DE102007016420B4 - Test bench and procedure for testing a drive train - Google Patents

Abstract

Method for testing a drive train (12) of a motor vehicle on a test stand (10), wherein an output (21, 23, 25, 27) of the drive train (12) is coupled to at least one first electrical machine (28, 30, 32, 34) which is controlled so as to generate predetermined load torques on the output (21, 23, 25, 27), nominal values of the load torques being predetermined by a first model (42) as a function of predetermined properties of the motor vehicle and a travel distance of the motor vehicle, wherein a drive (38) of the drive train (12) is coupled to a second electric machine (36) which generates drive torques on the drive (38) and which is controlled such that the generated drive torques correspond to setpoint values (SW_M_A) which are determined by a second model (44) can be generated, which simulates a producible by an internal combustion engine, in the drive train (12) effective torque, characterized in that the second elektrisc he machine (36) is coupled to a secondary mass (50) of a dual mass flywheel (46) and that of the second model ...

Description

The invention relates to a method for checking a drive train of a motor vehicle according to the preamble of claim 1 and to a test stand according to the preamble of claim 8.

Here, a powertrain is understood here to be the sum or a part of the sum of components with which in a motor vehicle the force and the torque of an internal combustion engine are transmitted to the drive wheels of the motor vehicle. Typical components are gearboxes, transfer cases, cardan shafts, differentials and axle shafts. In this sense, for example, a change gear is already understood as a drive train.

For a distinction between a drive and an output is assumed in this application of a power flow from the engine to the wheels, although the physical power flow in the push mode can also run in the opposite direction. In other words: If one considers a component of the drive train which has a first side facing the internal combustion engine and a second side facing a drive wheel of the motor vehicle, then the first side represents the drive and the second side represents the output.

From the US 2003/0167143 A1 a test stand for testing a drive train is known, in which a first electric machine generates predetermined load torques at an output of the drive train to be tested, while a second electric machine can be coupled to the drive of the drive train and the generated drive torques correspond to predetermined target values.

The DE 43 28 537 A1 also discloses a test stand and a method for testing a drive train in which a dual-mass flywheel between engine and transmission can be simulated by means of a control in its mode of action.

A generic method and a generic test stand is in each case from EP 1 037 030 B1 known. In the known method and test stand, the drive train is driven by an internal combustion engine. The torques acting on a particular vehicle and driving profile from a road via wheels of the vehicle in the drive train are applied by electric motors as electric machines, which are coupled instead of the wheels to the drive train. The control of the electric motors is such that each electric machine generates predetermined load torques at the output. In this case, setpoint values of the load torques are predetermined by a first model as a function of predefined properties of the motor vehicle and a travel distance of the motor vehicle such that the wheel torques acting on the drive train in real driving operation are reproduced as realistically as possible. In the known subject matter, tire slip is already taken into account in the control of the first electric machines on the output side of the drive train.

Such powertrain test stands are used, for example, in the development of a gearbox to be able to test the change gear in conjunction with the other components of the drive train and the internal combustion engine.

To shorten the entire development time of a motor vehicle and / or a drive train of the motor vehicle, a parallel development of internal combustion engine and change gear is desired. However, at an early stage of development, the internal combustion engine will still not have a marketable maturity in terms of performance and durability. A long-term testing of a change gear may then require several internal combustion engines. These are very expensive because they are not yet available in the early stage of development as series products and require a lot of work to produce them. As a result, the costs for a durability test of the gearbox are driven up.

This undesirable effect is further increased by downtimes for replacement of the internal combustion engine on the test bench and repairs. If, on the other hand, one waits until sufficiently stable internal combustion engines with the desired performance are available in larger numbers, the overall development time for the entire vehicle may be extended.

Against this background, the object of the invention is to provide a method and a test stand which reduces the mentioned disadvantages of a high cost and / or a prolonged overall development time for a vehicle.

This object is achieved in each case with the features of one of the independent claims.

In the invention, the internal combustion engine is replaced in powertrain test runs by the second electric machine. This generates drive torques on the drive of the drive train and is controlled in such a way that a torque that can be generated by an internal combustion engine and that is effective in the drive train is reproduced. In this case, setpoint values of the drive torques are generated by a second model, which simulates the torque output of an internal combustion engine.

The invention also makes it possible to test a drive train under dynamic endurance conditions even when an engine with the corresponding degree of maturity is not yet available. The resulting independence from the development process of the engine results in a desired streamlining of the entire vehicle development process. On the other hand, it also benefits the quality, since components of the powertrain can be tested from the outset with load specification data. In addition, this method leads to a significant cost reduction by saving expensive prototype engines and reducing downtime through repair work.

In addition, investigations are made possible by the flexibility of the electric drive, which would not be conceivable with internal combustion engines such. B. rattle tests. Here, both the non-uniformity in frequency and amplitude and speed and torque can be specified independently to simulate targeted any operating points. Otherwise, a separate test bench would be required for this type of experiment.

The invention is characterized in that the second electric machine is coupled to a secondary mass of a dual-mass flywheel and the desired values generated by the second model correspond to a torque that would be exerted on the secondary mass by an elastic coupling of the secondary mass with a primary mass of the dual-mass flywheel.

As a result, the primary mass and the elastic coupling are taken over into the modeling as it were. During the tests on the test bench, the electric machine then drives not the primary mass, but directly the secondary mass of the dual-mass flywheel. As will be explained in more detail below, the electric machine then only has to generate an alternating component with a reduced amplitude. As a result, a smaller electric machine with a smaller moment of inertia can be used. Another advantage is the ability to simulate torques of internal combustion engines whose mass moment of inertia is significantly lower than that of the electric drive used.

Further advantages will be apparent from the dependent claims, the description and the attached figures.

It is understood that the features mentioned above and those yet to be explained below can be used not only in the particular combination given, but also in other combinations or in isolation, without departing from the scope of the present invention.

drawings

Embodiments of the invention are illustrated in the drawings and are explained in more detail in the following description. In each case, in schematic form:

1 an overview of an embodiment of a test bench according to the invention;

2 a structure of control of the test bench,

3 a representation of a complete dual-mass flywheel and a part thereof, which is used in a preferred embodiment of the invention; and

4 a block diagram of a second model, the method aspects of the invention disclosed.

In detail, the shows 1 a test bench 10 on which a powertrain 12 a motor vehicle is mounted. In the embodiment of 1 indicates the drive train 12 a change gear 14 with a transfer case and a front axle differential, a cardan shaft 16 , a rear axle differential 18 and drive shafts 20 . 22 . 24 . 26 with power take-offs 21 . 23 . 25 and 27 on. When used in a motor vehicle are the outputs 21 to 27 connected to wheels of the motor vehicle. For testing the powertrain 12 on the test bench 10 are the outputs 21 . 23 . 25 . 27 each to a first electric machine 28 . 30 . 32 . 34 coupled. It is understood, however, that the invention also drive trains in other embodiments, for example in a front-wheel drive, a rear-wheel drive with front engine, mid-engine or rear engine and corresponding four-wheel variants is applicable.

The simulation of the wheel-side loads already takes place on existing powertrain test stands with sufficient accuracy and corresponds to the state of the art, for example according to the aforementioned EP 1 037 030 B1 ,

At the test bench 12 after 1 the driving motor is not a combustion engine, but as a second electric machine 36 realized with a drive 38 of the powertrain 12 connected is. The first and second electric machines 28 to 34 and 36 may alternatively be operated as electric motors or generators to both driving and decelerating torques on the drives 21 to 27 and the drive 38 to be able to produce.

The electrical machines 28 to 34 and 36 be from a controller 40 controlled. The control 40 is particularly adapted to at least one of the first electric machines 28 to 34 to control so that these predetermined load torques on the coupled drives 21 to 27 produce. In addition, the controller 40 set up the second electric machine 36 so that the generated drive torques correspond to a torque that is an internal combustion engine in the drive train 12 would produce.

Like the schematic representation of the controller 40 in the 2 clarifies, is the controller 40 in particular set to setpoints SW_M_B the load torque through a first model 42 as a function of vehicle parameters FP1,... FPn, ie as a function of properties of a motor vehicle, and as a function of a travel distance of the motor vehicle, ie as a function of distance parameters SP1,..., SPm. Here, a preferred embodiment is characterized in that the desired values of the load torques are predetermined by the first model taking into account a tire slip. Typical vehicle parameters include wheelbase, gauge, tire parameters such as coefficients of friction and vehicle mass. Typical route parameters are, for example, curve radii, pitch angles and coefficients of friction of the roadway, the list in each case not being meant exhaustively.

In addition, the controller 40 set to setpoint values SW_M_A the drive torque through a second model 44 to generate that in the drive train producible by the internal combustion engine 12 simulates effective torque. The second model processes this 44 In one embodiment, a signal FW, which represents a torque request of a driver and a signal n_36, the rotor speed of the second electric machine 36 represents. blocks 43 and 44 in the controller 40 represent the formation of drive signals to which the first electric machines 28 . 30 . 32 . 34 and the second electric machine 36 be controlled in response to the setpoint values SW_M_B and SW_M_A in order to generate corresponding torques for the respective setpoint specification.

In a preferred embodiment, the second electric machine 36 to a secondary mass of a dual mass flywheel as a drive 38 of the powertrain 12 coupled.

Such a dual mass flywheel 46 is in the 3a shown schematically and has an elastic coupling 52 between a primary mass 48 and a secondary mass 50 , The primary mass 48 has a moment of inertia JP and is rotationally fixed to the crankshaft when used in a motor vehicle 49 connected to the internal combustion engine. The elastic coupling 52 has a spring constant C_PS and transmits the torque of the primary mass 48 on the secondary mass 50 which has an inertia moment JS. The secondary mass 50 is about another elastic coupling 54 having a spring constant C_SG with a transmission input shaft 55 connected. In this constellation is the dual mass flywheel 46 in a drive train 12 used when installed in a motor vehicle.

In a preferred embodiment, the in 3b is shown, the second electric machine 36 in test bench tests on the secondary mass 50 coupled to the dual mass flywheel and the effect of the primary mass 48 and the elastic coupling 52 gets through the second model 44 simulated. In other words, this embodiment is characterized in that the with the second model 44 generated setpoints SW_M_A correspond to a torque that of an elastic coupling 52 the primary mass 48 with the secondary mass 50 on the secondary mass 50 would be exercised.

The advantages of this embodiment are clear from the following background:
As is known, the discontinuous mode of operation of a conventional internal combustion engine forms in the time course of its speed and its torque. In particular, the working cycles of each cylinder with the ignition frequency are formed, so that the speed and torque curves each have a fluctuating DC component (mean value) and one oscillating with the ignition frequency or a multiple of the ignition frequency alternating component.

An electrical machine, which should simulate such courses as realistically as possible, must be able to generate in particular the DC component and in addition to be able to generate the AC component.

The latter requires, for example, comparatively fast angular momentum changes (high dynamics) with the comparatively large torque values required for this purpose. However, an electric machine capable of producing such torques usually has a substantial greater moment of inertia than a combustion engine of equal power.

However, such a larger moment of inertia is disadvantageous because a realistic replica of the alternating component requires as similar moments of inertia of the respective masses to be accelerated and braked. The moment of inertia of the second electric machine 36 should come as close as possible to the moment of inertia of the internal combustion engine.

Furthermore, it applies that the acceleration torques resulting from dynamic processes of the second electric machine 36 already have to be applied to their own acceleration and therefore at the interface to the rest of the powertrain 12 are no longer available, proportional to the moment of inertia of the second electric machine 36 are. The same power in principle larger moment of inertia of the second electric machine 36 therefore requires a large torque for self-acceleration, which undesirably increases the difference to the smaller moment of inertia of an internal combustion engine. Since a high dynamics is required (ie: a simulation of vibrations up to 450 Hz) is the lowest possible moment of inertia of the second electric machine 36 he wishes.

In this context, the design provides the following advantages: the fact that the second electric machine 36 torques to be generated correspond to a torque, that of an elastic coupling 52 the secondary mass 50 with a primary mass 48 of the dual mass flywheel 46 on the secondary mass 50 would be exercised, the requirement for the maximum torque from the second electric machine decreases 36 is to provide. There are two reasons:
On the one hand, the physical moment of inertia JP of the primary mass falls 48 of the dual mass flywheel 46 path. It is not used in the bench test and therefore does not need to be slowed down or accelerated. In other words: The physically real coupling of the second electric machine 36 to the secondary mass 50 means that the real primary mass 48 and the means (springs) for the elastic coupling 52 between primary mass 48 and secondary mass 50 when driven by the second electric machine 36 be omitted.

This results in a desired approximation of the relevant moments of inertia: Looking at the before the change gear 14 arranged moments of inertia, the result is in the case of the drive by the internal combustion engine, the sum of the moments of inertia of the engine, the primary mass 48 and the secondary mass 50 as a relevant moment of inertia. In contrast, in the case of the drive by the second electrical 36 Machine is the sum of the moments of inertia of the second electric machine 36 and the secondary mass 50 as a relevant moment of inertia.

By aligning the relevant moments of inertia, the similarity of the second electrical machine 36 generated interaction with the rest of the drive train 12 further improved with the interaction generated by the internal combustion engine, so that a realistic vibration behavior can be achieved with a minimum of control intervention in the test bench trial.

It also creates a real primary mass 48 a certain low-pass effect in relation to the alternating component. This means that for a real existing primary mass 48 the alternating component before the primary mass 48 has a larger amplitude than behind the unit of primary mass 48 and elastic coupling 52 that is, before the secondary mass 50 ,

By omitting the primary mass 48 must be the second electric machine 36 only produce the smaller amplitudes, which in a complete powertrain 12 in front of the secondary mass 50 occur. In other words, the fact that the design of the secondary mass 50 simulating occurring torque, the second electric machine must 36 only generate an alternating component with correspondingly reduced amplitude.

As a result, a smaller electric machine 36 be used with a smaller moment of inertia, which is more similar to the moment of inertia of the internal combustion engine.

Another advantage is the ability to simulate torques of internal combustion engines whose mass moment of inertia is significantly lower than that of the electric drive used.

4 shows a block diagram of the second model 44 , A goal of simulating the torque values of the internal combustion engine with the second model 44 in particular, is all for possible damage to the transmission 14 to simulate relevant burdens. On the drive side, these consist essentially of the average torque output by the internal combustion engine and the torsional vibrations which the internal combustion engine transmits via the dual mass flywheel 46 in the drive train 12 brings.

The second model 44 consists of three submodels 56 . 58 and 60 together, who the Describe internal combustion engine with its nonuniform output torque and the mechanical transmission path to the transmission input. The following is the physical structure of the second model 44 explained before on details of the engine model 56 will be received.

With the combustion engine model 56 the torque M_VM of the internal combustion engine is simulated. In this case, a mean torque value (DC component M_DC) and an average-free value (AC component M_AC) are formed.

Both values are in a link 62 added. The block 64 represents the formation of the torque average M_DC as a function of a driver's request FW, an engine speed n and a CAN bus engagement. In the driver's request and in the CAN bus intervention, torque requirements are formed. The engine speed n would also be the (fictitious) speed of the (imaginary, on the test bench 10 not real existing) primary mass 48 , which still has the full, undamped ripple.

The alternating component M_AC is displayed in the block 66 formed essentially as a function of the rotational speed n and a dimension F for the combustion chamber filling. As a measure F, for example, an opening angle of a throttle in question.

Then in the block 58 that influences the non-existent primary mass 48 considered, the speed n of the internal combustion engine modeled. This speed n would be the drive of a real combustion engine, the speed of the primary mass 48 of the dual mass flywheel 46 correspond. According to the equation of motion for a rotating mass, its speed or angular frequency is proportional to the time integral of the sum of the acting torques. On the primary mass 48 would be the sum of the torque M_VM of the internal combustion engine and of the drive train via the elastic coupling 52 between secondary mass 50 and primary mass 48 on the primary mass 48 applied torque M_ZMS act. In the block 58 becomes the fictitious angular frequency of an imaginary primary mass 48 a dual mass flywheel 46 by normalizing said integral to a fictitious moment of inertia JP of the imaginary primary mass 48 educated.

The resulting (fictitious) speed n of the imaginary primary mass 48 involved in a powertrain 12 in the motor vehicle would correspond to the speed n of the internal combustion engine, is used as an input to the engine model 56 recycled.

Furthermore, this fictitious speed n in the following block 60 used to, one of the fictitious primary mass 48 via an elastic coupling 52 on the real existing secondary mass 50 to transmit transmitted torque M_ZMS.

The last-mentioned torque M_ZMS would be in a real drive train 12 in a motor vehicle, the primary mass 48 and the secondary mass 50 of the dual mass flywheel 46 against elastic restoring moments rotate against each other, wherein the angle of rotation as a time integral of the sum of the angular frequencies of the primary mass 48 and secondary mass 50 of the dual mass flywheel 46 results. The to the angular frequency of the secondary mass 50 proportional speed of the secondary mass 50 corresponds to the drive of the secondary mass 50 through the second electric machine 36 its rotor speed n_36 and is measured as such or is otherwise, for example, from the control of the second electric machine 36 , known.

The time integral of the difference between the real rotor speed n_36 and the notional speed n, or the angular frequency proportional thereto, after a multiplication by the spring constant C_PS, supplies the coupling between the primary mass 48 and secondary mass 50 that ultimately on the secondary mass 50 applied torque M_ZMS, that of the second electric machine 36 is to provide. The torque M_ZMS thus calculated thus represents the desired set value SW_M_A for the torque that is generated by the second electric machine 36 on the real existing secondary mass 50 of the dual mass flywheel 46 should be exercised.

When modeling the setpoint SW_M_A = M_ZMS, it was assumed that the torque was transmitted via a rotating primary mass 48 of the dual mass flywheel 46 was transferred. In an intermediate step of the modeling, the sum of the on the primary mass was 48 considered acting torques. There goes the secondary mass 52 calculated setpoint M_ZMS with opposite sign than that of the drive train 12 over the elastic coupling 52 between secondary mass 50 and primary mass 48 on the primary mass 48 retroactive torque.

The desired value SW_M_A = M_ZMS determined in this way is used to control the second electrical machine 36 used. That of the second electric machine 36 on the drive train 12 applied torque then largely corresponds to the torque that the internal combustion engine on the drive train 12 would exercise.

The two-mass flywheel model from the blocks 58 and 60 forms the mass moment of inertia of the internal combustion engine and the primary mass 48 of the dual mass flywheel 46 as well as the influence of elastic coupling 52 between the primary mass 48 and the secondary mass 50 of the dual mass flywheel 46 to.

Through this modeling of the dual mass flywheel 46 can also be simulated internal combustion engines whose mass moment of inertia is significantly lower than that of the electric drive used.

That in the block 64 of the combustion engine model 56 mimicked (simulated) average output torque M_DC is determined mainly by the driver's request FW (torque demand by the driver) and the speed n. Also important are dynamic influences due to self-acceleration and combustion build-up (response time), as they occur during load changes. Assuming identical boundary conditions and an identical setpoint course, the mean torque M_DC can be adjusted by a direct comparison with measurements of internal combustion engine test benches.

The mean torque M_DC of the internal combustion engine can therefore optionally be taken from a characteristic map (driver request, torque requirements of other control devices or functions, represented by a CAN bus input) or calculated from a more complete thermodynamic real-time model, which also includes other factors such as ignition times and or lambda values of combustion chamber fillings considered the internal combustion engine. In such a real-time model, the average torque is calculated according to the thermodynamic laws. It is then relatively accurate not only for an engine in series production but also for a motor in an earlier development phase. The parameterization of the motor model is relatively simple with the basic physical motor data.

The CAN bus input is used to take into account torque requirements of other ECUs. It is desirable that all functions that directly affect the drivetrain load be replicated. These are, on the one hand, functions for responding to the requirements of other control devices (eg torque reduction during a shift in an automatically shifting change gearbox), on the other hand, functions that independently influence the torque build-up and the torque output (eg a load cycle shock absorption).

Mean-engine-free gas and mass torque curves for mapping the internal combustion engine alternating component M_AC are superimposed on the mean engine torque M_DC over the entire torque M_VM, which is shown in block 66 be formed.

The alternating component M_AC results as a result of its discontinuous combustion process, the kinematic conditions between piston and crankshaft movement, ie the gas and mass torque curves and forms a cause for torsional vibrations in the drive train. These are critical for continuously loaded components such as shafts because they dramatically increase the number of cycles. Therefore, it is important to consider the AC component M_AC in the simulation. The amplitude of the gas torque depends on the pressure curves of the individual cylinders, whereas the amplitude of the mass moment is speed-dependent. As the speed increases, the proportion of the mass momentum in the total engine torque increases.

In order to be able to carry out the modeling in real time, a preferred embodiment sees a characteristic model 66 that works with the premise of a rigid crankshaft. In order to take account of the operating state-dependent weighting of gas and mass-flow influences, these are preferably modeled in separate characteristic curves and then superimposed. The amplitudes are each defined separately in maps via speed or throttle and speed, the actual vibration is given as a normalized time course over 720 ° crank angle.

A preferred internal combustion engine model 56 takes account of charge-switching processes and thus offers the time dependency, which is particularly important in dynamic processes. Since the nonuniformity of mass and gas forces is added to the mean torque, there is a high degree of flexibility in the parameterization.

Claims (9)

Method for testing a drive train ( 12 ) of a motor vehicle on a test bench ( 10 ), whereby an output ( 21 . 23 . 25 . 27 ) of the drive train ( 12 ) with at least one first electrical machine ( 28 . 30 . 32 . 34 ), which is controlled so as to be at the output ( 21 . 23 . 25 . 27 ) generates predetermined load torques, wherein setpoint values of the load torques are generated by a first model ( 42 ) as a function of predetermined properties of the motor vehicle and a driving distance of the motor vehicle are predetermined, wherein a drive ( 38 ) of the drive train ( 12 ) with a second electric machine ( 36 ) coupled to the drive ( 38 ) Generates driving torques and which is controlled so that the generated Drive torques correspond to setpoint values (SW_M_A) which are determined by a second model ( 44 ), which can be generated by an internal combustion engine, in the drive train ( 12 ) simulates effective torque, characterized in that the second electric machine ( 36 ) to a secondary mass ( 50 ) of a dual mass flywheel ( 46 ) and that of the second model ( 44 ) correspond to a torque (M_ZMS) that is dependent on an elastic coupling (SW_M_A) 52 ) of the secondary mass ( 50 ) with a primary mass ( 48 ) of the dual mass flywheel ( 46 ) on the secondary mass ( 50 ) would be exercised. Method according to claim 1, characterized in that that of the second model ( 44 ) generated setpoints (SW_M_A) taking into account a sum of an average torque (M_DC) of the internal combustion engine and an AC component (M_AC) to the torque (M_VM) of the internal combustion engine. Method according to claim 2, characterized in that as input variables for the second model ( 44 a speed (n) of the internal combustion engine, a torque request (FW) by a driver and a torque request by functions of the powertrain are used. Method according to one of the preceding claims, characterized in that the setpoint values of the load torques by the first model ( 42 ) are given as a function of predetermined properties of the motor vehicle and a driving distance of the motor vehicle, taking into account a tire slip. A method according to claim 1, characterized in that in the generation of the desired values by the second model ( 44 ) is modeled in an intermediate step, the value of a torque (M_VM), the engine from the internal combustion engine to a primary mass ( 48 ) of a dual mass flywheel ( 46 ) would be exercised. A method according to claim 5, characterized in that in an intermediate step, a speed (n) is modeled, with the primary mass ( 48 ) of the dual mass flywheel ( 46 ) would rotate when driven by the internal combustion engine. Method according to Claim 6, characterized in that the modeled rotational speed (n) is used as the input value for an internal combustion engine model ( 56 ) and additionally, together with a measured value for a rotational speed (n_ZMS) of the secondary mass ( 50 ), serve as an input quantity for the generation of the setpoint values of the torque (M_ZMS), which in the elastic coupling ( 52 ) of the secondary mass ( 50 ) with the primary mass ( 48 ) of the dual mass flywheel ( 46 ) would be effective. Test bench ( 10 ) for testing a drive train ( 12 ) of a motor vehicle, with a first electrical machine ( 28 . 30 . 32 . 34 ) and with a controller ( 40 ), which is adapted to the first electric machine ( 28 . 30 . 32 . 34 ) so that they have predetermined load torques at an output ( 21 . 23 . 25 . 27 ) of a powertrain to be tested ( 12 ), the controller ( 40 ) is further adapted to set values of the load torques by a first model ( 42 ) as a function of predetermined properties of the motor vehicle and a driving distance of the motor vehicle, the test bench ( 10 ) a second electric machine ( 36 ) provided with a drive ( 38 ) of the drive train ( 12 ), and the controller ( 40 ) is adapted to the second electric machine ( 36 ) such that the generated drive torques correspond to setpoint values (SW_M_A) which are determined by a second model ( 44 ), which can be generated by an internal combustion engine, in the drive train ( 12 ) simulates effective torque, characterized in that the second electrical machine ( 36 ) to a secondary mass ( 50 ) of a dual mass flywheel ( 46 ) and the test bench ( 10 ) is set up with the second model ( 44 ) To generate setpoint values (SW_M_A) corresponding to a torque that is dependent on an elastic coupling ( 52 ) of the secondary mass ( 50 ) with a primary mass ( 48 ) of the dual mass flywheel ( 46 ) on the secondary mass ( 50 ) would be exercised. Test bench ( 10 ) according to claim 8, characterized in that the test bench ( 10 ) is adapted to perform a method according to one of claims 2 to 7.

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