# Patent

## Title
Self-amplifying disc brake and method for actuating a self-amplifying disc brake

## Abstract

A self-energizing disk brake is described. The brake includes an electric actuator in which an activation force is amplified using a self-energizing device arranged between the electric actuator and a brake lining. The system includes a brake application unit for applying the brake lining to a brake disk with an application movement of the brake lining, relative to the brake disk, having a component extending parallel to a rotational axis of the brake disk and another component extending tangentially to the rotational axis of the brake disk. At least one electromotive drive for activating the brake application unit and at least two brake plungers of the brake application unit are provided. Pressure surfaces of the brake plungers have a recess with a ramp shaped contour for engagingly receiving a rolling element.

## Background
BACKGROUND AND SUMMARY OF THE INVENTION
The invention relates to a self-energizing disc brake having an electric actuator in which an application force is amplified using a self-energizing device, and to a method for activating a self-energizing brake.
Self-energizing brakes are known in a wide variety of types. For example a classic design of self-energizing brakes are drum brakes in which a brake shoe is arranged in a leading fashion so that the friction forces between the brake lining and drum support the tensioning forces.
In contrast, in disc brakes it was assumed in the past that it was in fact a significant advantage of this design of brake that, with brake linings which act exclusively perpendicularly to the circumferential brake disc and on which only an activation device which acts parallel to the axis of the brake disc acts with a force which is aligned in such a way, there is no self-energizing effect. This was the case to an even greater extent in disc brakes for heavy utility vehicles in which the activation is preferably carried out hydraulically or pneumatically.
However, if disc brakes with activation devices which are operated electromotively are also to be used in relatively heavy commercial vehicles, the self-energizing disc brake becomes an option since it provides the possibility that, owing to the self-energization of the brake, the electric motor can be given smaller dimensions than would be possible with a non-self-energizing disc brake.
Self-energizing disc brakes are known in a wide variety of designs. However, the majority of solutions describe operational principles which permit self-energization but, due to a lack of price competitiveness as well as due to their awkward complex design, are not suitable for implementing a disc brake for heavy commercial vehicles which is ready for series production and can be manufactured economically, and often therefore have not passed the stage of theoretical ideas.
Against this background, the present invention provides an electromechanically operated, self-energizing disc brake which can be manufactured cost-effectively with a simple design. It preferably also provides the advantage that the power demand of the electromotive drive is minimized compared to comparable, directly electromechanically activated brakes by using efficient self-energization, even in the boundary region of the coefficient of friction of the brake lining.
The invention achieves this result using a brake application unit. Having two or more brake plungers, each plunger having pressure surfaces with a recess and ramp shaped contours into which engage rolling elements.
Accordingly, each of the pressure surfaces of the at least two or more brake plungers is provided, on its side facing the brake lining unit, with a recess with a ramp shaped contour into each of which a rolling element engages which is supported at one end on the ramp shaped contour of the pressure surfaces of the brake plungers and at the other end on the brake lining unit.
In this way, the self-energizing brake is easily combined with an adjustment device. The brake is applied in the direction of the brake disc or perpendicularly to the brake disc by the brake plungers by changing their length in the axial direction. The application of the brake in the circumferential direction or parallel to the disc is carried out in a different way, for example with a crank. The brake plungers can also be used to adjust for brake lining wear. For braking operations this advantageously results in a very wide variety of options which will be explained in more detail below.
The invention preferably also implements a configuration of the brake application unit such that it converts uniform rotations of an output shaft of the electromotive drive during a brake application movement into a movement of the brake lining, the movement component of which movement is nonlinear at least in the tangential direction (direction U).
Advantageous embodiments of the invention will be described below.
The invention not only reduces the manufacturing costs of a brake system for utility vehicles but it also significantly minimizes the power demand of the electromotive drive in relation to comparable directly electromechanically activated brakes by using efficient self-energization even in the boundary region of the coefficient of friction of the brake lining. According to particularly advantageous embodiments, it is possible to significantly reduce the power demand compared to other known self-energizing concepts.
It is also possible to meet the same power requirements compared to modern compressed air brakes and also to satisfy the same predefined installation conditions and weight prescriptions.
The adjustable ramp system can also be used to implement a reliable parking brake which adjusts automatically even when friction elements shrink owing to cooling. A further significant advantage of the invention is therefore the fact that with the proposed disc brake a reliably acting parking brake is also implemented without additionally necessary activation components.
For this purpose, the ramp angle with the greatest degree of self-energization must be dimensioned in such a way that self-energization is possible even at the lowest envisioned coefficient of friction of the brake lining.
When the brake is inserted, there is therefore an exclusively mechanically holding effect of the brake. If brake linings and/or the brake disc shrink or if there is a drop in the coefficient of friction which occurs during the shut-off phase, the brake and the self-energization of the brake are automatically adjusted in order to keep the vehicle in a stationary state.
The electromotive drive is preferably coupled to an open-loop and/or closed-loop control device which is configured to perform respectively open-loop or closed-loop control of the position of the actuator element or brake lining. In this context, the position of the brake lining unit is set according to predefined values of a superordinate unit (for example a control device).
This open- and/or closed-loop control device is preferably operated as follows:
The basis of the preferred closed-loop control concept is a braking or deceleration closed-loop control process of the vehicle such as is customary in contemporary EBS closed-loop-controlled vehicles with a compressed air brake system.
In such brake systems, the driver or an autonomous vehicle system presets a braking request or deceleration request which is converted into a “braking” signal which is processed by the EBS system and converted into a corresponding actuation of the wheel brake actuators (pneumatic cylinders or electric motor) which brings about corresponding activation of the brakes.
In pneumatically activated disc brakes, a pure pressure control process of the activation cylinder of the respective brake is usually carried out according to the relationship:

    
    
        brake pressure→cylinder force→tensioning force frictional force which can be determined and reproduced within sufficiently tight limits of precision.
    
    


In self-energizing, electromagnetically activated brakes, this sufficient precision is generally no longer provided between the actuator manipulated variable and the frictional force.
The motor current is frequently used as the activator manipulated variable of electromechanically self-energizing brakes of this kind. However, such large tolerances of the achievable braking effect arise from the engine efficiency levels, which are for example also temperature dependent, and from the efficiency level of the step down gear mechanism as well as finally the efficiency level of the amplification method in conjunction with the variations in the coefficient of friction of the brake linings, that it no longer appears possible to control the braking effect by using the motor current.
It has been proposed to measure the frictional force and carry out closed-loop control on it directly for example in International patent document WO 03/100282 describing the self-energizing brake which has wedge activation and is described in European patent document EP 0 953 785 B1.
According to this method there is the problem of finding a suitable measuring method for determining the frictional force. Furthermore, there is the difficulty of the frictional source being influenced to a very high degree by brake oscillations and wheel oscillations and thus constituting a controlled variable which can be controlled only with difficulty.
The aim is therefore to find a closed-loop control method for self-energizing brakes which is well suited in particular also to disc brakes and which avoids the problems associated with closed-loop control of the frictional force.
To summarize, the present invention implements a method for actuating a self-energizing brake in which an activation force applied by the actuator is amplified using a self-energizing device arranged between the actuator and brake lining, wherein the actuator is coupled to an open-loop or closed-loop control device which is configured to actuate the actuator in order to set the position of the brake lining units in such a way. The present method is distinguished by the fact that during the closed-loop control process tolerance-conditioned braking force differences, referred to as third controlled variable, among the wheel brakes on which the closed-loop control process is performed by the brake system, are determined and compensated.
The invention also provides a method for carrying out a parking braking operation, with a brake according to the invention in which during a parking braking operation the brake is applied in a simple manner solely using the brake plungers until the rolling elements have moved the lining units against the disc, after which the self-energizing effect starts without the crank being activated.
The brake plungers or pistons can also be used for relatively minor compensation braking operations.
As far as possible, sensor systems which are already present and are reliable and proven are used to sense the signals which are necessary for the closed-loop control.
A First Variant Will be Explained First
Solution 1: Third Controlled Variable
The solution described below provides a brake system in which, between the “braking or deceleration” vehicle controlled variable and the “current or actuator position” actuator manipulated variable, a third controlled variable is introduced which is intended to compensate the tolerance-conditioned braking force differences among the wheel brakes on which closed-loop control is carried out by the brake system.
This third controlled variable is sensed individually for each vehicle wheel and compared with the values determined at the other wheels.
When there are inadmissible deviations from the defined values of the EBS system, these predefined values (motor current or actuator position) for the individual brakes have a correction factor superimposed on them individually, with which correction factor the existing braking force differences are compensated.
This adaptation process is carried out, if appropriate, in relatively small increments over a plurality of brake activation processes.
The wheel slip of the respective vehicle wheel is preferably evaluated as a third controlled variable.
In this method it is advantageously unnecessary to generate a precise relationship between the wheel slip and braking force, but rather the wheel slip characteristic variables which occur at the individual wheels are adjusted to form a specific predefined set point value for the EBS system. In particular, in this context the wheel slip characteristic variables of the brakes of the individual axles are adjusted as precisely as possible. The matching of the wheel slip characteristic variables of the axles to one another takes place in a second step taking into account the possibly different predefined values of the brake system for the individual axles.
Alternatively, the tensioning force acting on the brake can also be determined as the third controlled variable. The tensioning force can be determined at the components of the brake which pick up force, for example at the brake caliper, by measuring deformation paths or component stresses. In the process, the necessary sensor can be arranged in the interior of the brake and integrated, if appropriate, into an electronic control system which is arranged within the brake.
Solution 2: Open Loop Control Using Actuator Position or Motor Current Combined with Tolerance Compensation
A second approach to the solution is based on the existing control algorithm of contemporary EBS systems in which only the actuator manipulated variable of pressure is replaced by another system-specific manipulated variable. The actuator position and motor current are particularly appropriate as system-specific manipulated variables.
In the description of the prior art, the large tolerance variation, which makes this method more difficult to apply, has already been mentioned. It is therefore necessary to largely eliminate the tolerance influences present in this effect chain.
This is preferably brought about with one or more of the following measures:

    
    
        Before the brake actuator is activated, the venting play is overcome by using the adjustment device so that the venting play is already no longer present as a fault source at the start of the actual brake application movement by the brake actuator.
        The influence of the brake lining compression—which differs due to the wear state and temperature state—on the predefined set point values of the brake system is compensated by correction factors. For this purpose, the wear state of the two brake linings is determined precisely for each brake. Likewise, by evaluating the energy balance of the brakes, their thermal content and hence also the temperature of the brake lining are determined. This energy balance can be evaluated by the electronic brake system or by an electronic open-loop controller which is integrated into the brakes.
        Brake-specific variations in the relationship between the tensioning force and the widening of the caliper are compensated by a calibration process when the brakes are fitted. For this purpose, defined forces are applied to the brake caliper, for example during the final inspection on the assembly belt, and the widening which occurs in the process is determined or the actuator adjustment travel necessary for this is determined directly. The defined application of force is preferably carried out in such a way that force pickups are used in the brake caliper, for example instead of the brake disc, and the actuator is then actuated in order to generate the predefined tensioning forces. The relationship between the tensioning force and the actuator position which is detected in this way can then be stored, for example in an electronic system which is integrated into the brakes.
        When the motor current is applied as an actuator variable, the tolerance compensation which is described can be applied in the same way. The relationship between the tensioning force and the motor current is then determined during the calibration process and stored as described above. During this calibration process, the tolerance influences of the gear mechanism and electric motor are also eliminated at least for room temperature conditions. The temperature influence on the electric motor, for example on its permanent magnet, can in turn be compensated by the abovementioned thermal balance calculation.
    
    


The resulting normal force for a specific position of the self-energizing device is dependent on a large number of factors such as

    current venting play
    rigidity of the brake (caliper) perpendicular to the frictional surface
    in particular the variable rigidity of the lining which is dependent on


locations of linings
wear state, that is to say residual thickness
temperature
prior history (effect on compressibility)
take up of moisture

    
    
        variable temperature of caliper and disc during the braking process
        coefficient of friction between the brake lining and brake disc (effect on self-energizing effect and thus also on the normal force and on the frictional force). This is itself dependent, inter alia, on
    
    


temperature
speed
According to the teaching of the invention, selective actuation of the ramp position in order to bring about a specific pressing force is virtually impossible if the influence of the aforesaid parameters is disregarded entirely.
In contrast, by virtue of the invention, a desired brake lining pressing force can be brought about by selective travel control of the self-energizing device or of the brake lining, and it is thus possible to dispense with a difficult-to-implement adjustment of the setpoint value to the actual value of the frictional force or else to permit selective pilot control for a brake with a setpoint value/actual value comparison of the brake lining pressing force or other frictional force.
According to the invention, this is achieved by virtue of the fact that interference variables which influence the correlation between the ramp position or brake lining position and the brake lining pressing force are compensated by taking into account relevant parameters.
For this purpose, a characteristic curve is determined which defines a corresponding pressing force in accordance with a position of the self-energizing device, for example a ramp, or an actuation travel which is predefined by the actuator.
This characteristic curve is preferably updated continuously, in order, for example, to be able to take into account influences such as temperature and speed.
The application point of the brake lining on the brake disc is determined, for example using the current of an electric actuator or by calculating it from the current venting play and ramp geometry.
The positive gradient of the characteristic curve is adapted as a function of the ramp position or brake lining position to:

    a) Rigidity of the brake (caliper) perpendicular to the frictional surface can be determined experimentally or by calculation and is substantially constant.
    b) In particular the variable rigidity of the lining which is dependent on
    
        brake lining locations either by specification within a tolerable framework or by inputting/selecting corresponding parameters in an electronic control device when the brake lining is changed.
        wear state, that is to say residual thickness is sensed continuously (as in the prior art)
        temperature either by measurement or by calculation, for example using energy integration, cooling power etc.
        prior history (such as effect on compressibility) logging of the prior history of the brake lining (aging), for example using energy integration, maximum temperature or the like. Relationship between the rigidity of the brake lining and aging can be determined empirically.
    
    
    c) The temperature of the caliper and disc which varies during the braking process either by measurement (for example thermal elements) or calculation.
    d) Coefficient of friction between the brake lining and brake disc (effect on the self-energizing effect and thus on the normal force and on the frictional force).
     This is itself dependent, inter alia, on
    
        temperature
        speed
        empirical determination of the dependents.
    
    


Alternatively or additionally, closed-loop control of the brake can also be carried out by determining the normal force which acts between the brake lining and disc. The normal force can be determined, for example, by sensing the expansion of the calipers. If the actual normal force deviates from the desired normal force, the latter can be adapted by the described travel/force characteristic curve.
The brake application unit or ramp can be implemented in a defined fashion by an angle, either as a pressure ramp, traction ramp or traction/pressure ramp. In the case of a traction/pressure ramp in particular a self-locking system is advantageously selected as a drive, i.e. a high force which results in the direction of the activation from an unusually high/low coefficient of friction cannot lead to uncontrollable displacement of the ramp.
The described compensation of the interference variables can also be used for directly activated systems (activation force=support force).
As a further embodiment of the invention, there is provision for the electric motor to rotate a crank directly or by using at least one or more gear mechanisms, the crank having a crank tappet as output element which serves to move the brake lining unit, and the crank tappet is oriented parallel to the axis of the brake disc. The arrangement is compact and easy to implement in structural terms.
In this context, the electric motor preferably has an output shaft which is oriented parallel to the axis of the brake disc and by which the crank which acts on the brake lining unit is rotated directly or using further, intermediately connected gear mechanism elements.
According to a further variant of the invention, the brake application unit also has at least one, in particular two or more, brake plungers, used as adjustment for pistons, which are oriented parallel to the axis BA of the brake disc and which are supported at one of their ends on the brake caliper or by a bearing device on a component which is connected to the brake caliper, the bearing device permitting in each case at least some of the brake plungers to rotate about their longitudinal axis.
To summarize, the following advantages are provided, each independently and also in combination:

    Circumferential activation by a crank


coaxially arranged drive unit
preferably integrated electronic control system

    Simple combination of spindle actuation and/or crank activation application function by spindles—force stroke by means of crank application and adaptation braking using spindles crank activation for high-load braking operations parking brake function using spindles
    Reliable and uncomplicated parking brake function
    Pretensioning by spindles—automatic post-tensioning by the amplification system without crank activation.
    If appropriate additional post-tensioning by crank activation
    Addition of a highly amplifying ramp angle
    Use of a common drive Shiftable distribution gearing
    Automatically shifting (only application function with spindles)
    Extraneously shifted (parking function and partial load braking function with spindles)
    Variable self-energization
    Multi-stage, shiftable infinitely variable, automatically adaptive and/or extraneously controlled
    Controlled self-locking of the brake plungers
    Controlled, and in the event of faults automatic, switching over from self-locking to non-self-locking operation
    a) self-locking spindles and addition of a non-self-locking preliminary stage (folding ramp, spherical ramp etc.)
    b) non-self-locking spindles and addition of a preliminary stage which brings about self locking (self-locking gear stage etc.)
    Play-free drive


Measures for eliminating play in the force transmission path from the drive motor to the brake lining pressure plate.



BRIEF DESCRIPTION OF THE DRAWINGS
The invention is explained in more detail below with reference to the drawings using the exemplary embodiments. The said drawings:
FIGS. 1-3 are each sectional view sketches which illustrate the basic design of different variants of disc brakes according to the invention, with FIG. 1a showing an enlarged view of the mechanism at the center of FIG. 1;
FIG. 4 is a further sketch illustrating the basic function of the disc brakes shown in FIGS. 1 to 4;
FIG. 5 shows a sketch of a further embodiment of a disc brake according to the invention as a sectional view; and
FIG. 6 shows multiple views of a partial region of a pressure plunger, further shown in FIG. 6a with a pressure element for application against a brake lining unit.



DETAILED DESCRIPTION OF THE DRAWINGS
The exemplary embodiments shown in the figures are described below with their essential features.
The functional principle in FIG. 1 with a nonadjustable ramp gradient in the pressure surfaces 5, 6 of the brake plungers will be described in more detail first.
FIG. 3 shows a variant with an infinitely adjustable ramp gradient and FIG. 2 shows a variant with an incrementally adjustable ramp gradient. FIG. 4 illustrates the basic functional principle in conjunction with FIG. 1.
The disc brakes according to the invention are preferably based on a fixed caliper concept in which a single-part or multi-part brake caliper 1 (also referred to as brake housing) is attached to a wheel axle so as to be unmovable in relation to a brake disc 2.
According to the concepts described below, is described a fixed caliper brake with at least external—reaction-side—electromechanically activated and electronically closed-loop controlled adjustment for wear, which can be supplemented by a brake application side adjustment for wear using the brake plungers if the latter is used for this purpose.
The functional principle and the features described can theoretically also be applied for other types of brakes such as, for example, sliding caliper brakes or pivoting caliper brakes. Only the caliper head of the mechanically/pneumatically activated basic brake is replaced by the electromechanical brake application unit with self-energization.
A fixed saddle brake with a pneumatically actuated brake application unit of this type is presented, for example, by German patent documents DE 36 10 569 A1, DE 37 16 202 A1 or European patent document EP 0 688 404 A1. A fixed caliper brake with an electromotive adjustment element is presented by International patent document WO 02/14 708 A1. Such electromotive adjustment devices can be arranged when desired in each case at least on the reaction side in the proposed exemplary embodiments.
In FIG. 1, the brake caliper 1 is indicated only in its brake application side region. In practice, it preferably engages around the upper circumferential region of the brake disc in the manner of a frame and is attached to an axle flange (not shown here).
The brake caliper 1 has, on its side facing the brake disc 2 with a brake disc rotational axis, one or more, preferably two, openings 3, 4 and a corresponding number of brake plungers 5, 6 (two here) which are oriented parallel to the axis BA of the brake disc.
According to FIGS. 1 to 4, two brake plungers 5, 6 are arranged parallel to one another in each case.
The two brake plungers 5, 6 or adjustment pistons are each supported directly or by intermediately connected elements; plain bearing shells 9, 10 here, are supported on the rear wall 11, facing away from the brake disc, of the brake caliper. Preferably balls 7, 8 with plain bearing shells 9 are used as bearing devices.
The bearing devices are configured in such a way that they permit the brake plungers 5, 6 or adjustment pistons to rotate about their own longitudinal axis LA.
In this context, spherical segment-like (connecting link-like) recesses are formed in each case in the brake plunger 5, 6 and in the brake caliper, in one of which recesses the plain bearing shells 9, 10 are inserted (into the one in the brake caliper here) so that the balls 7, 8 can rotate relative to the plain bearing shell.
Alternatively, the balls 7, 8 can also be embodied as spherical shoulders at the ends of the brake plungers 5, 6 facing the brake caliper (not illustrated here), which ends then engage in corresponding recesses in the brake caliper with plain bearing shells.
Flat plain bearings or annular bearings or the like (not illustrated here) are also conceivable instead of the balls and recesses.
The brake plungers 5, 6 each have a spindle 12 which is provided with an external thread and on which a sleeve-like nut 13 with a corresponding internal thread is rotatably arranged. This thread can be non-self-locking or self-locking depending on the configuration.
At their side facing away from the brake disc, the nuts 13 have a flange 31, compression springs 32 which concentrically embrace the nut 13 and exert a predefined force on the flange or prestress the flange 31 relative to the inner wall of the brake caliper acting in each case between the flange 31 and the inner wall of the brake caliper 1.
Alternatively, the entire mechanism is prestressed against the pressure plate.
According to FIGS. 1 to 4, the nut 13 of each brake plunger 5, 6 is arranged on the side facing the brake disc, and the spindles 12 are arranged on the side facing the interior of the brake caliper. An inverted arrangement may also be envisioned (not illustrated here).
By screwing the nut 13 onto the spindle, the axial length of each individual brake caliper 5, 6 which is embodied in this way can be adjusted, for example in order to compensate wear of the brake lining and when the brake linings come to bear against the brake disc 2.
On the side facing the brake disc, that is to say on the pressure faces, the brake plungers 5, 6, here the nuts 13, are each provided with a ramp-like recess or contour 14 whose lowest point is preferably in the region of the longitudinal axis of the brake plungers.
This is shown particularly well in FIG. 6. According to FIG. 6, two “raceways” or contours 14a, 14b which are rotated through 90° with respect to one another are provided in the nuts 13 of the brake plungers 5, 6 and they each have a different angle +α1, −α1; +α2, −α2 of aperture with respect to the longitudinal axis LA. Either one or other of the contours 14a, 14b can be used by rotating the brake plungers 5, 6 through 90°—the nuts 13 here—about their longitudinal axis when the brake is applied, which gives rise to different behaviors of the brake (this is explained in more detail below). One of the halves of each “double ramp” +α; −α can be used in the forward direction of travel and the other in the rearward direction of travel.
The recesses or raceways 14 are embodied in a spherical shape with a constant angle α of aperture or ramp angle α with respect to the longitudinal axis LA or else, for example, according to a particularly preferred variant they are preferably embodied in the manner of a variable spherical surface contour, for example oval spherical contour whose ramp angle α relative to the longitudinal axes LA of the brake plungers 5, 6 varies in the circumferential direction (relative to the longitudinal axis LA of the brake plungers)—for example incrementally (FIG. 6) or continuously.
In each case rolling elements 16 which are embodied as balls 16 in a preferred configuration here engage in the recesses 14.
The lowest point 14c (FIG. 6) of the ramps is advantageously configured in such a way that the rolling elements, in particular balls, are latched at the lowest point and can be released from the depression only counter to a minimum latching force, which ensures a defined rest state.
Alternatively, according to one alternative embodiment, rolling elements (for example barrels) may be envisioned which are cylindrical or shaped in some other way and which would then roll, for example, on a groove-like recess in the brake plungers. However, it would not be possible to implement all the embodiments of the invention which are represented in FIGS. 1 to 3, as will become clear below. However, it would be possible to implement an exemplary embodiment in the manner of FIG. 1 with a groove in the pressure surfaces.
The rolling elements 16 engage on their sides facing away from the brake disc in plain bearing shells 17 which are constructed in accordance with the embodiment of the rolling elements, in the manner of spherical heads here, and they are inserted into recesses with a corresponding shape in a pressure plate 18 which bears against the carrier plate 19 of a brake application side brake lining 20 with brake lining material 21 which is arranged in the brake caliper 1 parallel to the rotational axis BA of the brake disc, and so as to be movable in the circumferential direction U (or tangential or parallel direction with respect to the tangential) in relation to the brake disc 2.
A clamping spring 22 between the pressure plate 18 and nuts 13 holds the pressure plate 18 against the nuts 13 under prestress. Alternatively it is also conceivable to prestress the pressure plate in some other way, for example at the housing (caliper).
In order to drive the brake, an electric drive motor 23 is used, downstream of which a step down gear mechanism 24 is preferably arranged, the output shaft 25 of which step down gear mechanism 24 acts on a further gear mechanism 26, in particular a planetary gear mechanism which is arranged centrally between the spindles.
In this context, the output shaft 25 drives a sun wheel 27 of the planetary gear mechanism 26 which entrains planet gears 28. The planet gears 28 mesh (as illustrated in detail in FIG. 1a) with the sun gear 27 and an internally and externally toothed ring 29. Depending on the switched state (switching capability not illustrated here) they either cause the planet star 33 or the ring 29 to rotate. The ring 29 meshes an external toothing with gearwheels 30 which are fitted onto the spindles 12 or integrally formed thereon.
In order to switch over the drive (for example an electric motor) it is possible to provide a spring-loaded ball catch mechanism (not illustrated here). The switching over process can also be implemented in some other way (for example electromagnetically).
In an axial prolongation of the planet star 33, a crank 34 which is embodied in a cylindrical fashion here and arranged parallel to the axis of the brake disc is provided, the crank 34 engaging, on its side facing the brake disc 2, with a crank tappet 35, embodied off-center (eccentrically) and also oriented parallel to the axis BA of the brake disc, in a corresponding opening 36 in a brake lining unit, in which case the opening 36, which has for example a cross section which corresponds to the cross section of the crank tappet 35 or else is embodied, for example, in the manner of a connecting link, in particular an elongated hole (for example perpendicular to the plane of the figure here).
In the exemplary embodiment in FIG. 1, the actuation device or brake application unit is composed of the two adjustment pistons or brake plungers 5, 6 which have variable lengths for the purpose of adjusting for wear and which have, in their pressure surface facing the brake disc 2, the recesses 14 in the manner of ramp contours on which the rolling elements 16 run, the rolling elements 16 transmitting the brake application force generated by the brake to the brake lining unit or to the pressure plate which rests on the brake lining.
In the pressure surface of the brake lining unit or the pressure plate 18, the rolling elements 16 are held by an oppositely configured ramp profile (not illustrated here) or in the plain bearing bed (plain bearing shells 17)—illustrated here and preferred since the rolling bodies are guided particularly securely.
The brake lining unit, composed here of the single-part or multi-part combination of the brake lining 20 and pressure plate 18 is pressed in a sprung fashion against the brake plungers and adjustment pistons 5, 6 in such a way that the rolling elements 16 arranged between them are clamped in elastically between the brake lining unit and the brake plunger.
The brake is activated after an application process of the brake lining 20 against the brake disc by displacing the pressure plate together with the brake lining 20 parallel to the frictional surface of the brake disc in the direction of rotation or the circumferential direction thereof.
This displacement is preferably brought about by the crank drive 35, 36 which acts approximately centrally on the pressure plate 18 of the brake lining 18, 20 with an output tappet and crank tappet 35, and is mounted parallel to the axis of rotation of the brake disc in the brake application housing, brake caliper 1.
The crank drive is actuated by the electric drive, for example the electric motor 23, with a gear mechanism 24 arranged downstream.
FIG. 1 is defined by a constant ramp angle α. A particularly simple structural design is achieved here which is defined by a robust design, good functional reliability and low manufacturing costs. In particular, an electric motor 23 with surprisingly low output power can be used. In this context, the balls 16 can be inexpensive rolling elements which as it were orientate themselves in the ramp surface. In order to increase the load bearing capacity the balls can also run in adapted ball tracks.
A variant with rollers as rolling elements 16 would, in contrast, have a particularly small hysteresis (not illustrated here).
FIG. 2 differs from the variant in FIG. 1 initially in that the ramp angle of the recesses 14 in the circumferential direction about the longitudinal axis of the adjustment nuts or brake pistons 5, 6 is not constant, but rather variable so that, depending on the rotational position of the nuts 13, a steep ramp angle α with a different value is present. For this purpose, ball tracks with differing gradients can be arranged for the various rotational positions.
As a result, the brake application characteristic can easily be changed by rotating the nuts 13, for example by using a separate adjustment actuator 39, preferably of an electromechanical design (for example a further, relatively small electric motor) which rotates the nut or nuts 13 using an output shaft 40 with a gearwheel 41, for example by using the output wheel 41 to drive one of the nuts 13, for example on an external toothing of its flange, and by the other nut 13 being entrained by a belt drive 42 which is wrapped around both nuts 13.
In this way, a degree of self-energization which can be achieved even in the boundary region of the coefficient of friction can be increased from FIG. 1 to FIG. 2. However, the switching over can only take place in the released state since the nuts 13 cannot rotate during the brake application processes.
According to FIG. 3, the ramp angle changes continuously in a tangential fashion around the longitudinal axis LA of the brake plungers 5, 6. This is used for automatic angle adaptation.
For this purpose, a second further planetary gear mechanism 37 which is offset axially with respect to the first planetary gear mechanism 26 is connected, on the one hand, between the crank 34 and the first planetary gear mechanism 26 and is in turn arranged centrally between the brake plunger, which has an output ring 38 driven by the planetary gears 43 and which entrains the externally toothed nuts 13, while the planet star 44 of this planetary gear mechanism in turn drives the crank or rotates about its longitudinal axis.
In this way, the following operation during the application of the brakes is possible:
The application of brakes is divided into the following phases:
1. overcoming of the venting play,
2. build up of braking force,
3. release of the brake and
4. setting of the venting play.
Phase 1 Overcoming of the Venting Play
Before a braking operation, the initial situation is as follows.
First, the crank 34 is in a home position (FIG. 1) in which it is held, for example, by a spring-loaded ball catch (not illustrated here).
A frictional torque or holding torque which is always greater than the spindle frictional torque is applied to the adjustment nuts 3 in this situation by the compression springs 32.
First, the drive motor 23 rotates the spindles 12 in the direction of rotation which applies the brakes. The planet star 33 is locked here in the gear mechanism 26 using the latched crank. The external gear or the internally and externally toothed ring 29 rotates the adjustment spindles 12 in the brake-applying direction until the brake lining 21 comes to rest on the brake disc 2.
The adjustment nuts 13 are secured here against rotation by a sufficiently high holding torque.
As a result of the reaction force which builds up, the adjustment spindles or brake plungers 5, 6 become blocked against the brake disc 2 which is preferably movable, but not necessarily, and is axially movable in the case of a fixed caliper, which adjustment spindles or brake plungers 5, 6 come to rest on the lining on the reaction side (not shown here).
Phase 2 Build Up of Brake Pressure
As a result of the blocked adjustment spindles 12, the drive torque acting on the crank 34 now increases so strongly that it is released from the latched position.
The crank 31 now displaces the brake lining in the direction of rotation with respect to the brake disc 2 until the position predefined by the controller is reached (FIG. 4).
In the process, the movement component of the brake lining behaves in a nonlinear fashion in the circumferential direction—parallel to the frictional surface of the brake discs—or tangentially or parallel to the tangential U of the crank tappet, because a greater distance is firstly traveled on the circular path of the crank tappet in the circumferential direction per time unit than as the movement of the crank tappet 35 progresses on its circular path. The gear mechanism with the crank drive is therefore configured in such a way that the angular movement on the electric motor and on the output tappet in the circumferential direction is not converted into a linear movement of the brake lining but rather into a delayed movement.
Three cases are now to be distinguished.
Case 1
The current coefficient of friction of the brake lining corresponds sufficiently precisely to the tangent of the angle of inclination of the ramp in the recesses 14 or in the pressure surfaces of the adjustment nuts 13.
The predefined position is reached in this case with only a small expenditure of adjustment force.
Case 2
The current coefficient of friction of the brake lining is considerably larger than the tangent of the angle of inclination of the ramp in the recesses 14 or pressure surfaces of the adjustment nuts 13.
As a result of the excessively large self-energization, the brake lining 20 becomes stronger and is entrained further by the rotational movement of the brake disc than corresponds to the predefined position.
A rotational force in the direction of movement of the brake disc 2 is applied to the crank 34 by the brake lining.
Since the electric drive motor 23 holds the sun gear 27 of the planetary gear mechanism 26 and of the second planetary gear mechanism 37 in the desired position, the further rotation of the crank 34 and thus of the planet star 44 of the second gear mechanism 37 brings about a rotation of the outer wheel or internally and externally toothed outer ring 38 of the second gear mechanism 37 and thus also of the two adjustment nuts 13.
The holding torque of the two adjustment nuts 13 is overcome in the process.
As a result of the rotation of the adjustment nuts 13, the effective angle α of inclination of the ramp is changed in the direction of decreasing self-energization until the effective self-energization is adapted sufficiently precisely to the current coefficient of friction of the brake lining.
Case 3
If the current coefficient of friction of the brake lining is considerably smaller than the tangent of the angle α of inclination of the ramp in the recesses 14 on the pressure surfaces of the adjustment nuts 13, the brake lining is not sufficiently entrained by the low self-energization. A relatively high drive torque is thus necessary at the crank 34 in order to move the brake lining 20.
Owing to the reaction torque which becomes effective at the ring gear of the gear mechanism 24, the adjustment spindles 5, 6 are rotated in the direction of increasing the self-energization process until the tangent of the effective angle α of inclination of the ramp is moved again in sufficiently precise correspondence with the coefficient of friction of the brake lining.
Phase 3 Release of the Brake
In order to release the brake, the crank 34 and thus the brake lining 20 are moved back into the latched position by the electric drive motor.
The force necessary for this at the crank is low since the self-energization has been adapted in the previous braking process.
When the crank 34 latches into the latched position, a jump in torque is produced.
Evaluating the operational data of the electric drive motor (for example rotational speed, power drain) makes it possible to detect that the last position has been reached.
Phase 4 Setting of the Venting Play and Checking it
Since the crank 34 is now latched in a frictionally locking fashion, the gear mechanism 26 is actuated again as the electric drive motor 23 continues its backward rotational movement, and the adjustment spindles 12 are now rotated back by a defined amount by the gear mechanism 26 in order to release the brake and to generate the venting play.
By applying the brake linings 20 to the brake disc 2 in the first phase the venting play is checked, and by defined backward movement out of this position the venting play is set.
The measurement of the wear value is made possible by evaluating the position signal of the electric drive motor according to venting play settings.
The system which is provided for braking when reversing will be explained briefly below.
Braking when Reversing:
Forward travel and reverse travel are differentiated by suitable measures, for example corresponding evaluation of the rotational signal, for example of the wheel speed sensor (for example an ABS sensor) at a control device (not illustrated here) at or in the brake or at a superordinate control device of the brake system which is connected to the electric motor and/or actuates it.
After the termination of phase 1, the crank 34 is actuated in the rotational direction which corresponds to the rotational direction of the brake disc, by actuating the brakes. The invention will be considered once more from another direction below.
Firstly, the basic principle in FIG. 1 with a nonadjustable ramp gradient will be explained once more in more detail. In order to implement this embodiment, it is firstly necessary to provide a recess 14 with a ramp shape in the brake plungers 5, 6.
An opposingly shaped ramp is correspondingly formed in the brake lining pressure plate 18 or, better still, the rolling element 16 is rotatably mounted in the brake lining pressure plate 18, or a ramp is formed in the brake lining pressure plate 18 and the rolling element is mounted in the brake plunger (not illustrated here).
So that the rolling elements 16 run up on the recesses of the nuts 13 of the brake plungers 5, 6 and thus push the brake lining 20 against the brake disc it is necessary to bring about displacements of the brake lining pressure plate with the brake lining in the circumferential direction, preferably by using an adjustment element (here a crank 34) which is arranged coaxially with respect to the longitudinal axis of the brake disc and parallel thereto. The nuts 13 preferably do not rotate during the actual braking process.
A dual ramp profile (in the circumferential direction U and counter to the circumferential direction U) in the brake plunger 5, 6 permits a self-energization effect here in both directions of travel.
The crank drive 14 is preferably driven by the electric motor 23 with the gear mechanism 24, 26 connected downstream.
It is envisioned to provide a separate drive for the brake plungers or else to combine the adjustment drive and crank drive (FIG. 1 and FIG. 2). The latter has the advantage that only a single drive motor is required for both functions.
It is also envisioned here to overcome the venting play with the separate adjustment drive (phase 1 of the functional description).
It is also envisioned to overcome the venting play with the crank drive using a particularly “steep ramp” at the start of displacement.
The dual ramp profile (recess 14) in the adjustment pistons 5, 6 permits a self-energization effect here in both directions of travel. It is possible to implement control of the displacement of the brake lining as a function of the direction of rotation of the wheel.
According to FIG. 3, the recess 13 or ramp in the brake plunger is embodied as a truncated cone-like hollow element. The rolling element is in turn mounted in the pressure plate 18. It is thus possible to adapt the ramp gradient to the coefficient of friction of the lining by rotating the brake plungers 5, 6. The rotation of the brake plungers 5, 6 is carried out by a separate drive 39 or automatically by a branching gear mechanism 26 which can transmit the rotational movement generated by the drive motor 23 in output rotational movements both to the crank 34 and to the rotational device 12 of the brake plungers 5, 6.
The branching gear mechanism 26 is preferably a planetary gear mechanism. A displacement force which becomes active at the crank tappet 35 (displacement of the brake lining unit by using the crank 34 when the self-energization is too low or pulling of the crank 34 by the brake lining unit when the self-energization is too high) brings about reaction torques in the branching gear mechanism 26 and the reaction torques attempt to bring about rotational movements at the input shaft and/or at the brake plungers 5, 6. If a sufficiently high holding force is then applied to the input shaft (for example by the drive motor which holds the position of the input shaft using its electronic controller), rotation occurs at the brake plungers 5, 6.
Given a suitable assignment of the direction of rotation of the adjustment pistons 5, 6 to the direction of the application of force to the crank tappet, the ramp gradient is rotated to relatively steep ramp angles when the self-energization is too large (brake lining unit pulls on the crank tappet), and when the self-energization is too low (crank tappet pushes the brake lining unit) it adjusts to relatively obtuse ramp angles, i.e. with the effect of increasing the self-energization.
In a version with an incrementally adjustable ramp gradient, at least two ramp paths which have different gradients and are arranged at an angle are provided. In this context the rolling elements are in turn slide-mounted in the brake lining pressure plate 18.
The ramp gradient is adapted to the coefficient of friction of the brake lining by switching over the brake plunger 5, 6 to the better adapted ramp gradient after a previous braking process during which it was necessary to switch over.
The brake plunger 5, 6 is rotated by a separate drive or automatically, for example similarly to the way described above.
The switching-over process is triggered after the end of the braking, in which case the adjustment rotational movement which acts on the brake plungers via the gear mechanism is elastically stored in a transmission element and is not implemented until the brake is released owing to the block on rotation of the brake plunger which then decreases again.
The block on rotation can be produced by frictional forces which act on the spindle as a result of the braking force or as a result of holding forces which are exerted by the electrical drive motor or an engaged clutch, for example an electromagnetic clutch, to the brake plunger itself or an element of the projection device or preferably by the balls or rolling elements which are located outside the center of the brake plunger in braking processes in a ramp path, and generates a holding torque using the braking force transmitted by the brake plunger, the positively locking accommodation in the ramp path (track) and the position which is eccentric to the center of the brake plunger.
The tracks for the ramp paths are expediently embodied in such a form that the track depth is low in the region of small brake application forces, i.e. low eccentricity of the ball or of the rolling element, and a large track depth is implemented toward the outer diameter of the brake plunger in order to achieve a high load-bearing capacity.
With this solution it is possible for direct switching over during the braking process also to occur in the region of low braking forces. Only when relatively high braking forces are present will the ball or the rolling element assume a position in the ramp track in which it is no longer possible to switch over during the braking process.
A crank drive is preferably used to drive the brake lining unit. As an alternative to a crank drive, other brake application elements such as an eccentric arrangement and the like are also envisioned if they bring about a nonlinear movement of the brake lining unit in the circumferential direction.
The electromagnetic brake is controlled in each case by a computer unit on the brake, which computer units may be networked or, for example, by using a superordinate computer on the vehicle for one or more brakes.
A linear drive with a largely analogous arrangement is alternatively also envisioned. Instead of the crank tappet, a gearwheel segment which engages in a toothed rack on the brake lining back (not illustrated here) is fitted onto the drive shaft here.
However, the nonlinear drive is preferably used.
FIG. 5 shows a further variant of the self-energizing disc brake according to the invention which corresponds largely to the exemplary embodiment in FIG. 1.
As in the exemplary embodiment in FIG. 1, the activation device or brake application unit is composed of the two adjustment pistons or brake plungers 5, 6 which have variable lengths for the purpose of adjustment for wear and which, in their pressure surface facing the brake disc 2, have the recesses 14 in the manner of ramp contours on which the rolling elements 16 run, the rolling elements 16 transmitting the brake application force generated by the brake to the brake lining unit or to the pressure plate resting on the brake lining.
In addition, an engageable clutch, here for example a magnetic clutch 46, in particular a clutch with bistable-action actuating magnets, is provided and is designed to shift the crank 34 in and out of the drive train between the actuator (electric motor) and brake lining unit, for example on an axially movable radial toothing 48. In this way it is possible, for example, to firstly brake in a selective way for parking braking operations or even exclusively only using the brake plungers 5, 6 or else it is possible, for example, for relatively small adaptation braking operations to be carried out solely by rotating the brake plungers 5, 6 or by changing the axial length of the brake plungers. If, on the other hand, a “normal” service braking process is initiated, the clutch is switched over and the braking process is carried out by the crank 34.
In addition, according to FIG. 5 a switching device 47 is provided for rotating the brake plungers, here the nuts, from one raceway 14a to the other raceway 14b. This switching device 47 can be configured as a separate electric motor or else as a switching magnet or the like which rotates one of the nuts 13 through 90° using, for example, a toothed rack or the like, in which case the other of the nuts 13 is entrained, for example by a crown gear 45.
It is also to be noted that the present brake designs can also be considered to be particularly advantageous in terms of their control behavior.
If, for example, a normal force closed-loop control process is carried out which is considered not to be usable according to the prior art as the only closed-loop control, it has an advantageous effect that this normal force can be determined very precisely by, for example, supporting the brake plungers on the brake caliper (parallel force to the longitudinal axis of the brake plungers) by, for example, arranging corresponding sensors on the brake plungers and/or adjacent elements.
The following list of reference symbols is provided to simplify understanding of the foregoing specification and drawing
















Brake caliper
1



Brake disc
2



Openings
3, 4



Brake plunger
5, 6



Ball bearings
7, 8



Plain bearing shells
9, 10



Rear wall
11



Spindle
12



Nut
13



Ramp-like recess
14



Rolling element
16



Plain bearing shells
17



Recesses
15



Pressure plate
18



Carrier plate
19



Brake lining
20



Brake lining material
21



Clamping spring
22



Drive motor
23



Step-down gear mechanism
24



Output shaft
25



Gear mechanism
26



Sun gear
27



Planetary gears
28



Ring
29



Gearwheels
30



Flange
31



Pressure springs
32



Planet star
33



Crank
34



Crank tappet
35



Opening
36



Gear mechanism
37



Ring
38



Adjustment actuator
39



Output shaft
40



Gearwheel
41



Belt drive
42



Planetary gears
43



Planet star
44



Crown gear
45



Magnetic clutch
46



Switching device
47



Ramp angle
α



Longitudinal axis
LA



Circumferential direction
U



Axis of brake disc
BA

## Claims

1. A self-energizing disk brake comprising:
a brake application unit for applying the brake lining to one side of a brake disk with an application movement of the brake lining relative to the brake disk, the application movement having a component extending parallel to a rotational axis of the brake disk and another component extending parallel to a friction surface of the brake disk in a circumferential direction relative to the rotational axis of the disk brake, the brake application unit including a self-energizing device;
an electric actuator in which an activation force applied thereto is amplified using the self-energizing device, the self-energizing device being arranged between the electric actuator and a brake lining;
at least one electromotive drive driven by the electric actuator for activating the brake application unit; and
at least two brake plungers of the brake application unit;
wherein
the self-energizing device includes pressure surfaces of the at least two brake plungers provided, on sides facing the brake lining, with a recess having a ramp shaped contour for engagingly receiving a rolling element, the rolling element being supported at one end by the ramp shaped contour and at another end by the brake lining, and
the electric actuator rotates a crank one of directly and through at least one gear mechanism, the crank being arranged to rotate about an axis parallel to the rotational axis of the brake disk to move the brake lining parallel to the friction surface of the brake disk in the circumferential direction relative to the rotational axis of the brake disk, wherein the crank engages a correspondingly aligned opening in the pressure plate.

2. The disk brake as claimed in claim 1, wherein the at least one electromotive drive activates the brake application unit to drive the at least two brake plungers to change an axial length of the at least two brake plungers.

3. The disk brake as claimed in claim 1, further comprising a further electromotive drive for driving the at least two brake plungers to change an axial length of the at least two brake plungers.

4. The brake disk as claimed in claim 1, wherein the at least two brake plungers are oriented parallel to the rotational axis of the brake disk and are supported at their ends by one of a brake caliper and a bearing device disposed on a component connected to the brake caliper, said bearing device permitting at least one of the brake plungers to rotate about its longitudinal axis.

5. The disk brake as claimed in claim 1, wherein the rolling elements each engage a corresponding recess formed in the brake lining unit.

6. The disk brake as claimed in claim 1, wherein the rolling elements each engage in a ramp contour formed in the brake lining unit.

7. The disk brake as claimed in claim 1, wherein the rolling elements comprise balls and the recesses are in the form of conical surfaces.

8. The disk brake as claimed in claim 1, wherein the rolling elements comprise balls and the recesses comprise raceways.

9. The disk brake as claimed in claim 1, wherein the rolling elements comprise rollers and the recesses comprise ramps.

10. The disk brake as claimed in claim 1, wherein the recesses comprise at least one of dual direction ramps and conical surfaces.

11. The disk brake as claimed in claim 1, wherein a lowest point of the ramp shaped contour is advantageously configured to latch the rolling elements therein, and to provide a minimum latching force to oppose release of the rolling elements from the ramp shaped contour.

12. The disk brake as claimed in claim 7, wherein the recesses comprising the conical surfaces have a constant angle of aperture with respect to a longitudinal axis of the brake plungers, in the circumferential direction around the longitudinal axis of the brake plungers.

13. The disk brake as claimed in claim 7, wherein the recesses comprising the conical surfaces have changing angles of aperture with respect to a longitudinal axis of the brake plungers, in the circumferential direction around the longitudinal axis of the brake plungers.

14. The disk brake as claimed in claim 7, wherein the recesses comprising the conical surfaces have one of an incrementally and continuously changing angle of aperture with respect to a longitudinal axis of the brake plunger, in the circumferential direction around the longitudinal axis of the brake plungers.

15. The disk brake as claimed in claim 1, further comprising plain bearing shells inserted into the recesses in the pressure plate of the brake lining unit.

16. The disk brake as claimed in claim 1, wherein the at least two brake plungers are each supported at ends facing away from the brake disk in the interior of the brake caliper by bearing devices comprising at least one of ball bearings, flat plain bearings and annular guiding bearings of the brake caliper.

17. The disk brake as claimed in claim 1, the at least two brake plungers each have a spindle with an external thread on which a nut with a corresponding internal thread is rotatably arranged.

18. The disk brake as claimed in claim 17, wherein the nut passes through one of an opening in the brake caliper and a closure plate on the brake caliper, and further comprise a flange disposed on a side facing away from the brake disk, and pressure springs arranged between the flange and the internal wall of the brake caliper.

19. The disk brake as claimed in claim 1, wherein the brake lining has a pressure plate bearing against a brake lining carrier plate of the brake lining.

20. The disk brake as claimed in claim 1, wherein the opening in the pressure plate is arranged centrally between the at least two brake plungers.

21. The disk brake as claimed in claim 1, wherein the opening in the pressure plate comprises a connecting link which extends perpendicularly to a plane through the at least two brake plungers.

22. The disk brake as claimed in claim 1, further comprising a device for automatically rotationally orienting the at least two brake plungers which are provided with a variable conical aperture angle in the recesses, for rotationally orienting the nuts as a function of a coefficient of friction.

23. The disk brake as claimed in claim 1, wherein the device is designed to orient the at least two brake plungers with a further gear mechanism.

24. The disk brake as claimed in claim 1, wherein the device is designed to automatically orient the brake plungers with a further drive with an electric motor.

25. The disk brake as claimed in claim 1, wherein the electromotive drive is coupled as an actuator to at least one of an open-loop and closed-loop control device configured to perform at least one of a corresponding open-loop and closed-loop control of a position of an activator element of the brake lining.

26. The disk brake as claimed in claim 1, further comprising an engageable clutch configured to switch the crank in and out of the drive train between the brake lining unit and the electromotor.

27. The disk brake as claimed in claim 1, wherein, during a parking braking operation, the at least two brake plungers exclusively apply a parking brake.