DE102021113291B4 - Method and device for determining at least one force acting on a drive unit and/or at least one drive torque acting on a drive unit - Google Patents

Abstract

Method for determining at least one force (3) acting on a drive assembly (2) and/or at least one drive torque (25) acting on the drive assembly (2), the drive assembly (2) being mounted on at least one vibration damping device (4) in a receiving device ( 5), wherein a deflection state and/or a time profile of the deflection state of the vibration damping device (4) is determined by directly measuring a distance (7) between test areas (8a, 8b), the test areas (8a, 8b) being arranged in this way are that the distance (7) is determined by the deflection state, and at least one acceleration (11) of the receiving device (5) is determined, characterized in that a parameter adaptation step is provided in order to determine an initial expansion of the vibration damping device (4).

Description

The invention relates to a method and a device for determining at least one force acting on a drive unit and/or at least one drive torque acting on a drive unit, the drive unit being mounted on at least one vibration damping device in a receiving device.

A generic method and a corresponding device are from DE 10 2009 023 903 A1 known.

The U.S. 3,800,599A describes a similar method for determining a force acting on a drive unit and a device for this purpose.

From the DE 100 52 248 C1 a system for mounting an oscillating mass, in particular for mounting an oscillating mass in a motor vehicle, is known. A sensor with a sensor element made of a piezoelectric film is provided, the output signal of which depends on periodic forces acting on a bearing and is used to control or regulate the bearing. Piezoelectric sensors usually have a zero point drift, making it difficult to detect a quasi-stationary bearing force and thus a drive torque.

From the DE 10 2018 209 825 A1 a warehouse for the collection of vehicle-specific data emerges. In this case, the bearing comprises a bearing body which is formed at least partially from a dielectric elastomer material.

The EP 1 424 225 A2 discloses a bearing for a control arm of a wheel suspension of a vehicle, at least one sensor arranged in or on the bearing determining the relative movement of the vehicle parts connected to one another by the bearing.

From the EP 2 218 934 A1 discloses a vibration damping device and a method for evaluating the performance of the vibration damping device having a main body made of an anti-vibration material.

Methods and devices are therefore known from the prior art which can be used to determine forces to which, for example, a drive unit of a motor vehicle is exposed.

It is known from the general state of the art that drive units in motor vehicles are mounted in a vibration-damped manner. Vibration damping of this type protects the drive assembly from transmission of external shocks and also reduces transmission of vibrations from the drive assembly to the vehicle.

A disadvantage of the methods and devices known from the prior art is that they can only unreliably determine the forces acting on the drive assembly. However, an unreliable determination of the forces acting on the drive unit can have negative effects on the determination of parameters relevant to driving safety, such as a drive torque of the drive unit.

In this context, the ISO standard 26262 describes the requirements for electrical or electronic, in particular programmable electronic systems in motor vehicles. With electrification and increasing drive-by-wire technology, there is an increased risk that errors in the control will lead to the undesirable implementation of variables, in particular drive torques, in the drive system. This can cause serious accidents. The erroneous implementation of a variable in the drive train should be detected as quickly and reliably as possible. Furthermore, a safety-related controller should take countermeasures to ensure that the driver can control the vehicle.

The present invention is based on the object of creating a method and a device for determining at least one force acting on a drive unit and/or at least one drive torque acting on a drive unit, which avoid the disadvantages of the prior art and in particular are able a reliable compared to the prior art to enable a more precise determination of the force acting on the drive assembly or of the drive torque acting on the drive assembly.

According to the invention, this object is achieved by the features mentioned in claim 1.

In the method according to the invention for determining at least one force acting on a drive assembly and/or at least one drive torque acting on the drive assembly, the drive assembly is mounted on at least one vibration damping device in a receiving device. A deflection state and/or a time profile of the deflection state of the vibration damping device is determined by directly measuring a distance between test areas. The test areas are arranged in such a way that the distance is determined by the deflection state. At least one acceleration of the receiving device is also determined.

The method according to the invention has the advantage that, in order to determine the force acting on the drive unit, in particular a driving force reacting on the drive unit in the case of a linearly acting drive unit, and/or the drive torque acting on the drive unit, in addition to the deflection and/or the course of the deflection over time of the vibration damping device, in particular the resulting spring force, an acceleration acting synchronously on the receiving device is also determined and taken into account. By taking into account the acceleration of the receiving device, the force and/or the drive torque acting on the drive unit can also be determined with greater precision. On the other hand, ignoring the acceleration of the receiving device leads to an erroneous determination of the force and/or drive torque.

The method according to the invention therefore enables the determination of forces and/or drive torques in drive systems and can be used, for example, in the field of motor vehicle technology and general drive technology, particularly in connection with functional safety. As a result, the method described here contributes to meeting the requirements set out in the ISO standard 26262. The solution according to the invention can also be used as part of condition monitoring for, for example, industrial or marine gears, in order to detect, for example, drive torques, imbalances or incorrectly aligned clutches.

In an advantageous development of the method according to the invention, it can be provided that an influence of the gravitational acceleration is taken into account when determining the at least one acceleration. It is particularly advantageous to take account of the effect of gravitational acceleration on the recording device, since this represents an influencing variable which is sufficiently determined, for example, by an inclination of the drive unit and the recording device with respect to the gravitational field. It is therefore subject to fluctuations which can be linked to a position of the vehicle and can therefore be determined.

Furthermore, in the method according to the invention, a parameter adaptation step is provided in order to determine an initial expansion of the vibration damping device. It is particularly advantageous to determine the initial expansion or an unstressed spring length of the vibration damping device, since this represents a further central parameter of the vibration damping device in addition to the already mentioned parameters of the spring constant and the damping constant, which is determined in a step provided for this purpose.

In a further advantageous embodiment of the method according to the invention, it can be provided that at least one parameter of the vibration damping device is identified by means of a dynamic model for a movement of the drive assembly. Parameters of the vibration damping device, such as a spring constant and/or a damping constant, can be subject to variability over time. For a particularly reliable determination of the force acting on the drive unit and/or the drive torque acting back on the drive unit, it is advantageous if the parameters of the vibration damping device are not only assumed on the basis of specification values, for example, but rather as part of the method, in particular regularly be determined. As a result, the reaction forces exerted by the vibration damping device on the drive unit can be determined particularly reliably. In this case, the dynamic model enables the extraction of the at least one parameter from time profiles of measurement data of the deflection state. Such a data record can be recorded in particular during normal operation of the drive unit.

An advantageous development of the method according to the invention can consist in determining the drive torque of the drive unit based on the determined deflection state and/or the time profile of the deflection state of the vibration damping device and/or the determined acceleration and/or the identified parameters. In order to determine an operating state of the drive unit, it is particularly advantageous to determine the drive torque exerted by the drive unit. In particular, it is advantageous if the determination of the drive torque is carried out particularly reliably using the determined reaction forces on the basis of the deflection state of the vibration damping device, the determined acceleration and the identified parameters.

If, in a further embodiment of the method according to the invention, the dynamic model maps at least two degrees of freedom of the movement of the drive assembly in the vibration damping device, one of the two degrees of freedom can be used to check the plausibility of the other. This results in an even more precise determination of the force acting on the drive assembly and/or of the drive torque acting back on the drive assembly

In an advantageous development of the method according to the invention, the vibration damping device can be formed by an elastomer bearing and the distance can be measured by means of a dielectric elastomer sensor. The use of elastomeric bearings for vibration damping has the advantage that they can be produced inexpensively and at the same time have all the desired properties of a vibration damping device. Using a dielectric elastomer sensor also enables an advantageously simple and direct measurement of the distance between the test areas. It is of particular advantage here if the test areas also represent the edge areas of the elastomer sensor and are designed, for example, as layers of the elastomer bearing. As a result, a deformation of the elastomer bearing and thus the deflection state of the anti-vibration device can be determined directly.

Well-known dielectric elastomer sensors contain at least two thin electrodes between which an elastomer layer is applied as a dielectric. An applied force leads to a deformation of the elastomer layer in one or more spatial directions, which leads to a change in the electrical capacitance that can be measured between the electrodes. A deformation state and thus the deflection state of the elastomer bearing are thus determined. An alternative distance measurement or measurement of the spring deflection also provides information about the state of deformation of the elastomer bearing.

It is known that elastomeric materials may creep under load, i.e. the state of deformation changes over time under constant load. This leads to a change in the initial expansion or the unstressed spring length of an elastomeric bearing. In addition, there is oxygen diffusion into the interior of the elastomeric material. Oxidation reactions cause aging-related changes in stiffness, which take place differently in the interior of the elastomer material than near the surface due to the lower oxygen concentration. Thus, the initial expansion or the unstressed spring length as well as the spring stiffness of an elastomeric bearing change over time. Different levels of humidity and temperature also affect the properties of the elastomer material. These parameter changes can be taken into account when determining the bearing forces from the state of deformation.

In a particularly preferred embodiment, the deformation state of the at least one elastomer bearing of the drive unit is determined with the aid of a direct distance measurement using a dielectric elastomer sensor. With at least one acceleration, with an acceleration signal of the acceleration preferably including the influence of gravitational acceleration and with the help of a dynamic model for the movement of the drive unit in the elastomer bearings, the parameters of the at least one elastomer bearing are identified in order to detect creep or settling phenomena and changes in the spring stiffness or the spring constant and/or the damping or the damping constant. Based on the currently determined state of deformation, the current acceleration and the identified parameters, the drive torque is then determined, also with the help of a dynamic model.

In a further embodiment of the invention, the dynamic model maps at least two degrees of freedom of the movement of the drive unit in the elastomeric bearings. This creates at least two equations, which can be used to determine the drive torque and, based on an error in the equation, a plausibility between the currently determined state of deformation, the current len acceleration and the identified parameters. Sensor signals describing the current state of deformation and the current acceleration can preferably each be filtered with identical, linear low-pass filters.

Claim 7 specifies a device for determining at least one force acting on a drive assembly and/or at least one drive torque acting on a drive assembly.

In the device according to the invention for determining at least one force acting on a drive assembly and/or at least one drive torque acting on a drive assembly, the drive assembly is mounted on at least one vibration damping device in a receiving device. The device has at least one distance measuring device, which is set up for the direct measurement of a distance and/or a time profile of the distance between test areas. The test areas are arranged in such a way that the distance is determined, preferably at all times, by a deflection state of the vibration damping device. Furthermore, the device according to the invention has a computing device that is set up to determine the deflection state from the distance. Furthermore, the device according to the invention has at least one acceleration measuring device, which is set up to determine an acceleration of the recording device.

The device according to the invention enables an advantageous exact and precise determination of the force acting on the drive unit and/or the drive torque acting on the drive unit, in that the deflection state of the vibration damping device and thus its reaction force is advantageously determined quickly and precisely by means of a direct measurement. For this purpose, the device has a computing device. In addition, in the device according to the invention, an acceleration of the receiving device can be taken into account by means of the acceleration measuring device when determining the force acting on the drive unit and/or the drive torque acting on the drive unit. In combination, this allows a particularly precise determination of the acting force and/or the acting drive torque and thus, for example, a particularly reliable control of an operating state of the drive unit.

According to the invention, the vibration damping device has at least one elastomer bearing and the distance measuring device is designed as a dielectric elastomer sensor, resulting in a sensor elastomer bearing. Such elastomer bearings for vibration damping are cost-effective and at the same time have all the desired properties of a vibration damping device. An equally simple and direct measurement of the distance between the test areas can be carried out using a dielectric elastomer sensor. If the test areas also represent the edge areas of the elastomer sensor and are designed, for example, as layers of the elastomer bearing, deformation of the elastomer bearing and thus the deflection state of the anti-vibration device can be determined directly.

In an advantageous development of the device according to the invention, it can be provided that the computing device is set up to identify at least one parameter of the vibration damping device using a dynamic model for a movement of the drive unit. Using the dynamic model, the at least one parameter can be extracted from measurement data of the deflection state over time, it being possible for a corresponding data set to be recorded, for example during operation of the drive assembly.

In a further advantageous embodiment of the device according to the invention, it can be provided that the computing device is set up to calculate the drive torque of the to determine the drive unit. The drive torque of the drive unit can be determined particularly reliably with a computing device designed in this way.

Furthermore, at least two vibration damping devices can be provided and the acceleration measuring device can have at least two acceleration sensors, the at least two acceleration sensors being set up to detect a vertical acceleration of the recording device at a position of the at least two vibration damping devices. The use of at least two vibration damping devices and at least two acceleration sensors enables the setting up and in particular the acquisition of data from at least two movement models of the drive unit or one movement model of the drive unit with at least two degrees of freedom. Due to the mutual dependencies of the movement models mentioned, they can be used to check their mutual plausibility, which in turn contributes to increasing the precision of the force and/or torque determination.

In a further advantageous embodiment of the device according to the invention, it can be provided that the dielectric elastomer sensor has at least two electrodes, preferably thin compared to the elastomer bearing, with an elastomer layer being arranged as a dielectric between the electrodes. The design of the dielectric elastomer sensor as part of the elastomer bearing has the advantage that the deformation of the elastomer bearing can be measured quickly, precisely and directly as a result.

The invention also relates to a motor vehicle with a drive unit having the features mentioned in claim 12.

The motor vehicle according to the invention has a drive unit which is mounted on at least one vibration damping device in a receiving device. A device for determining at least one force acting on the drive unit and/or at least one drive torque acting on the drive unit according to one of claims 8 to 13 is provided. Additionally or alternatively, in the motor vehicle according to the invention, at least one force acting on the drive unit and/or at least one drive torque acting on the drive unit is determined using a method according to one of claims 1 to 7.

The motor vehicle according to the invention has the advantage that an operating state of the drive unit can be determined using the particularly precise values that are detected using the method according to the invention and/or the device according to the invention. This leads to a very high level of operational reliability and/or driving safety of the motor vehicle according to the invention.

In the following, exemplary embodiments of the invention are shown in principle with reference to the drawing.

• 1 eine schematische Darstellung einer erfindungsgemäßen Vorrichtung;
• 2 eine schematische Darstellung einer weiteren Ausführungsform der erfindungsgemäßen Vorrichtung in einem Kraftfahrzeug;
• 3 eine vereinfachte diagrammmäßige Darstellung eines Anfahrvorganges; und
• 4 eine prinzipmäßige Darstellung eines Blockschaltbilds eines Tiefpassfilters.

It shows:

• 1 a schematic representation of a device according to the invention;
• 2 a schematic representation of a further embodiment of the device according to the invention in a motor vehicle;
• 3 a simplified diagrammatic representation of a starting process; and
• 4 a basic representation of a block diagram of a low-pass filter.

1 shows a schematic representation of a device 1 for determining at least one force 3 acting on a drive unit 2 . The drive unit 2 is mounted on at least one vibration damping device 4 in a receiving device 5 .

The device 1 has at least one distance measuring device 6, which is set up for the direct measurement of a distance 7 between test areas 8a, 8b. The test areas 8a, 8b are arranged in such a way that the distance 7 and/or a time profile of the distance 7 is determined by a deflection state of the vibration damping device 4. Device 1 also has a computing device 9, which is set up to determine the deflection state from distance 7, and an acceleration measuring device 10, which is set up to determine an acceleration of recording device 5, which is identified by the arrow labeled with reference number 11 .

In the 1 The device 1 shown is particularly suitable for carrying out a method for determining at least one force 3 acting on the drive assembly 2 mounted on the vibration damping device 4 in the receiving device 5. In the method, a deflection state and/or a time profile of the deflection state of the vibration damping device 4 determined by a direct measurement of the distance 7 between the test areas 8a, 8b. The test areas 8a, 8b are arranged in such a way that the distance 7 is determined by the deflection state. Furthermore, at least one acceleration of the recording device 5 identified by the arrow 11 is determined.

2 shows a schematic representation of a further embodiment of the device 1 according to the invention in a motor vehicle 12, in which the drive unit 2 is mounted on three vibration damping devices 4 in the receiving device 5 on a body. Furthermore, the motor vehicle 12 has the device 1 for determining the at least one drive torque 25 acting on the drive unit 2 . Furthermore, at the in 2 motor vehicle shown, which determines at least one drive torque 25 acting on the drive unit 2 by means of the aforementioned method.

• - Es werden keine Kurvenfahrt und kein Wanken betrachtet. Das Antriebsaggregat 2 und das Kraftfahrzeug 12 besitzen jeweils nur drei Freiheitsgrade, d.h. es handelt sich um ein ebenes Problem:
  • - Translatorische Bewegung in Richtung einer x-Achse 14 parallel zur Fahrbahnoberfläche.
  • - Translatorische Bewegung in Richtung einer z-Achse 15 senkrecht zur Fahrbahnoberfläche.
  • - Drehung (Nicken) um eine in der 2 als Kreis mit innenliegendem Punkt dargestellten y-Achse 16 (Nickachse, senkrecht zur Fahrbahnoberfläche).
• - Die in dem Ausführungsbeispiel nichtlinear ausgebildeten Elastomerlager 4a, 4b, 4c werden um einen Arbeitspunkt als linearisiert betrachtet, d.h. mittels linearen Federsteifigkeiten und linearen Dämpfungskonstanten beschrieben.
• - Verkopplung zwischen Einfederungen eines der Elastomerlager 4a, 4b, 4c und Lagerkräften in verschiedenen Raumrichtungen werden nicht betrachtet.
• - Ein Schwerpunkt 17 des Antriebsaggregats 2 liegt auf der Höhe von aggregateseitigen Anschlagpunkten 18 der Elastomerlager 4a, 4b, 4c, ein Abstand l des Schwerpunktes 17 von den aggregateseitigen Anschlagpunkten 18 ist auf beiden Seiten gleich groß.
• - Trägheitswirkungen von rotierenden, drehbeschleunigten Teilen in dem Antriebsaggregat 2 sind in dem Ausführungsbeispiel nicht berücksichtigt. Eine Berücksichtigung kann beispielsweise mittels einer Messung einer Winkelbeschleunigungen der drehbeschleunigten Teile und durch zusätzliche Terme in den nachfolgend angegebenen Gleichungen einfach realisiert werden.

In the present case, motor vehicle 12 is an electric vehicle. The drive assembly 2 has an electric machine and a transmission in a manner not shown in detail and is attached to a body 13 of the motor vehicle 12 with the aid of the vibration damping device 4, which in the present case has several elastomeric bearings 4a, 4b, 4c. For the sake of simplicity, some simplifications described below are used:

• - No cornering and no rolling are considered. The drive unit 2 and the motor vehicle 12 each have only three degrees of freedom, ie it is a plane problem:
  • - Translational movement in the direction of an x-axis 14 parallel to the road surface.
  • - Translational movement in the direction of a z-axis 15 perpendicular to the road surface.
  • - Rotation (pitch) by one in the 2 y-axis 16 (pitch axis, perpendicular to the road surface) shown as a circle with a point inside.
• The elastomeric bearings 4a, 4b, 4c, which are nonlinear in the exemplary embodiment, are considered to be linearized around an operating point, ie are described by means of linear spring stiffness and linear damping constants.
• - Coupling between deflection of one of the elastomer bearings 4a, 4b, 4c and bearing forces in different spatial directions are not considered.
• - A center of gravity 17 of the drive unit 2 is at the level of unit-side attachment points 18 of the elastomeric bearings 4a, 4b, 4c, a distance l of the center of gravity 17 from the unit-side attachment points 18 is the same on both sides.
• - Effects of inertia of rotating, rotationally accelerated parts in the drive unit 2 are not taken into account in the exemplary embodiment. It can easily be taken into account, for example, by measuring an angular acceleration of the rotationally accelerated parts and by using additional terms in the equations given below.

in the in 2 In the exemplary embodiment shown, the drive unit 2 is supported on the body 13 via three elastomeric bearings 4a, 4b, 4c. The rear vertical elastomer bearing 4a consists of a vertical spring with spring stiffness c _h and a vertical damper with damping constant d _h . The front vertical elastomer bearing 4b consists of a vertical spring with spring stiffness c _v and a vertical damper with damping constant d _v . The rear horizontal elastomer bearing 4c consists of a horizontal spring with spring stiffness c _x and a horizontal damper with damping constant d _x .

The variables z _h and z _Lh describe the heights 20 and 19 of the upper and lower stop points of the rear vertical elastomer bearing 4a in the inertial system, for example above an ideal, level road surface. The variables z _v and z _Lv correspondingly describe the heights 21 and 22 of the upper and lower stop points of the front vertical elastomer bearing 4b. The variable z _0h describes the unstressed length of the rear vertical elastomeric bearing 4a or the rear vertical spring c _h and the variable z _ov describes the unstressed length of the front vertical elastomeric bearing 4b or the front vertical spring c _v .

die Lagerkraft des vorderen vertikalen Elastomerlagers F_zv ergibt sich zu $$F_{zv}=c_v\left(z_{Lv}+z_{0v}-z_v\right)+d_v\left({\dot{z}}_{Lv}-{\dot{z}}_v\right).$$

The bearing force of the rear vertical elastomer bearing F _zh thus results in $$f_{\mathrm{e.g}H}=c_H\left({\mathrm{e.g}}_{LH}+{\mathrm{e.g}}_{0H}-{\mathrm{e.g}}_H\right)+{\mathrm{i.e}}_n\left({\dot{\mathrm{e.g}}}_{LH}-{\dot{\mathrm{e.g}}}_H\right),$$

the bearing force of the front vertical elastomer bearing F _zv results in $$f_{\mathrm{e.g}v}=c_v\left({\mathrm{e.g}}_{Lv}+{\mathrm{e.g}}_{0v}-{\mathrm{e.g}}_v\right)+{\mathrm{i.e}}_v\left({\dot{\mathrm{e.g}}}_{Lv}-{\dot{\mathrm{e.g}}}_v\right).$$

The drive unit 2 has the known mass m and the known mass moment of inertia J _y about the y-axis 16 (pitch axis, perpendicular to the road surface). Roadway 23 has gradient angle 24 with value α _St . A drive torque 25 of the drive assembly 2 with the value M _A acts back on the drive assembly 2 via the drive shafts and is supported via the elastomer bearings 4a, 4b. The drive torque 25 of the drive unit 2 can be determined based on the determined deflection state of the vibration damping device 4 and/or the acceleration 11 determined by means of the acceleration measuring device 10 and/or the identified parameters.

The dynamics in the z-direction 15 and the pitching are described below:

$$\begin{array}{l}m\ddot{z}=\frac{m}{2}\left({\ddot{z}}_h+{\ddot{z}}_v\right)\\ =c_h\left(z_{Lh}+z_{0h}-z_h\right)+d_h\left({\dot{z}}_{Lh}-{\dot{z}}_h\right)+c_v\left(z_{Lv}+z_{0v}-z_v\right)+\\ d_v\left({\dot{z}}_{Lv}-{\dot{z}}_v\right)-mg\text{ cos}\left({\alpha }_{St}\right)\end{array}$$

For example, as a result of road bumps 26, the center of gravity 17 of the drive unit 2 can move in the vertical direction, the following applies: $$\mathrm{e.g}=\frac{1}{2}\left({\mathrm{e.g}}_H+{\mathrm{e.g}}_v\right)$$

$$\begin{array}{l}m\ddot{\mathrm{e.g}}=\frac{m}{2}\left({\ddot{\mathrm{e.g}}}_H+{\ddot{\mathrm{e.g}}}_v\right)\\ =c_H\left({\mathrm{e.g}}_{LH}+{\mathrm{e.g}}_{0H}-{\mathrm{e.g}}_H\right)+{\mathrm{i.e}}_H\left({\dot{\mathrm{e.g}}}_{LH}-{\dot{\mathrm{e.g}}}_H\right)+c_v\left({\mathrm{e.g}}_{Lv}+{\mathrm{e.g}}_{0v}-{\mathrm{e.g}}_v\right)+\\ {\mathrm{i.e}}_v\left({\dot{\mathrm{e.g}}}_{Lv}-{\dot{\mathrm{e.g}}}_v\right)-mG\text{cos}\left(a_{St}\right)\end{array}$$

$$\begin{array}{l}{J}_y\ddot{\vartheta }=\frac{J_y}{2l}\left({\ddot{z}}_h-{\ddot{z}}_v\right)\\ =c_{h}l\left(z_{Lh}+z_{0h}-z_h\right)+d_{h}l\left({\dot{z}}_{Lh}-{\dot{z}}_h\right)-c_{v}l\left(z_{Lv}+z_{0v}-z_v\right)-\\ d_{v}l\left({\dot{z}}_{Lv}-{\dot{z}}_v\right)-M_A\end{array}$$

The drive torque or drive torque M _A leads to a twisting angle 28 of the drive unit 2 with the value _θ about the y-axis 16 (pitching). It applies approximately: $$\vartheta =\frac{1}{2l}\left({\mathrm{e.g}}_H-{\mathrm{e.g}}_V\right)$$

$$\begin{array}{l}{J}_y\ddot{\vartheta }=\frac{J_y}{2l}\left({\ddot{\mathrm{e.g}}}_H-{\ddot{\mathrm{e.g}}}_v\right)\\ =c_{H}l\left({\mathrm{e.g}}_{LH}+{\mathrm{e.g}}_{0H}-{\mathrm{e.g}}_H\right)+{\mathrm{i.e}}_{H}l\left({\dot{\mathrm{e.g}}}_{LH}-{\dot{\mathrm{e.g}}}_H\right)-c_{v}l\left({\mathrm{e.g}}_{Lv}+{\mathrm{e.g}}_{0v}-{\mathrm{e.g}}_v\right)-\\ {\mathrm{i.e}}_{v}l\left({\dot{\mathrm{e.g}}}_{Lv}-{\dot{\mathrm{e.g}}}_v\right)-M_A\end{array}$$

$$\Delta z_v=z_v-z_{Lv}$$

The exemplary embodiment described above is based on elastomeric bearings 4a, 4b, 4c with at least two thin electrodes and an elastomer layer in between as the dielectric. A deformation of the elastomer layer leads to a change in the electrical capacitance that can be measured between the electrodes. A corresponding sensor system described below can be implemented inexpensively. The distance measuring device 6 is designed as a dielectric elastomer sensor. The state of deformation of the elastomer layer or, by means of suitable scaling, the state of deformation of the entire elastomer bearing can thus be determined by means of suitable evaluation. In this way, the distance 7 can be measured by means of the dielectric elastomer sensor. It is assumed that sensor signals Δz _h and Δz _v are available, which characterize the distances between the contact points of the elastomeric bearings 4a, 4b on the drive unit 2 and the body 13: $$\Delta {\mathrm{e.g}}_H={\mathrm{e.g}}_H-{\mathrm{e.g}}_{LH}$$

$$\Delta {\mathrm{e.g}}_v={\mathrm{e.g}}_v-{\mathrm{e.g}}_{Lv}$$

$$S_v={\ddot{z}}_{Lv}+g\text{ cos}\left({\alpha }_{St}\right)$$

in the in 2 The exemplary embodiment illustrated also includes two acceleration measuring devices 10 designed as acceleration sensors, which record the acceleration of the body 13 in the z-direction 15 at the body-side stop points of the front and rear vertical elastomer bearings 4b and 4a. In this way, vertical movements and pitching movements of the body 13 are detected. Corresponding sensors are typically based on the principle of seismic mass, i.e. they also react to the gravitational acceleration g. As a result, an influence of the acceleration due to gravity can be taken into account when determining the at least one acceleration 11 . There are thus sensor signals S _h and S _v of the acceleration sensors: $$S_H={\ddot{\mathrm{e.g}}}_{LH}+G\text{cos}\left(a_{St}\right)$$

$$S_v={\ddot{\mathrm{e.g}}}_{Lv}+G\text{cos}\left(a_{St}\right)$$

Equations (7) and (8) result from a division of the gravitational acceleration g into a slope acceleration 27a parallel to the x-axis 14 and an acceleration 27b parallel to the z-axis 15.

As an alternative to this, the sensor signals S _h and S _v can also be derived from the sensors of an ESP system of the motor vehicle 12 . Then no additional acceleration sensors are required, which minimizes the costs. In modern ESP sensor concepts, the acceleration of the body 13 in the x, y, and z directions 14, 16, 15 as well as rotation rates about the x, y, and z axes are determined. Provision can be made for the accelerations at the body-side stop points of the front and rear vertical elastomer bearings 4b, 4a to be calculated using these variables measured at an installation location of the ESP sensor system.

Preferably, the following in connection with 4 described low-pass filter according to Eq. (18) used. The rotational accelerations can thus be determined in a simple manner from the rates of rotation.

multiplizierten Gl. (7) und (8) von der Gl. (2), so ergibt sich: $$\begin{array}{l}\frac{m}{2}\left({\ddot{z}}_h-{\ddot{z}}_{Lh}\right)+\frac{m}{2}\left({\ddot{z}}_v-{\ddot{z}}_{Lv}\right)-\frac{2m}{2}g\text{ cos}\left({\alpha }_{St}\right)\\ =c_h\left(z_{Lh}+z_{0h}-z_h\right)+d_h\left({\dot{z}}_{Lh}-{\dot{z}}_h\right)+c_v\left(z_{Lv}+z_{0v}-z_v\right)+d_v\left({\dot{z}}_{Lv}-{\dot{z}}_v\right)-\\ mg\text{ cos}\left({\alpha }_{St}\right)-\frac{m}{2}{S}_h-\frac{m}{2}{S}_v\end{array}$$

If you subtract the with the factor $\frac{m}{}$$\frac{}{2}$

multiplied Eq. (7) and (8) from Eq. (2), this results in: $$\begin{array}{l}\frac{m}{2}\left({\ddot{\mathrm{e.g}}}_H-{\ddot{\mathrm{e.g}}}_{LH}\right)+\frac{m}{2}\left({\ddot{\mathrm{e.g}}}_v-{\ddot{\mathrm{e.g}}}_{Lv}\right)-\frac{2m}{2}G\text{cos}\left(a_{St}\right)\\ =c_H\left({\mathrm{e.g}}_{LH}+{\mathrm{e.g}}_{0H}-{\mathrm{e.g}}_H\right)+{\mathrm{i.e}}_H\left({\dot{\mathrm{e.g}}}_{LH}-{\dot{\mathrm{e.g}}}_H\right)+c_v\left({\mathrm{e.g}}_{Lv}+{\mathrm{e.g}}_{0v}-{\mathrm{e.g}}_v\right)+{\mathrm{i.e}}_v\left({\dot{\mathrm{e.g}}}_{Lv}-{\dot{\mathrm{e.g}}}_v\right)-\\ mG\text{cos}\left(a_{St}\right)-\frac{m}{2}{S}_H-\frac{m}{2}{S}_v\end{array}$$

After inserting Eq. (5) and Eq. (6) in Eq. (9) one obtains: $$\begin{array}{l}\frac{\mathrm{<figure-callout\; id="2"\; label="m"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">m</figure-callout>}}{2}\Delta \ddot{}{\mathrm{e.g}}_H+\frac{\mathrm{<figure-callout\; id="2"\; label="m"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">m</figure-callout>}}{2}\Delta \ddot{}{\mathrm{e.g}}_v+c_H\Delta {\mathrm{e.g}}_H+{\mathrm{i.e}}_H\Delta \dot{}{\mathrm{e.g}}_H+c_v\Delta {\mathrm{e.g}}_v+{\mathrm{i.e}}_v\Delta \dot{}{\mathrm{e.g}}_v=c_H{\mathrm{e.g}}_{0H}+c_v{\mathrm{e.g}}_{0v}-\\ \frac{m}{2}{S}_H-\frac{m}{2}{S}_v\end{array}$$

multiplizierten Gl. (7) von der Gl. (4) und addiert die die mit dem Faktor $\frac{J_y}{2l}$

multiplizierten Gl. (8), so ergibt sich: $$\begin{array}{l}\frac{J_y}{2l}\left({\ddot{z}}_h-{\ddot{z}}_{Lh}\right)-\frac{J_y}{2l}\left({\ddot{z}}_v-{\ddot{z}}_{Lv}\right)\\ =c_{h}l\left(z_{Lh}+z_{0h}-z_h\right)+d_{h}l\left({\dot{z}}_{Lh}-{\dot{z}}_h\right)-c_{v}l\left(z_{Lv}+z_{0v}-z_v\right)-\\ d_{v}l\left({\dot{z}}_{Lv}-{\dot{z}}_v\right)-M_A-\frac{J_y}{2l}{S}_h+\frac{J_y}{2l}{S}_v\end{array}$$

Similarly, subtracting those with the factor $\frac{J_y}{2l}$

multiplied Eq. (7) from Eq. (4) and adds the with the factor $\frac{J_y}{2l}$

multiplied Eq. (8), this results in: $$\begin{array}{l}\frac{J_y}{2l}\left({\ddot{\mathrm{e.g}}}_H-{\ddot{\mathrm{e.g}}}_{LH}\right)-\frac{J_y}{2l}\left({\ddot{\mathrm{e.g}}}_v-{\ddot{\mathrm{e.g}}}_{Lv}\right)\\ =c_{H}l\left({\mathrm{e.g}}_{LH}+{\mathrm{e.g}}_{0H}-{\mathrm{e.g}}_H\right)+{\mathrm{i.e}}_{H}l\left({\dot{\mathrm{e.g}}}_{LH}-{\dot{\mathrm{e.g}}}_H\right)-c_{v}l\left({\mathrm{e.g}}_{Lv}+{\mathrm{e.g}}_{0v}-{\mathrm{e.g}}_v\right)-\\ {\mathrm{i.e}}_{v}l\left({\dot{\mathrm{e.g}}}_{Lv}-{\dot{\mathrm{e.g}}}_v\right)-M_A-\frac{J_y}{2l}{S}_H+\frac{J_y}{2l}{S}_v\end{array}$$

After inserting Eq. (5) and Eq. (6) in Eq. (11) one obtains: $$\begin{array}{l}\frac{J_y}{2l}\Delta \ddot{}{\mathrm{e.g}}_H-\frac{J_y}{2l}\Delta \ddot{}{\mathrm{e.g}}_v+c_{H}l\Delta {\mathrm{e.g}}_H+{\mathrm{i.e}}_{H}l\Delta \dot{}{\mathrm{e.g}}_H-c_{v}l\Delta {\mathrm{e.g}}_v-{\mathrm{i.e}}_{v}l\Delta \dot{}{\mathrm{e.g}}_v=c_{H}l{\mathrm{e.g}}_{0H}-\\ c_{v}l{\mathrm{e.g}}_{0v}-M_A-\frac{J_y}{2l}{S}_H+\frac{J_y}{2l}{S}_v\end{array}$$

An advantage of the method described is that the gradient angle 24 with the value α _St is received via the sensor signals _Sh and S _v of the acceleration sensors. Eq. (10) and (12) apply both to a horizontal and to an uphill or downhill roadway 23. An additional sensor system for the gradient angle 24 with the value α _St is not required. In the same way, varying pitch angles of the body 13, for example as a result of different loading conditions of the motor vehicle 12, during acceleration and braking processes or bumps in the road 26, are entered via the sensor signals S _h and S _v of the acceleration sensors.

$\Delta \ddot{}{z}_h\text{ und }\Delta \ddot{}{z}_v$

beschreiben den zeitlichen Verlauf der Auslenkung der Schwingungsdämpfungseinrichtung 4. Zusammen mit den Sensorsignalen S_h und S_v der Beschleunigungsmesseinrichtung 10, welche auf die Aufnahmeeinrichtung 5 wirkende Beschleunigungen bestimmt, lässt sich entsprechend Gl. (12) das auf das Antriebsaggregat 2 rückwirkende Antriebsdrehmoment 25 ermitteln.Eq. (10) and (12) represent a dynamic model for the movement of the drive unit 2 in the vibration damping device 4. According to the two Eqs. (10) and (12), the described dynamic model maps two degrees of freedom of the movement of the drive assembly 2. The sensor signals Δz _h and Δz _v , which characterize the distances between the contact points of the elastomeric bearings 4a, 4b on the drive unit 2 and the body 13, and their derivatives over time $\Delta \dot{}{\mathrm{e.g}}_H,\Delta \dot{}{\mathrm{e.g}}_v,$

$\Delta \ddot{}{\mathrm{e.g}}_H\text{and}\Delta \ddot{}{\mathrm{e.g}}_v$

describe the time course of the deflection of the vibration damping device 4. Together with the sensor signals S _h and S _v of the acceleration measuring device 10, which determines accelerations acting on the recording device 5, Eq. (12) Determine the drive torque 25 acting on the drive assembly 2.

• Die ungespannte Länge z_0h des hinteren vertikalen Elastomerlagers 4a bzw. dessen Feder c_h und die ungespannte Länge z_0v des vorderen vertikalen Elastomerlagers 4b bzw. dessen Feder c_v ändern sich über der Zeit infolge von Setzerscheinungen der Elastomerlager 4a, 4b. Die Federsteifigkeiten c_h und c_v ändern sich ebenfalls über der Zeit, z.B. infolge von Oxidationsreaktionen im Elastomermaterial. Insofern ist in gewissen Zeitabständen eine Adaption dieser Größen erforderlich. Diese Parameteradaption kann am stehenden Kraftfahrzeug 12 durchgeführt werden. Bis auf den Anteil der Erdbeschleunigung sind sämtliche Beschleunigungen in Richtung der z-Achse 15 gleich Null, ebenso die Geschwindigkeiten und die Gl. (10) vereinfacht sich zu: $$c_h\Delta z_h+c_v\Delta z_v=c_h{z}_{0h}+c_v{z}_{0v}-\frac{m}{2}{S}_h-\frac{m}{2}{S}_v$$
  

Using the dynamic model for the movement of the drive unit 2 in the vibration damping device 4 according to Eqs. (10) and (12), at least one parameter of the vibration damping device 4 is identified. The parameter identification is described below using the example of a vehicle 12 designed as an electric vehicle:

• The unrestrained length z _0h of the rear vertical elastomeric bearing 4a or its spring c _h and the unrestrained length z _0v of the front vertical elastomeric bearing 4b or its spring c _v change over time due to the settling of the elastomeric bearings 4a, 4b. The spring stiffnesses c _h and c _v also change over time, for example as a result of oxidation reactions in the elastomer material. In this respect, an adaptation of these variables is necessary at certain time intervals. This parameter adaptation can be carried out when the motor vehicle 12 is stationary. Except for the portion of the gravitational acceleration, all accelerations in the direction of the z-axis 15 are equal to zero, as are the velocities and Eq. (10) simplifies to: $$c_H\Delta {\mathrm{e.g}}_H+c_v\Delta {\mathrm{e.g}}_v=c_H{\mathrm{e.g}}_{0H}+c_v{\mathrm{e.g}}_{0v}-\frac{m}{2}{S}_H-\frac{m}{2}{S}_v$$
  

und $\frac{J_y}{2l}{S}_v,$

da beide Sensorsignale S_h und S_v nur noch die Erdbeschleunigung enthalten: $$c_{h}l\Delta z_h-c_{v}l\Delta z_v=c_{h}l{z}_{0h}-c_{v}l{z}_{0v}{}_h-M_A$$

In Eq. (12) the terms also cancel each other out $\frac{J_y}{2l}{S}_H$

and $\frac{J_y}{2l}{S}_v,$

since both sensor signals S _h and S _v only contain the acceleration due to gravity: $$c_{H}l\Delta {\mathrm{e.g}}_H-c_{v}l\Delta {\mathrm{e.g}}_v=c_{H}l{\mathrm{e.g}}_{0H}-c_{v}l{\mathrm{e.g}}_{0v}{}_H-M_A$$

Electric vehicles usually have a creeping behavior, ie, in accordance with a conventional motor vehicle 12 with a torque converter, the motor vehicle 12 creeps away as soon as a brake pedal is released in a selector lever position D “drive”. This is realized by a vehicle control requesting the drive torque 25 with the value M _A corresponding to a specified creep torque M _K from the drive assembly 2 designed as an electric machine. For reasons of energy, building up a creep torque M _K only makes sense if the brake pedal has not been pressed too hard and it is therefore probable that motor vehicle 12 will creep.

3 shows a simplified start-up procedure. A braking torque M _Br , which is set via the service brakes of motor vehicle 12 and is specified by a driver of the same, the drive torque _MA of drive assembly 2 and a vehicle speed v _F , are plotted on a vertical axis. The time t is plotted on a horizontal axis. At the beginning of the curve shown, the brake pedal was still pressed hard. The build-up of a creeping torque does not make sense because the set braking torque M _Br is far above the value for the creeping torque M _K. A drive torque M _A =0 is set at the beginning of the process. It is proposed that a first measurement of the sensor signals Δz _h , Δz _v , S _h and S _v takes place at a point in time T ₁ . By inserting these values and M _A = 0 into Eq. (13) and (14) result in the first two conditional equations for the parameters z _oh , c _h , z _ov and c _v .

As a result of a reduction in the actuation of the brake pedal, the braking torque M _Br decreases and the vehicle controller specifies a drive torque M _A =M _K for the electric machine or the drive assembly 2 in accordance with the creep torque. It is proposed that at time T ₂ , after the drive torque has been built up and a certain amount of time has elapsed until the drive unit has settled in the elastomeric bearings 4a, 4b, a second measurement of the sensor signals Δz _h , Δz _v , _Sh and S _v done. The motor vehicle 12 is still at a standstill. Substituting these values and M _A = M _K into Eq. (13) and (14) result in two further conditional equations for the parameters z _0h , c _h , z _0v and c _v .

The four parameters z _0h , c _h , z _0v and c _v can be calculated with a total of four conditional equations. The damping constants d h and d _v can then be calculated from the relevant natural frequency ω of the movement of the drive unit 2 in the elastomeric bearings _{4a, 4b, an associated loss angle δ of the respective elastomeric bearing 4a, 4b and the spring stiffness c h and} c _v .

The measurements are preferably carried out several times during a start-up process and the sensor signals are averaged in order to minimize noise. It is also advantageous to average or filter the calculated parameter values over several start-up processes, since the real values only change slowly over time.

The method can be carried out during each start-up process, and the calculated parameters z _0h , c _h , z _0v and c _v can be compared with the values averaged over several measurements. If one or more deviations are above predetermined thresholds, a malfunction of the electric machine or of the drive unit 2 is concluded. This means that the electric machine does not correctly convert the drive torque M _A =M _K . A malfunction can thus be detected even before motor vehicle 12 drives off. An error reaction can thus reliably prevent an undesired vehicle movement.

The times shown above for the measurements are examples in the 3 to see. The first measurement for the drive torque M _A = 0 can also be carried out, for example, when transferring the selector lever from a P “Park” position to an N “Neutral” position or to the D “Drive” position. In order to leave the P “Park” position, it is usually necessary to actuate the brake pedal, which continues to ensure that motor vehicle 12 is stationary. When leaving the P “Park” position and transitioning to the N “Neutral” position, a mechanical parking lock in the drive unit 2 is typically deactivated, as a result of which the drive torque is reduced to M _A =0 even on a downhill road 23 .

For the second measurement, a drive torque _MA ≈ 0 can be built up as a test when the vehicle is stationary in the D "drive" position or in an R "reverse" position while the brake pedal is actuated with sufficient force. In this case, it must be ensured that the amount of the drive torque M _A remains far enough below the braking torque M _Br and that the motor vehicle 12 therefore does not move. As a further safety measure, the detection of an unwanted vehicle movement by checking the electric motor or wheel speeds is suggested. A countermeasure can be an immediate shutdown of the drive torque _MA . This alternative allows different measurements with a variation of the drive torque _MA , which improves the parameter determination. In addition, it is also possible with electric vehicles that do not have any creeping. It is also possible to carry out the tests as part of regular maintenance. In this case, a high absolute value drive torque M _A can be set in both effective directions as a test, which improves the parameter identification and also allows non-linearities of the elastomeric bearings to be detected.

In addition to determining the spring stiffnesses c _h and c _v , the parameter adaptation step described above mainly serves to determine the unstressed length z _0h of the rear vertical elastomer bearing 4a or its spring c _h and the unstressed length z _0v of the front vertical elastomer bearing 4b or its spring c _v to be determined, since these change over time as a result of settling of the elastomeric bearings 4a, 4b. The initial expansion or the unstressed spring length of the vibration damping device is determined. The gradient angle 24 with the value α _St is received via the sensor signals S _h and S _v of the acceleration sensors. It is advantageous that the parameter adaptation step can be carried out both when the roadway 23 is horizontal and when it is ascending or descending, an additional measurement of the gradient angle 24 is not required.

In addition or as an alternative to this, methods of parameter identification of dynamic systems can be used, with time profiles of the sensor signals Δz _h (t), Δz _v (t), Sh ( _t ) and S _v (t) and of the drive torque M _A (t) over of time t. For example, in this case the period between the time T ₁ and the time T ₂ of the 2 to be viewed as. Eqs. (10) and (12). The least squares method, which is based on a time-discrete system description based on Eqs. (10) and (12). The damping constants _dh and _dv can then also be better estimated under dynamic operating conditions. The dynamic model can use the time characteristics of the sensor signals to determine the energy dissipated due to material damping in the elastomer bearings 4a, 4b and, together with a measured ambient temperature, can conclude the elastomer temperatures based on the model. Temperature-dependent parameters can thus be determined more precisely.

The methods for parameter identification presented above can also be used to test the condition of the elastomeric bearings 4a, 4b. If one of the identified parameters is outside a predefined range, a defect in the associated elastomer bearing 4a, 4b is concluded and this is stored in the error memory of the vehicle control as part of the diagnostics.

The method described above can also be used in conventional motor vehicles 12 with an internal combustion engine and torque converter or an automated starting clutch. The first measurement for the drive torque _MA = 0 can be made as described above, for example when moving the selector lever from the P "Park" position to the N "Neutral" position or to the D "Drive" position or R the " Perform reverse position. The second measurement for a drive torque M _A ≈0 can be carried out after a creep or starting torque has built up, with motor vehicle 12 preferably still standing still. The drive torque M _A is preferably determined from the instantaneous speed ratios in the torque converter or using a contact force or contact pressure of a slipping starting clutch. A corresponding procedure is proposed in particular for ship and industrial transmissions.

A first measurement for the drive torque M _A =0 takes place with the clutches open. The second measurement for a drive torque M _A ≈0 is carried out when the clutch is slipping, with the drive torque M _A being determined using the contact pressure force or the contact pressure of the slipping clutch. Inertial effects of accelerated rotating parts must be taken into account. In the event of vibrations in the drive assembly 2, for example caused by an internal combustion engine, the sensor signals Δz _h , Δz _v , S _h and S _v are preferably filtered in order to remove higher frequencies and the main excitation of the internal combustion engine.

• Wurden die Parameter identifiziert, so kann mittels Gl. (12) während des Fahrbetriebes basierend auf den zeitlichen Verläufen der Sensorsignale Δz_h(t), Δz_v(t), S_h(t) und S_v(t) das Antriebsmoment M_A(t) bestimmt werden.

The determination of the drive torque is explained below:

• Once the parameters have been identified, Eq. (12) the drive torque M _A (t) can be determined during driving operation based on the time profiles of the sensor signals Δz _h (t), Δz _v (t), S _h (t) and S _v (t).

$$\Delta z_{vm}\left(t\right)=\Delta z_v\left(t\right)-z_{0v}$$

A conversion to modified sensor signals Δz _hm and Δz _vm is useful for this purpose: $$\Delta {\mathrm{e.g}}_{Hm}\left(t\right)=\Delta {\mathrm{e.g}}_H\left(t\right)-{\mathrm{e.g}}_{0H}$$

$$\Delta {\mathrm{e.g}}_{vm}\left(t\right)=\Delta {\mathrm{e.g}}_v\left(t\right)-{\mathrm{e.g}}_{0v}$$

Eq. (12) simplifies to $$\begin{array}{l}\frac{J_y}{2l}\Delta \ddot{}{\mathrm{e.g}}_{Hm}-\frac{J_y}{2l}\Delta \ddot{}{\mathrm{e.g}}_{vm}+c_{H}l\Delta {\mathrm{e.g}}_{Hm}+{\mathrm{i.e}}_{H}l\Delta \dot{}{\mathrm{e.g}}_{Hm}-c_{v}l\Delta {\mathrm{e.g}}_{vm}-{\mathrm{i.e}}_{v}l\Delta \dot{}{\mathrm{e.g}}_{vm}=\\ -M_A-\frac{J_y}{2l}{S}_H+\frac{J_y}{2l}{S}_v\end{array}$$

The sensor signals contain noise, so that a numerical formation of the derivations, for example by forming the difference, is typically problematic. It is therefore proposed to filter the sensor signals with identical, stable linear low-pass filters. In the Laplace domain, the low-pass filters are described, for example, by the transfer function G(s): $$G\left(s\right)=\frac{1}{a_2{\mathrm{<figure-callout\; id="2"\; label="s"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">s</figure-callout>}}^2+a_{1}s+1}$$

In the case of safety-related controls, a maximum error response time is usually specified, which is made up of the error detection time and the required shutdown time of the system. If the method and/or the device according to the invention are used in order to detect an incorrectly converted drive torque, then it is proposed that the dynamics of the low-pass filter according to Eq. (18) adapt to the error detection time. For example, a settling time for the step response of the low-pass filter is half the permitted error detection time. Otherwise, the filter dynamics are specified by the desired dynamics of the ascertained (filtered) drive torque.

$$\Delta z_{vm}*\left(s\right)=G\left(s\right)\Delta z_{vm}\left(s\right)$$

$$S_h*\left(s\right)=G\left(s\right)S_h\left(s\right)$$

$$S_v*\left(\text{s}\right)=G\left(s\right)S_v\left(s\right)$$

The quantities Δz _hm (s), Δz _vm (s), _Sh (s) and S _v (s) describe the sensor signals in the Laplacian range, the corresponding filtered sensor signals Δz _hm *(s), Δz _vm *(s) , S _h *(s) and S _v *(s) result from: $$\Delta {\mathrm{e.g}}_{Hm}*\left(s\right)=G\left(s\right)\Delta {\mathrm{e.g}}_{Hm}\left(s\right)$$

$$\Delta {\mathrm{e.g}}_{vm}*\left(s\right)=G\left(s\right)\Delta {\mathrm{e.g}}_{vm}\left(s\right)$$

$$S_H*\left(s\right)=G\left(s\right)S_H\left(s\right)$$

$$S_v*\left(\text{s}\right)=G\left(s\right)S_v\left(s\right)$$

The advantage of this procedure is that the derivatives sΔz _hm *(s), s ² Δz _hm *(s), sΔz _vm *(s) and s ² Δz _vm *(s) of the filtered sensor signals Δz _hm *(s) and Δz _vm *(s) can be determined directly from the filter transfer functions G(s).

und eine dritte Ausgangsgröße 33 durch dessen zweite Ableitung $\Delta z_{hm}^{{}^{··}}*\left(\text{t}\right)$

gegeben. 4 shows an example of the block diagram of a low-pass filter with the transfer function G(s) according to Eq. (18) in the time domain. An input variable 29 is given here by a sensor signal Δz _hm (t), a first output variable 31 is provided by the filtered sensor signal Δz _hm *(t), a second output variable 32 by its first derivative $\Delta {\mathrm{e.g}}_{Hm}^{{}^{·}}*\left(\text{t}\right)$

and a third output quantity 33 by its second derivative $\Delta {\mathrm{e.g}}_{Hm}^{{}^{··}}*\left(\text{t}\right)$

given.

Ein zweiter Transformationsblock 35 entspricht einem Integrator $\frac{1}{s}.$

Ein dritter Transformationsblock 36 entspricht ebenfalls einem Integrator $\frac{1}{s}.$

Ein vierter Transformationsblock 37 entspricht einem Multiplikator $\frac{1}{a_2}.$

Ein fünfter Transformationsblock 38 entspricht einem Multiplikator $\frac{a_1}{a_2}.$

Ein erster Verrechnungsblock 39 entspricht einer Addition. Ein zweiter Verrechnungsblock 40 entspricht einer Subtraktion.A first transformation block 34 corresponds to a multiplier $\frac{1}{a_2}.$

A second transformation block 35 corresponds to an integrator $\frac{1}{s}.$

A third transformation block 36 also corresponds to an integrator $\frac{1}{s}.$

A fourth transformation block 37 corresponds to a multiplier $\frac{1}{a_2}.$

A fifth transformation block 38 corresponds to a multiplier $\frac{a_1}{a_2}.$

A first calculation block 39 corresponds to an addition. A second calculation block 40 corresponds to a subtraction.

mit dem Laplace-transformierten Antriebsmoment M_A(s) und dem gefilterten, Laplace-transformierten Antriebsmoment M_A*(s): $$M_A*\left(s\right)=G\left(s\right)M_A\left(s\right)$$

By Laplace transformation of Eq. (17), multiply by G(s) and insert Eq. (19) to (22) are obtained $$\begin{array}{l}\frac{J_y}{2\mathrm{<figure-callout\; id="2"\; label="l\; s"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">l</figure-callout>}}{s}^{}{}^2\Delta {\mathrm{e.g}}_{Hm}*\left(\text{s}\right)-\frac{J_y}{2\mathrm{<figure-callout\; id="2"\; label="l\; s"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">l</figure-callout>}}{s}^{}{}^2\Delta {\mathrm{e.g}}_{vm}*\left(\text{s}\right)+c_{H}l\Delta {\mathrm{e.g}}_{Hm}*\left(\text{s}\right)+{\mathrm{i.e}}_{H}ls\Delta {\mathrm{e.g}}_{Hm}*\left(\text{s}\right)-\\ c_{v}l\Delta {\mathrm{e.g}}_{vm}*\left(\text{s}\right)-{\mathrm{i.e}}_{v}ls\Delta {\mathrm{e.g}}_{vm}*\left(\text{s}\right)\\ =-M_A*\left(\text{s}\right)-\frac{J_y}{2l}{S}_H*\left(\text{s}\right)+\frac{J_y}{2l}{S}_v*\left(\text{s}\right)\end{array}$$

with the Laplace-transformed drive torque M _A (s) and the filtered, Laplace-transformed drive torque M _A *(s): $$M_A*\left(s\right)=G\left(s\right)M_A\left(s\right)$$

By Laplace inverse transformation of Eq. (23) one obtains a conditional equation for the filtered drive torque M _A *(t), based on the filtered sensor signals and their time derivatives determined from the filter transfer functions G(s).

• Mit den modifizierten Sensorsignalen Δz_hm und Δz_vm vereinfacht sich die Gl. (10) zu: $$\begin{array}{l}\frac{\mathrm{<figure-callout\; id="2"\; label="m"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">m</figure-callout>}}{2}\Delta \ddot{}{z}_{hm}+\frac{\mathrm{<figure-callout\; id="2"\; label="m"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">m</figure-callout>}}{2}\Delta \ddot{}{z}_{vm}+c_h\Delta z_{hm}+d_h\Delta \dot{}{z}_{hm}+c_v\Delta z_{vm}+d_v\Delta \dot{}{z}_{vm}=\\ -\frac{m}{2}{S}_h-\frac{m}{2}{S}_v\end{array}$$
  

The plausibility is proven below:

• With the modified sensor signals Δz _hm and Δz _vm , Eq. (10) to: $$\begin{array}{l}\frac{m}{2}\Delta \ddot{}{\mathrm{e.g}}_{Hm}+\frac{m}{2}\Delta \ddot{}{\mathrm{e.g}}_{vm}+c_H\Delta {\mathrm{e.g}}_{Hm}+{\mathrm{i.e}}_H\Delta \dot{}{\mathrm{e.g}}_{Hm}+c_v\Delta {\mathrm{e.g}}_{vm}+{\mathrm{i.e}}_v\Delta \dot{}{\mathrm{e.g}}_{vm}=\\ -\frac{m}{2}{S}_H-\frac{m}{2}{S}_v\end{array}$$
  

Under ideal conditions, Eq. (25) fulfilled, in reality deviations are to be expected due to errors in modelling, in parameter identification and in the acquisition of the sensor signals. To check the plausibility of the sensor signals and the identified parameters, Eq. (25) introduced an equation error ε(t): $$\begin{array}{l}\frac{\mathrm{<figure-callout\; id="2"\; label="m"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">m</figure-callout>}}{2}\Delta \ddot{}{\mathrm{e.g}}_{Hm}+\frac{\mathrm{<figure-callout\; id="2"\; label="m"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">m</figure-callout>}}{2}\Delta \ddot{}{\mathrm{e.g}}_{vm}+c_H\Delta {\mathrm{e.g}}_{Hm}+{\mathrm{i.e}}_H\Delta \dot{}{\mathrm{e.g}}_{Hm}+c_v\Delta {\mathrm{e.g}}_{vm}+{\mathrm{i.e}}_v\Delta \dot{}{\mathrm{e.g}}_{vm}+\frac{m}{2}{S}_H+\\ \frac{m}{2}{S}_v=e_{\mathrm{e.g}}\left(t\right)\end{array}$$

mit dem Laplace-transformierten Gleichungsfehler ε_z(s) und dem gefilterten Laplace-transformierten Gleichungsfehler ε_z*(s): $${\epsilon }_z*\left(s\right)=G\left(s\right){\epsilon }_z\left(s\right)$$

By Laplace transformation of Eq. (26), multiply by G(s) and insert Eq. (19) to (22) are obtained $$\begin{array}{l}\frac{m}{2}{\mathrm{<figure-callout\; id="2"\; label="s"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">s</figure-callout>}}^2\Delta {\mathrm{e.g}}_{Hm}*\left(\text{s}\right)+\frac{m}{2}{\mathrm{<figure-callout\; id="2"\; label="s"\; filenames="DE102021113291B4\_0054.png,DE102021113291B4\_0055.png"\; state="\{\{state\}\}">s</figure-callout>}}^2\Delta {\mathrm{e.g}}_{vm}*\left(\text{s}\right)+c_H\Delta {\mathrm{e.g}}_{Hm}*\left(\text{s}\right)+{\mathrm{i.e}}_{H}s\Delta {\mathrm{e.g}}_{Hm}*\left(\text{s}\right)+\\ c_v\Delta {\mathrm{e.g}}_{vm}*\left(\text{s}\right)+{\mathrm{i.e}}_{v}s\Delta {\mathrm{e.g}}_{vm}*\left(\text{s}\right)+\frac{m}{2}{S}_H*\left(\text{s}\right)+\frac{m}{2}{S}_v*\left(\text{s}\right)=e_{\mathrm{e.g}}*\left(s\right)\end{array}$$

with the Laplace-transformed equation error ε _z (s) and the filtered Laplace-transformed equation error ε _z *(s): $$e_{\mathrm{e.g}}*\left(s\right)=G\left(s\right)e_{\mathrm{e.g}}\left(s\right)$$

A conditional equation for the filtered equation error ε _z *(t) is obtained by Laplace inverse transformation, based on the filtered sensor signals and their time derivatives. Eq. (27) describes the vertical dynamics. As a result of the bumps in the road 26, the drive unit 2 will constantly move in the elastomeric bearings 4a, 4b, 4c. The filtered equation error ε _z *(t) thus represents a good measure of the quality of the calculations of the parameters. If the absolute value of the equation error is greater than a specified threshold, then an error in the sensor system or the mechanics is concluded, for example a hitting of the Drive unit 2 on the body 13 or a foreign body that is stuck between the drive unit 2 and the body 13. Ultimately, it is then also concluded that there is an error in the determination of the filtered drive torque M _A *(t). The filtered drive torque M _A *(t) can thus be checked for plausibility.

If a correctly determined, filtered drive torque M _A *(t) deviates from the driver's desired torque or the filtered driver's desired torque by a predetermined threshold, then a fault in the drive system, for example in the drive unit 2, is concluded and fault reactions are triggered, for example an emergency Operation with reduced drive torque M _A (t) or switching off the system. If the equation error ε _z *(t) is too large, this diagnosis is no longer meaningful and an error reaction should also take place.

A correctly determined, filtered drive torque M _A *(t) can contain dynamic torque components that result from the movements of the drive unit 2 in the vibration damping device 4, for example by drive shafts twisting cyclically as a result of the movements of the drive unit 2 and the drive torque acting on the drive unit 2 25 influence. The movements of the drive unit 2 can be determined based on the time profiles of the modified sensor signals Δz _hm (t) and Δz _vm (t) or the filtered sensor signals Δz _hm *(t), Δz _vm *(t) and together with the torsional spring stiffness and the torsion spring damping of the drive shafts to determine these dynamic torque components. If the filtered drive torque M _A *(t) is corrected by these dynamic torque components, the drive torque generated by the drive assembly 2 through energy conversion remains. It is also possible to correct the filtered drive torque M _A *(t) by torque components that arise from the effects of inertia of rotating, rotationally accelerated parts in the drive unit 2 . The corrected drive torque determined in this way can then be compared with the driver's desired torque or the filtered driver's desired torque, which improves the error detection.

For the sake of simplicity, the exemplary embodiment contains a few simplifications. In reality, elastomeric bearings show a non-linear behavior with a frequency dependency. It is therefore proposed to modify the transfer functions G(s) of the low-pass filters to determine the filtered sensor signals Δz _hm *(s) and Δz _vm *(s) or to determine the derivatives sΔz _hm *(s), sΔz _vm *(s). . The changed amplitude response of the modified low-pass filter can at least partially take into account the frequency dependency of the elastomeric bearings 4a, 4b. Multiplying by c _h , c _v , d _h , d _v then results in frequency dependencies in the forces, which simulates the frequency-dependent non-linearities of the elastomers. An increasing increase in rigidity over the deformation, the so-called progression, can be taken into account, for example, by increasing c _h , c _v with larger deformation Δz _hm *(s) and Δz _vm *(s).

In order to determine the filtered drive torque M _A *(t), it is proposed that the low-pass filter also be adapted in the event of vibrations in the drive unit, for example caused by an internal combustion engine or tooth meshing. This is intended to remove higher frequencies and the main excitation of the combustion engine. The low-pass filters are preferably adapted to the current speed of drive unit 2 .

The dynamics of the drive assembly 2 can be adjusted in a corresponding manner 2 Evaluate in the direction of the x-axis 14.

The bearing force F _x of the rear horizontal elastomer bearing results from the spring stiffness c _x , the damping constant d _x and the unstressed length x ₀ of the rear horizontal elastomer bearing 4c or the horizontal spring $$f_x=c_x\left(x_L+x_0-x\right)+{\mathrm{i.e}}_x\left({\dot{x}}_L-\dot{x}\right).$$

The variables x and x _L describe the positions 14 and 41 of the assembly-side and body-side attachment point of the rear horizontal elastomer bearing in the inertial system.

The following applies to the center of gravity 17 of the drive assembly 2 in the horizontal direction: $$m\ddot{x}=c_x\left(x_L+x_0-x\right)+{\mathrm{i.e}}_x\left({\dot{x}}_L-\dot{x}\right)-mG\text{sin}\left(a_{St}\right)$$

A sensor signal Δx is available which characterizes the distance between the contact points of the elastomer bearing 4c, which are located on the drive unit 2 and the body 13: $$\mathrm{\Delta }x=x-x_L$$

Here, too, the acceleration measuring device 10 designed as an acceleration sensor is preferably installed, which records the acceleration of the body 13 in the direction of the x-axis 14 at the body-side stop point of the horizontal elastomer bearing 4c. The acceleration sensor also reacts to the gravitational acceleration g. The sensor signal S _x is therefore present: $$S_x={\ddot{x}}_L+G\text{sin}\left(a_{St}\right)$$

As an alternative to this, the sensor signal S _x can be derived from the sensors of an ESP system, as described above.

If one subtracts Eq. multiplied by the mass m. (31) from Eq. (29), this results in: $$\begin{array}{l}m\left(\ddot{x}-{\ddot{x}}_L\right)-mG\text{sin}\left(a_{St}\right)=c_x\left(x_L+x_0-x\right)+{\mathrm{i.e}}_x\left({\dot{x}}_L-\dot{x}\right)-\\ mG\text{sin}\left(a_{St}\right)-m{S}_x\end{array}$$

After inserting Eq. (30) in Eq. (32) one obtains: $$m\mathrm{\Delta }\ddot{}x+c_x\mathrm{\Delta }x+{\mathrm{i.e}}_x\mathrm{\Delta }\dot{}x=c_x{x}_0-m{S}_x$$

The spring stiffness c _x and the unstressed length x ₀ of the elastomer bearing 4c can be determined by means of two measurements with the motor vehicle 12 stationary and different pitch angles α _St . It is also possible to average the results of more than two measurements. The damping constant d _x can then be calculated from the relevant natural frequency ω of the movement of the drive unit 2 in the x-direction, an associated loss angle δ of the elastomer bearing 4c and the spring stiffness c _x . Also based on Eq. (33) a filtered equation error ε _x *(t) can be determined. If the equation error is greater than a predetermined threshold in terms of amount, this indicates an error in the sensors or the Close the mechanics according to the procedure shown above. If signals from the ESP sensors are used, it is proposed to use the filtered equation errors ε _z *(t) and ε _x *(t) to test the ESP sensors.

Unbalanced or misaligned couplings can be determined as follows:

und die gefilterte Lagerkraft des vorderen vertikalen Elastomerlagers $$F_{zv}*\left(t\right)=-c_v\mathrm{\Delta }{z}_{vm}*\left(t\right)-d_v\Delta z_{vm}^{\cdot }*\left(t\right)$$

Individual terms from Eq. (27) describe the filtered bearing force of the rear vertical elastomer bearing 4a: $$f_{\mathrm{e.g}H}*\left(t\right)=-c_H\mathrm{\Delta }{\mathrm{e.g}}_{Hm}*\left(t\right)-{\mathrm{i.e}}_H\Delta {\mathrm{e.g}}_{Hm}^{\cdot }*\left(t\right)$$

and the filtered bearing force of the front elastomeric vertical bearing $$f_{\mathrm{e.g}v}*\left(t\right)=-c_v\mathrm{\Delta }{\mathrm{e.g}}_{vm}*\left(t\right)-{\mathrm{i.e}}_v\Delta {\mathrm{e.g}}_{vm}^{\cdot }*\left(t\right)$$

The filtered bearing force of the horizontal elastomer bearing 4c can be determined in a corresponding manner with a modified sensor signal Δx _m = Δx - x ₀ or a filtered modified sensor signal Δx _m *: $$f_x*\left(t\right)=-c_x\mathrm{\Delta }{x}_m*\left(t\right)-{\mathrm{i.e}}_x\Delta x_m^{\cdot }*\left(t\right)$$

It is proposed to detect imbalances or incorrectly aligned clutches during operation based on a speed-dependent, out-of-phase oscillation of the filtered bearing force of the horizontal elastomeric bearing 4c compared to the oscillations of the filtered bearing forces of the two vertical elastomeric bearings 4a, 4b. In the case of a rotating radial force, there are dependencies in the vibration amplitudes of the three forces. An imbalance leads to a rotating radial force that increases with the speed and thus to an increase in the vibration amplitudes. If the clutch is not correctly aligned, this increase is typically absent, i.e. an imbalance or an incorrectly aligned clutch can be detected on the basis of the vibration amplitudes, a speed dependency of the vibration amplitudes and the phase offset of the vibrations. This approach is also possible for marine and industrial gears.

The method described can also be used by determining the state of deformation of the at least one elastomer bearing 4a, 4b, 4c by measuring a local force or a local surface pressure in or on the elastomer. The relationship between the local force or the local surface pressure and the state of deformation over time also changes as a result of aging-related creep phenomena and inhomogeneous changes in stiffness. The relationship between the state of deformation and the total bearing force also changes over time. The method can be used to establish a connection between the measured variable, i.e. the local force or a local surface pressure, and the total bearing force by means of parameter identification.

The dielectric elastomer sensor briefly described above can have at least two electrodes, which are preferably thin compared to the elastomer bearing 4a, 4b, 4c, between which an elastomer layer is arranged as a dielectric.

Claims (12)

Method for determining at least one force (3) acting on a drive assembly (2) and/or at least one drive torque (25) acting on the drive assembly (2), the drive assembly (2) being mounted on at least one vibration damping device (4) in a receiving device ( 5), wherein a deflection state and/or a time profile of the deflection state of the vibration damping device (4) is determined by directly measuring a distance (7) between test areas (8a, 8b), the test areas (8a, 8b) being arranged in this way are that the distance (7) is determined by the deflection state, and at least one acceleration (11) of the receiving device (5) is determined, characterized in that a parameter adaptation step is provided in order to determine an initial extension of the vibration damping device (4). procedure after claim 1 , characterized in that when determining the at least one acceleration (11) an influence of the acceleration due to gravity is taken into account. procedure after claim 1 or 2 , characterized in that at least one parameter of the vibration damping device (4) is identified by means of a dynamic model for a movement of the drive unit (2). Procedure according to one of Claims 1 , 2 or 3 , characterized in that the drive torque (25) of the drive assembly (2) is determined based on the determined deflection state and/or the determined time profile of the deflection state of the vibration damping device (4) and/or the determined acceleration (11) and/or the identified parameters becomes. Procedure according to one of Claims 1 until 4 , characterized in that the dynamic model maps at least two degrees of freedom of the movement of the drive unit (2) in the vibration damping device (4). Method according to claim one of Claims 1 until 5 , characterized in that the vibration damping device (4) has at least one elastomer bearing (4a, 4b, 4c), the measurement of the distance (7) being carried out by means of a dielectric elastomer sensor. Device (1) for determining at least one force (3) acting on a drive assembly (2) and/or at least one drive torque (25) acting on the drive assembly (2), the drive assembly (2) being supported on at least one vibration damping device (4) in a recording device (5), having - at least one distance measuring device (6) which is set up for the direct measurement of a distance (7) and/or a time profile of the distance (7) between test areas (8a, 8b), the test areas (8a, 8b) are arranged in such a way that the distance (7) is determined, preferably at all times, by a deflection state of the vibration damping device (4), - a computing device (9) which is set up to determine the deflection state from the distance (7). , and - at least one acceleration measuring device (10), which is set up to determine an acceleration (11) of the recording device (5), characterized in that the vibration damping device (4) has at least one elastomer bearing (4a, 4b, 4c) and the distance measuring device ( 6) is designed as a dielectric elastomer sensor. Device (1) after claim 7 , characterized in that the computing device (9) is set up to identify at least one parameter of the vibration damping device (4) by means of a dynamic model for a movement of the drive unit (2). Device (1) after claim 7 or 8th , characterized in that the computing device (9) is set up to calculate the drive torque based on the determined deflection state and/or the determined time profile of the deflection state of the vibration damping device (4) and/or the determined acceleration (11) and/or the identified parameters (25) of the drive unit (2) to determine. Device (1) after claim 7 , 8th or 9 , characterized in that at least two vibration damping devices (4) are provided and the acceleration measuring device (10) has at least two acceleration sensors, the at least two acceleration sensors being set up to detect a vertical acceleration of the recording device (5) at a position of the at least two vibration damping devices (4) capture. Device (1) according to one of Claims 7 until 10 , characterized in that the dielectric elastomer sensor has at least two electrodes, preferably thin in comparison to the elastomer bearing (4a, 4b, 4c), an elastomer layer being arranged as a dielectric between the electrodes. Motor vehicle (12) with a drive assembly (2) which is mounted on at least one vibration damping device (4) in a receiving device (5), characterized in that a device (1) for determining at least one force ( 3) and/or at least one drive torque (25) acting on the drive unit (2), according to one of Claims 7 until 11 is provided and/or at least one acting on the drive unit (2). Force (3) and/or at least one drive torque (25) acting on the drive unit (2) by means of a method according to one of Claims 1 until 6 is determined.

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