-linearities (03): a. Non-linearity (03) to overcome an air gap (068) between the brake lining (063) and the friction surface, b. non-linearity (03) to determine the point of contact of the friction surface and the brake lining (063), c. Non-linearity (03) to achieve a minimum braking effect, d. non-linearity (03) to generate an increasing braking torque, e. Non-linearity (03) for operation with reduced electrical power requirements, f. non-linearity (03) to quickly achieve high braking effects, g. non-linearity (03) for measuring and/or adjusting parameters, h. Non-linearity (03) to reduce electrical and mechanical loads when starting the pad lift, i. non-linearity (03) to compensate for brake fading, j. Non-linearity (03) for wear adjustment (02). Machine, in particular transport device, vehicle, elevator or

Bicycle comprising a braking device according to one of Claims 1 to 15. Machine according to Claim 16, comprising a further, in particular electronic, braking device, characterized in that the further braking device is designed as a, in particular spring-loaded, parking brake.