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    • 专利摘要:
    In order to improve the identification of system parameters of a test setup of a test stand, in particular with regard to the quality of the identification, it is provided that the test setup (PA) is excited dynamically on the test stand (1) by providing the test setup (PA) with a dynamic input signal (u (u). t)) and thereby measured values (MW) of the input signal (u (t)) of the test setup (PA) and a resulting output signal (y (t)) of the test setup (PA) are detected, from the detected input signal (u ( t)) and output signal (y (t)) with a non-parametric identification method a frequency response (G (Ωk)) of the dynamic behavior of the test setup (PA) between the output signal (y (t)) and the input signal (u (t) ), from the frequency response (G (Ωk)) a model structure of a parametric model, which maps the input signal (u (t)) to the output signal (y (t)), is determined on the basis of the model structure and a parametric identification method e at least one system parameter (SP) of the test setup (PA) is determined and the at least one identified system parameter (SP) is used to carry out the test run.
    公开号:AT520554A4
    申请号:T51086/2017
    申请日:2017-12-29
    公开日:2019-05-15
    发明作者:Sangili Vadamalu Raja；Christian Beidl Dr；Bier Maximilian
    申请人:Avl List Gmbh；
    IPC主号:
    专利说明:

    Summary
    In order to improve the identification of system parameters of a test setup of a test bench, in particular with regard to the quality of the identification, it is provided that the test setup (PA) on the test bench (1) is excited dynamically by providing the test setup (PA) with a dynamic input signal (u ( t)) is switched on and measurement values (MW) of the input signal (u (t)) of the test setup (PA) and a resulting output signal (y (t)) of the test setup (PA) are recorded from the recorded input signal (u ( t)) and output signal (y (t)) with a non-parametric identification method a frequency response (G (Q k )) of the dynamic behavior of the test setup (PA) between the output signal (y (t)) and the input signal (u (t )) is determined from the frequency response (G (Q k )), a model structure of a parametric model that maps the input signal (u (t)) to the output signal (y (t)) is derived, based on the model structure and a parametric Identifikationsmeth or at least one system parameter (SP) of the test setup (PA) is determined and the at least one identified system parameter (SP) is used to carry out the test run.
    Fig. 1
    /
    AV-3965 AT
    Test bench and method for performing a dynamic test run for a test setup
    The subject invention relates to a test bench and a method for carrying out a dynamic test run for a test setup on a test bench, the test setup comprising at least one torque generator that is mechanically connected to at least one torque sink on the test bench by means of a coupling element, and wherein the torque generator, the coupling element and the torque sink is described with system parameters characterizing the dynamic behavior.
    The development of drive units, such as internal combustion engines or electric motors, of drive trains with such drive units or of drive train components with such drive units, largely take place on test benches. Likewise, the calibration of a control function or regulation function of a vehicle is usually carried out, for example to meet legal requirements, such as the emission behavior on a test bench. To carry out a test on the test bench, the test object, i.e. the drive unit or the drive train or the drive train component, is connected to a load machine (usually an electric motor, also called a dynamometer) on the test bench in order to be able to operate the test object against a load , The test object and load machine are usually coupled using coupling elements such as test bench shafts, coupling flanges, etc. dynamic behavior of the system responds. Of course, suggestions with a natural frequency of the dynamic system are critical on the test bench, since this can create critical states that can even damage or destroy certain parts, in particular the coupling elements, of the test setup on the test bench. Knowledge of the dynamic behavior of the test setup is therefore important for carrying out test tests on a test bench.
    However, controllers are also used on the test bench to control components of the test setup, in particular the loading machine and a drive unit, for carrying out the test. For the design of the controller, precise knowledge of the dynamic behavior of the test setup is also desirable in order to be able to adapt the controller behavior to it and / or to ensure the stability of the control system.
    Last but not least, so-called observers are often used on a test bench to calculate non-directly measured or directly measurable quantities of the test set-up from other accessible or available measurement quantities. An example of this is the inner / 2 (5 1
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    Torque of the drive unit, i.e. the torque actually generated and not the torque that is often required or used on the test bench for the test run.
    For the design of a controller and / or an observer, a model of the dynamic system, i.e. the test setup, is generally required, which requires a sufficient knowledge of the dynamic system.
    The dynamic behavior of the test set-up on the test bench is primarily determined by the inertia of the components of the test set-up (in particular the test specimen and the load machine) and by the stiffness, possibly also the damping, the coupling between the test specimen and the load machine (i.e. between the mass-loaded components of the test set-up), for example, the torsional rigidity of a test bench shaft. These parameters are often determined individually for each component or are known from data sheets for the respective component. In practice, however, the use of these known parameters is often unsatisfactory when carrying out the test run and has often led to poor results. The reason for this is that the test rig often adapts to the mechanical structure of the test structure, which changes the dynamic system. For example, other measuring sensors, for example a torque sensor on the test bench shaft, are used, or mechanical components are exchanged or added or removed on the test bench. For example, an adapter flange between two components of the test setup can be changed. The properties of components in the test setup can also change due to aging, which also affects the behavior of the dynamic system.
    From DE 10 2006 025 878 A1 it is therefore already known to determine the parameters of the dynamic behavior directly on the test bench. For this purpose, the test setup is dynamically excited by a pseudo-stochastic speed excitation and the parameters of a model of the dynamic system, in particular stiffness and damping of a connecting shaft, are determined using methods of identification theory. With the identified parameters of the model of the test setup, the behavior of the test setup can be described with sufficient accuracy and used for the design of an observer or a controller, but also for monitoring the system. In this approach, a parametric model of the test setup is used, i.e. a model that contains the parameters of the dynamic system and that maps the input / output behavior of the dynamic system. The parameters are determined as the poles of the dynamic system. A difficulty of this method is that the resulting torques have to be measured as output variables of the dynamic system due to the speed excitation, which is difficult in practice on the test bench. Apart from this, certain assumptions about the model structure must be made in advance in order to be able to determine the parameters of an assumed model. Becomes an unsuitable model3 / 26
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    If the structure is selected, the real dynamic behavior is only partially or inaccurately reproduced by the model. In practice, however, the correct choice of the model structure is a difficult task, especially in the case of more complex test setups with several masses and couplings in between, and can only be done by specialists, which limits the applicability of the method. In addition, no noise of a measurement signal (e.g. a measured speed) is taken into account during identification, which can lead to a poorer identification result. In addition, identification in DE 10 2006 025 878 A1 takes place in an open loop, although the measurement signals on the test bench were measured in a closed control loop. This can also reduce the quality of identification. Last but not least, pseudo-stochastic speed excitation cannot be set to the desired or required frequency range. This can lead to certain frequencies not being excited at all or more than required frequencies being excited, which can also have a negative effect on the quality of identification.
    It is therefore an object of the present invention to improve the identification of system parameters of a test setup of a test bench, particularly with regard to the quality of the identification.
    This object is achieved according to the invention in that the test set-up is dynamically excited on the test bench, in that a dynamic input signal is applied to the test set-up and measurement values of the input signal of the test set-up and a resultant output signal of the test set-up are recorded from the recorded input signal and output signal with a not -parametric identification method, a frequency response of the dynamic behavior of the test setup between the output signal and the input signal is determined, from the frequency response a model structure of a parametric model that maps the input signal to the output signal is derived, based on the model structure and a parametric identification method, at least one system parameter of the Test setup is determined, and the at least one identified system parameter is used to carry out the test run. This enables a systematic identification of the required system parameters, whereby a basic characterization of the dynamic behavior of the test setup is carried out, from which the model structure on which the test setup is based can be derived. With the non-parametric identification, a suitable choice of the model structure can be ensured regardless of the complexity of the test setup. The following parametric identification then uses the knowledge of the model structure to determine the system parameters. An additional advantage can be seen in the fact that the non-parametric identification as well as the parametric identification uses the same measured variables, which facilitates the implementation of the identification method.
    / 26 3 '
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    With the non-parametric identification method, measurement noise of the input signal and / or the output signal can also be taken into account, as a result of which the quality of identification can be improved. In addition, the non-parametric identification method can also be used to determine a variance of the measurement noise of the output signal and / or a variance of the measurement noise of the input signal and / or a covariance of the noise between input and output. These variances are then also available for use in the parametric identification process.
    Model parameters of the parametric model are advantageously determined with the parametric identification method and from this by comparing the parametric model with a physical system model with the at least one system parameter, which determines at least one system parameter of a system component of the test setup. This can be facilitated if the parametric model is broken down into partial models, each with model parameters, and by comparing at least one partial model with a physical partial model with the at least one system parameter, which at least one system parameter is determined from the model parameters of the partial model.
    The at least one identified system parameter can be used to carry out the test run by using the at least one system parameter for designing a controller for at least one component of the test setup. Or by using the at least one system parameter for a design of a filter that either filters a setpoint for a controller for at least one component of the test setup or for a control deviation supplied to a controller for at least one component of the test setup. Or by monitoring a change in the at least one system parameter over time. Or by using the at least one system parameter in order to adapt the dynamic behavior of the test setup to a desired dynamic behavior.
    The subject invention is explained in more detail below with reference to FIGS. 1 to 6, which show exemplary, schematic and non-limiting advantageous embodiments of the invention. It shows
    1 shows a test stand with a dual-mass oscillator as a test setup, Fig. 2a, 2b examples of frequency responses of a dual-mass oscillator, Fig. 3 shows a controller for regulating a component of the test setup, Fig. 4 shows the use of a filter to carry out a test run on the test bench, Fig. 5 shows an example of a three-mass oscillator and FIG. 6 shows an implementation of the invention on the test bench.
    To carry out a test run, the invention is based on a test setup PA on a test bench 1, with a test specimen with a torque generator DE, for example
    -4 / 2 (5th
    AV-3965 AT a drive unit such as an internal combustion engine 2, and an associated torque sink DS, for example a load machine 3 (dynamometer), as a load, as shown in a simple case in Fig.1. The torque generator DE is, for example, an internal combustion engine 2, but can also be a drive train with an internal combustion engine 2 and / or also an electric motor, or a part thereof. The test object comprises at least one torque generator DE. Torque generator DE and torque sink DS are mechanically connected to one another via at least one coupling element KE, for example a test bench shaft 4, for transmitting a torque. The coupling element 4 can also comprise a plurality of non-negligible mass-bearing mechanical components, for example coupling flanges, gears, etc. This means that the test setup PA can become arbitrarily complex with regard to the components that influence the dynamics. The dynamic behavior of the test setup PA is determined in a known manner primarily by the inertia J of the components of the test setup PA (i.e. by the non-negligible, mass-laden parts) and the type of coupling (stiffness c, damping d) between them. To carry out a test with such a test setup PA, it is important to know the dynamic system parameters in order to know the dynamic behavior of the test setup PA. For example, knowledge of the resonance frequencies ω κ of the test setup PA is important in order to avoid excitation in the range of the resonance frequencies ω κ . Apart from this, the test run on test bench 1 is intended to simulate the behavior of the test specimen that would result if the test specimen was installed in a real vehicle and moved with the real vehicle. It is therefore important in this context that the dynamic behavior of the test object on test bench 1 roughly corresponds to the dynamic behavior in the vehicle in order to enable realistic test runs. If the dynamic system parameters of the test setup PA are known, then specific measures can be taken on the test bench 1, for example mechanical measures such as additional or different masses, stiffness and / or damping, or control engineering measures, such as adding filters and / or controllers, in order to do this adapt dynamic behavior on test bench 1 to real behavior. The system parameters are also required for the design of possible controllers for controlling certain components of the test setup PA, for example a dyno controller R D for the loading machine 3, in order to be able to optimally adapt the controller behavior to the specific dynamic behavior of the test setup PA.
    The invention is based on the fact that dynamic system parameters describing the dynamic behavior, in particular moments of inertia J, torsional stiffness c, torsional damping d, resonance frequencies ω κ or eradication frequencies ω-τ, of the concrete test setup PA, are, at least in part, not known and before a test test is carried out to be determined on test bench 1.
    / 26
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    For this purpose, according to the invention, the basic character of the dynamic behavior of the test setup PA is first determined using a non-parametric identification method, from which a model structure of a model of the test setup PA with system parameters SP of the test setup PA describing the dynamic behavior is derived. The system parameters SP of the model are then determined using the model structure with the system parameters SP using a parametric identification method. With nonparametric identification, only the measured input signals u (t) and measured output signals y (t) are examined.
    The frequency response (characterized by the amplitude, and possibly also the phase, by frequency) of the test setup PA is determined using the non-parametric identification method. For the frequency response, the physical dynamic system (here the test setup) is known to be excited with a dynamic signal u (t) (input signal) with a certain frequency content and the response y (t) (output signal) is measured on the test setup PA. The input signal u (t) is, for example, a speed ω D of the torque sink DS (for example load machine 3) and the output signal y (t) is, for example, a shaft torque T sh on the test stand shaft 4 or a speed ω E of the torque generator DE. Typically, measuring sensors MS are also provided on a test bench 1 in order to detect the measured values MW of certain measured variables (input signal u (t), output signal y (t)) (FIG. 6), for example a speed sensor 5 for detecting the speed ω 0 of the loading machine 3 and / or a torque sensor 6 for detecting a shaft torque T sh , as shown in Fig.1. Other measured variables can also be measured, such as the rotational speed ω E or the torque T E generated by a torque generator DE (eg internal combustion engine 2) or the torque T D of the torque sink DS.
    However, it is not important what is used as the input signal u (t) and what is used as the output signal y (t). The methodology described below is independent of this.
    The frequency response is determined in a known manner from the Fourier transform of the input signal u (t) and the output signal y (t). If U (k) denotes the Fourier transform of the input signal u (t) at frequency k = ^ k and Y (k) denotes the Fourier transform of the output signal y (t) at frequency k = ^ k , then the frequency response G results (k) as the quotient of the Fourier transforms Y (k) of the output signal y (t) and U (k) of the input signal u (t). It is also known to take into account noise of the input signal u (t) and the output signal y (t). Noise results, for example, from measurement noise when measuring physical quantities, from deviations between a setpoint specification on the test bench and the adjustment of the setpoint at the test bench, through process noise, etc. Denotes n u (t) the noise at the input and n y (t) the noise at the output, the input signal u (t) in the time domain as 7/26 6
    AV-3965 AT u (t) = u 0 (t) + n u (t) or in the frequency domain as U (k) = U 0 (k) + N u (k) and the output signal y (t) im Time domain also as y (t) = y 0 (t) + n y (t) or in the frequency domain as Y (k) = Y 0 (k) + N y (k), with the noise-free signals u 0 , y o resp U 0 , Y 0 and the noise signals n u , n y and N u , N y .
    In order to approximate the frequency response G in the presence of input and output noise, there are various known non-parametric identification methods, for example spectral analysis or the local polynomial method (LPM). Spectral analysis evaluates either the amplitude spectrum or the power spectrum of the frequency response, e.g. in L. Ljung, "System Identification: Theory for the User", 2nd Edition Prentice Hall PTR, 1999 or in Thomas Kuttner, "Praxiswissen Schwungungsmesstechnik", p.325-335, Springer Vieweg 2015. The method of local polynomials is described, for example, in R. Pintelon, et al., “Estimation of non-paramteric noise and FRF models for multivariable systems - Part I: Theory”, Mechanical Systems and Signal Processing, volume 24, Issue 3, p. 573-595, 2010. The determination of the frequency response G is briefly explained below using the example of LPM.
    In LPM, the frequency response G is locally approximated by a local frequency k by a polynomial. This is done for all frequencies jω k of frequency response G. If a generalized frequency Q k is used, with Q k = jω k for the time-continuous case and Q k = e j “ k for the time-discrete case, the input-output behavior of the dynamic system (test setup PA) can be in the form
    Y (k) = G (Q k ) U (k) + T (Q k ) + V (k) can be written. G (Q k ) denotes the Fourier transform of the transfer function of the dynamic system (i.e. the frequency response between the selected input and output), T (Q k ) a transient error in the output at frequency Q k , which is not due to the excitation, and V (k) the measurement noise of the output signal. U (k) and Y (k) are the Fourier transforms of the measured input signal u (t) and output signal y (t).
    The k of the frequency-dependent quantities Q k can be approximated by local polynomials a local frequency Q. The frequencies around Q k are indicated by the variable r = -n, -n + 1, ..., n, where n is specified or selected. This leads to / 26 7
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    Ο (Ω k „). G (Ok) + Σ g, (k) r 's = 1
    T (Qk "). T (Qk) + f <s (k) r s s = 1
    Y (k + r) G (Ok) + Σ 8s (k) r s s = 1
    U (k + r) + Τ (Ω ( ) + Σts (k) r s + V (k + 1) s = 1
    In it, g s and t s denote the 2 (R + 1) unknown parameters of the local polynomials of the order R (which is selected or specified). This gives a total of 2n + 1 equations for 2 (R + 1) unknowns (g s , t s ) based on the r for each frequency k. With a parameter vector 0 (k) = [θτ (Ω) g (k) g 2 (k) ··· g R (k) Τ (Ω) t (k) t 2 (k) ··· t R (k )] can the 2n +
    Equations in the matrix form Y (k) - O (k) 0 (k) are written using the matrix
    r R ] U (k + r) 1

    In it are the resulting 2n + 1
    Equations for the frequencies k are stacked on top of each other. The advantage of this method is that the transient component T (Q k ) can be estimated directly and does not have to be determined for certain frequency ranges by window approaches, such as in spectral analysis.
    The parameter vector 0 (k) can then, for example, in the sense of a least squares approximation from a parameter estimate using the equation 0 (k) = | Φ (Ε) Ν Φ / k) 'O (k) H Y (k) are estimated, where “() H ” denotes the adjoint matrix (transposed and complex conjugate).
    With the resulting residual β (Ω Ε + 1 ) = Y (k + r) [G (Q) U (k + r) + Τ (Ω)] the least squares approximation, the variance σ 2 (k) of the measurement noise of the output signal with σY (k) =
    2n +1 - R
    Σ | e ^ k +,) | 2 can be calculated.
    This approximation then also provides direct estimates for the frequency response Οτ (Ω) and also for the transient components Τ (Ω). With “ Λ ”, the following are always estimated / 26 8
    AV-3965 AT tongues. Depending on what is used as the input signal U (k) and as the output signal Y (k), there are of course different frequency responses G (Q).
    The case of additionally noisy input signals u (t) or the case of feedback of the noisy output signal y (t) to the input (for example in the case of a closed control loop), which also leads to a noisy input signal u (t), can also be covered. In order to avoid systematic errors in the parameter estimation in the presence of input noise, the parameter estimation in this case is advantageously carried out on the closed control loop. This is not a significant limitation, since technical systems such as a PA test setup on a test bench 1 are usually operated in a closed control loop. This is based on a reference signal s (t) (or the Fourier transform S (k)) which corresponds to the setpoint specification for a closed control loop. The relationship between the input signal u (t) and the reference signal s (t) results from the transfer function R of the controller and the current actual values y ist zu u = (s - y ist ) -R. The frequency response G ru (O k ) of the reference signal to the input U (k) and the frequency response G ry (Q k ) of the reference signal to the output Y (k) are then determined, as described above, which leads to estimated frequency responses G ru (Q k ) and (3 ry (Q k ) leads with G rz (Q k ) = [G ry (Q k ) G ru (Q k )] T. With VY (k) the measurement noise of the output signal Y (k) VU (k) denotes the measurement noise of the input signal U (k), with Z (k) = [Y (k) U (k)] T and VZ (k) = [V Y (k) V U (k) ] T the system equation can then be written in the form Z (k) = G. (Q) R (k) + Trz (Qk) + Vz (k), where Trz (Qk) = [T ^ (Qk) 1, (^)] T denotes the transient system error on the input and the output. An estimate of the frequency response G (Q) in the presence of input noise and output noise is then obtained from G (Q) = G. (Q) Gm '(Q k ) ,
    Analogous to the case with only output noise, a variance σ 2 (k) of the measurement noise of the output signal, a variance σ ^^) of the measurement noise of the input signal and a covariance σ ^ / k) of the noise between input and output can also be found in the case of input noise and output noise be determined.
    For the excitation of the dynamic system, the torque generator DE, like the internal combustion engine 2, is preferably towed, that is to say not fired. Actually operated torque generator DE, such as a fired internal combustion engine 2, would also be possible with excitation, but this would make identification more complex because the torque generator DE itself would introduce speed oscillations. For the excitation, therefore, the / 26 connected to the torque generator DE is preferably used
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    Torque sink DS, the load machine 3, which can also drive by design as an electric motor, is used and speed vibrations are impressed for excitation. The excitation can take place with various signals, such as, for example, a multisine signal (in which several frequencies are excited simultaneously at any time) or a chirp signal (in which a frequency is excited at any time, for example with a linear frequency increase). The pickup signal is the setpoint specification for test bench 1 for pickup operation. For example, speed setpoint specifications for the controller of the loading machine 3 are specified as the excitation signal.
    Some important properties of the dynamic system can be derived from a frequency response G (Q k ), as shown in FIGS. 2a and 2b, in which the amplitude A = | G (Q k ) | of the frequency response G (Q k ) is shown, is explained as an example. The frequency response G (Q k ) with the torque T D of the loading machine 3 as input u (t) and the speed w D of the loading machine 3 as output y (t) are shown in FIG. 2a, and the frequency response G in FIG. 2a (Q k ) with the torque T D of the load machine 3 as the input u (t) and the speed ω E of the internal combustion engine 2 as the output y (t). This means that upon excitation, the input u (t) and the output u (t) are measured and, as described, the frequency response G (Q k ) is determined from their Fourier transforms U (k), Y (k). Of course, different input signal / output signal combinations can be used, for example depending on which measured variables are available.
    For example, resonance frequencies ω κ and / or repayment frequencies ω Ε can be derived from the frequency response G (Q k ). Both frequencies can be determined by determining minima and maxima and the gradients in the frequency response G (Q k ). A repayment frequency ω-τ is therefore a minimum with a gradient reversal from negative to positive. A resonance frequency ω κ a maximum with gradient reversal from positive to negative. Of course, several or no eradication frequencies ω τ and / or resonance frequencies ω κ can occur in the frequency response G (Q k ). Furthermore, information on the zeros can also be derived from the frequency response G (Q k ). In principle, characteristic frequency responses are known for different systems, for example for a two-mass oscillation system (Fig. 2a, 2b) or a multi-mass oscillator. A specific model structure can be concluded by comparing the known frequency responses with the determined frequency response G (Q k ). After the frequency response G (Q k ) can be represented in the form of numerator polynomial to denominator polynomial, the comparison of the numerator and denominator polynomial (which corresponds to the model structure) can be determined on the basis of this comparison. The number of resonance frequencies ω κ can also be used to draw conclusions about the model structure. The number of resonance frequencies ω κ usually corresponds to the number of vibratable masses minus one. A two-mass oscillation system accordingly has a resonance frequency ω κ (as for example in FIGS. 2a, 2b).
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    Different characteristic frequency responses G (Q k ) result for different system configurations (with oscillating masses and mechanical couplings in between), which are however known. The characteristic frequency responses G (Q k ) can be stored, for example, in order to be able to infer a model structure of the present test setup PA by comparing the estimated frequency response G (Q).
    2a and 2b also show the respective determined covariances c ^ j (k).
    The advantage of this procedure can also be seen in the fact that one can get by for the nonparametric identification under certain circumstances even without measuring a torque on the test bench 1, for example to select a model structure. This enables a first estimate of the dynamic behavior of the test setup PA on the test bench 1 in a simple manner, in particular without including a torque measuring flange in the test setup PA.
    According to the invention, the non-parametric identification method is followed by a parametric identification method with which the dynamic input-output behavior of the dynamic technical system (test setup PA) is approximated with a model with system parameters SP that describe the dynamic behavior of the test setup PA. There are also known methods for this, both in the time domain and in the frequency domain, which are briefly explained below.
    The parametric identification is based on a model of the dynamic system with the model parameters θ, which calculates the output y (k) from the input u (k) and a fault. The mapping of the input u (k) to the output y (k) takes place with a system information model G (q, Θ) = g e q ', with system model parameters gθ and the backward shift k = 1 operator q -k . A disturbance (e.g. due to measurement noise) can be modeled with a disturbance model H (q, θ) and a probability distribution e, or a probability density function f e , inf, with H (q, Θ) = 1 + h e q 'and noise model parameters ho , It should be noted that k = 1 here that k does not denote a frequency as in non-parametric identification, but rather a time index of the discrete-time signals, eg u (k) and y (k). The model of the dynamic system can then be written in a discrete-time notation as y (k) = G (q, Θ) π ^) + H (q, Θ) β ^). The goal is therefore to use this model to estimate the output y (k) at the time k from known past data of the input u and the output y up to the time k-1 (ie past data). Ver-1112/26
    AV-3965 AT are the data Z K = {u (1), y (1), u (k-1), y (k-1), u (k), y (k)}. There are various known approaches for this.
    An example of a parametric identification method in the time domain is the so-called Prediction Error Method (PEM), as described for example in L. Ljung, "System Identification: Theory for the User", 2nd Edition Prentice Hall PTR, 1999. A known method in the frequency domain is the Maximum Likelihood Estimator Method (MLE).
    PEM is based on the model of the dynamic system y (k) = G (q, 0) u (k) + H (q, 0) e (k)
    K___________J v (k), where θ are the model parameters. Here v (k) denotes colored noise. If white smoking is used as the probability distribution e (k), then the colored noise v (k) can also be described as v (k) = H (q, 0) e (k) = h (t) e (k -1) = e (k) + m (k -1) can be written t = 0. Here m (k-1) is the mean up to time (k-1). This can be done in the form m (k-1) = v (k | k-1) = (H (q, Θ) - 1) e (k) = (1 -H mv (q, 0)) v ( k) can be rewritten with the inverse H inv of H. The estimate of the output y (k | k-1) from the data Z K can then be written in the following form:
    y (k 1 0 , Z k ) = G (q, 0) u (k) + v (k | k -1) = H inv ( q, 0) G ( q, 0) u (k) + (1 - H 1IW (q, 0)) y (k) .
    The estimation error then results in s (k, 0) = y (k) - y (k, 0). In order to determine the model parameters θ, a cost function ΰ (θ, Z K ) can be written that minimizes the weighted estimation error. For example, a mathematical norm l (), for example the Euclidean norm (2 norm) l () = 11 () | 2 , the weighted estimation error can be used. If s F (k, 0) = F (q) s (k, 0) denotes the weighted estimation error with the weights F (q), k- 1, then the cost function can be, for example, as J (0, Z) = - l (see F (k, 0)) can be written K k = 0. This cost function J is minimized to estimate the model parameters θ: 0 = argmin J (0, Z).
    In order to determine the sought system parameters SP of the dynamic system from the model parameters θ, the discrete-time model y (k) = G (q, 0) u (k) + H (q, 0) e (k) in an advantageous practical implementation with polynomials A (q) = 1 + ajq -1 + ... + a na q -na ,
    B (q) = 1 + bq -1 + ... + bnbq -nb and C (q) = 1 + qq -1 + ... + encq -nc also in the form y (k) = B (q) u (k) + C (q) e (k) (ARMAX model). The model parameters A (q) A (q)
    G H / 26
    AV-3965 AT
    Then we get θ = [α ... a na b C c n J · In the case of white ones
    Noise e (k) is reduced to y (k) = B (q) u (k) h - 1 - e (k) (ARX model)
    A (q) A (q)
    GH θ = [α ... a na b bJ. The order n a , n b , n c of the polynomials A, B, C is specified according to the model structure defined with the non-parametric identification.
    In the case of a dual-mass oscillator, for example in a test setup according to FIG. 1, the denominator polynomial A (q) is a polynomial with order n a = 2 and the numerator polynomial B (q) is a polynomial with order n b = 1.
    In order to arrive at the sought system parameters SP, an equivalence of the model parameters in the discrete-time model and the parameters of a physical model of the dynamic system is assumed.
    For example, the test setup PA according to Fig. 1 with internal combustion engine 2 (with inertia J E ), test stand shaft 4 (with torsional rigidity c and torsional damping d) and load machine 3 (with moment of inertia J D ) physically according to ω ο _ 1 J E s 2 + ds + c T D s JEJD s + d ( JE + J D ) s + c (J E + J d ) ω _ 1 ds + c T D s J E J D s H d (J E HJ D ) s HC ( J E HJ D ) can be modeled. Another example arises in the two-mass oscillator of FIG. 1 with the input signal ω 0 and the output signal ω ^, which results in a model ω Ε c + ds
    J E s + ds + c leads. In this case, the frequency response G would of course also have been determined with this input signal and output signal. This also shows that different models 20 also result from different input and / or output signals. These equations can be put into a discrete-time notation, which enables a comparison of the system parameters SP (J E , c, d, J D ) with the estimated model parameters θ = [^ ... a na b ··· b n J. The system parameters SP (J E , c, d, JD) can be determined from this.
    After the model structure is known from the frequency response G, the model can also be divided into 25 partial models, which makes it easier to determine the system parameters SP (J E , c, d, J D ). A two-mass oscillator can be subdivided, for example, into a first sub-model for the internal combustion engine 2, a second sub-model for the test bench shaft 4 and a third sub-model for the load machine 3. This means that the discrete-time model can also be subdivided into corresponding sub-models, which results in the sub-models being ordered
    AV-3965 AT reduced speaking. The model parameters of the sub-models are then estimated as described above.
    Physical sub-models are then used in the same way, which is described again below using the example of the dual-mass oscillator.
    For the first sub-model, the moment equilibrium for the internal combustion engine 2 (FIG. 1) is written in time-continuous notation in the form ω Ε = - (T E (s) - T sh (s)), J E s with the Laplace operator s , the speed ω E of the internal combustion engine 2, the internal torque T E of the internal combustion engine 2 and the shaft torque T sh . After the internal combustion engine 2 is preferably operated in a towed manner, the non-stochastic part of the model structure results in T E = 0 and consequently ω Ε = —— T h (s) or J E s ω Ε (k) ω (k-1) = - T (k) in time-discrete notation with the known sampling time Ts (ty J E typically in the kHz range). By comparison, a 1 = -1 and b 1 = Ts / J E are then obtained directly from the first partial model, from which the system parameter J E can be determined. This also enables the quality of the parameter estimation to be estimated. If the estimated model parameter a 1 is close to one, then a high identification quality can be assumed.
    In order to determine the system parameters SP of the coupling between the internal combustion engine 2 and the load machine 3, torsional damping d and torsional stiffness c, the second sub-model for the test bench shaft 4 is used. Starting from the torque equilibrium on the cut-out test stand shaft 4, the load machine 3 can be written with a Δω = ω E - ω 0 , with the speed ω 0 and again with the assumption T E = 0 c
    T sh (s) = (--- hd). o (s) or in time-discrete notation s
    T (k) - T sh (k -1) = (-cTs + d). O (k) - d. O (k -1). By comparison, the system parameters a 1 = -1, b 1 = - (cTs + d) and b 2 = -d again result from the model parameters of the second partial model, from which the sought system parameters c, d can be determined again. The model parameters a 1 , b 1 of the second partial model naturally do not correspond to the model parameters of the first partial model.
    In the same way, a third sub-model can be used for the loading machine 3 in order to determine the system parameter J D. The inertia J D of the loading machine can of course also be determined for the dual-mass oscillator from the resonance frequency ω κ , which is known from the non-parametric identification, for example, -14- / 26
    AV-3965 AT
    as J D = c = -------. The inertia J D of the loading machine 3 is often known, 2 C - R J J E
    so that the known quality of inertia can also be used to compare the identification quality.
    As already mentioned, the parametric identification can also take place in the frequency domain, for example using MLE. At MLE the model parameters θ are estimated, which maximize the so-called likelihood function. This known method is briefly explained below.
    MLE uses the measurement data of the output signal y = y 1 , y 2 , y N and an associated probability density function f ny of the measurement noise at the output, which is assumed to be known and is described by the model parameters θ. f (y ^ 0 ) describes the probability distribution function of the randomly dependent parts of the estimation problem. With a hypothetical model y 0 = G (u 0 , θ) that describes the excitations and the parameters, the likelihood function in the case of measurement noise at the output can be written as f (y | 6, u 0 ) = f (yy 0 ) , Here u 0 denotes the noiseless input. The unknown model parameters θ can then be determined by maximizing the likelihood function f: θ (Ν) = argmaxf (y 10, u o ).
    θ
    In the case of additional measurement noise at the input, the likelihood function can be written as f (y, u ^, y 0 , u 0 ) = f (y - y 0 ) f nu (u -u 0 ) with the probability density function fnu of the measurement noise on Entrance.
    Assuming that the noise at the input and the output have zero mean, are normally distributed and are independent of the frequencies, a Gaussian
    Cost function Υ ^ Θ, Ζ) =
    Σ k = 1 | e (Qk, frZ (k)) | 2 σΜ, Θ) (likelihood function f). Therein θ is the parameter vector and Z (k) = [Y (k) U (k)] denotes the available measurement data of the input U (k) and the output Y (k). e denotes an error over all frequencies Q k of the form e (Q, Θ, Ζ (^) = Y (k) - G (Q, Θ) υ ^) and a e the covariance of the
    Error e in the form σ (^, Θ) = σ (^ + | G (^, Θ) | 2 σ (^ - 2Re (G (Q) σ Π (^). Re denotes the real part.
    As can be seen, the variances and covariance of the measurement noise at the input σ ^), Output σ (^ And input-output σ ^, (Κ) / 26
    AV-3965 AT required. These can advantageously be obtained from the non-parametric identification as described above.
    Since the resulting optimization problem of maximizing the cost function V ML (likelihood function) is non-linear, the optimization is solved, for example, with the well-known Levenberg-Marquardt method. The convergence of the optimality of the optimization essentially depends on the initial values of the optimization. Estimated values of the sought system parameters SP can be used as initial values or other known initialization methods, for example from a generalized total least square method, can be used.
    If a parametric transfer function G (O, Θ) = B 9 ( ° k) is used, the
    Α θ (ok)
    Cost function for MLE to be rewritten. Here A and B are again polynomials A (q) = 1 + ajq '+ ... + a na q -na , B (q) = 1 + bq' + ... + bnbq -nb . The order na, n b of the polynomials again results from the frequency response G (O) estimated with the non-parametric identification. The optimization then again results in the model parameters 9 = [a ... a na bb nb ], which in turn are compared from a comparison with a physical model of the dynamic system (test setup PA) in order to refer to the system parameters SP (J E , c, d, J D ) to come. In the same way, partial models can also be used again in order to simplify the determination of the system parameters SP (J E , c, d, J D ).
    The partial models for the internal combustion engine 2 and the test bench shaft 4 in time-discrete NotaTs tion with the z-transformation result in ω Ε (k) = 1 - z -1 T sh (k) and
    Tsh (s) = (d - Ths) - dz 1
    - z -1
    Ao (k), from which the system parameters SP (J E , c, d, J D ) result from the
    T s estimated partial models for J E = -, J D b 1
    C o R - R J J E (b2 + b1)
    L · -
    Ts d = -b s result.
    The parametric identification thus supplies the system parameters SP (J E , c, d, J D in the case of the dual-mass oscillator) of the excited dynamic system (the test setup PA as shown, for example, in FIG. 1). With at least one system parameter SP, for example / 26
    AV-3965 AT a controller for certain components of the test setup can be designed using conventional controller design methods, e.g. a speed controller for the load machine 3 or a torque controller for the internal combustion engine 2.
    An exemplary closed control loop for the test setup PA according to FIG. 1 is shown in FIG. The controlled system 10 represents the test setup PA with the mass moments of inertia J E of the internal combustion engine 2 and J D of the load machine 3. The test stand shaft 4 is represented by the torsional rigidity c and torsional damping d. These system parameters SP were determined by the above non-parametric / parametric identification. The loading machine 3 is speed-controlled by a Dyno controller R D (for example a PI controller). For this purpose, a target speed w Dset is specified. The actual speed ω 0 that occurs is measured on test bench 1. The control error u Dsoll -ω 0 is corrected by the Dyno controller RD according to the specified controller behavior. The controller parameters for setting the desired control behavior on the specific dynamic system can be determined using a well-known controller design method.
    The resonance frequency (s) ω κ determined with the non-parametric identification, which also represent a system parameter SP, can also be used to design a filter F that is intended to prevent possible resonances on test bench 1. The aim of the filter F is to prevent excitation with a resonance frequency ω κ . In order to influence the dynamic behavior of the test setup PA on a test bench 1 with a filter F as little as possible, filters F are suitable for this purpose, which filter out frequencies within a narrow frequency range, for example so-called notch filters. For this purpose, the notch filter is designed so that frequencies in a narrow frequency range around the resonance frequency ω κ are filtered out. Such a filter F can be used in front of a controller R for a component of the test setup PA, for example the Dyno controller RD, for filtering the control deviation w supplied to the control R (difference between the setpoint SW and the actual value IW of the controlled variable) (as in FIG. 4) , or also for filtering the setpoint values SW (for example a setpoint speed ω ^ ιι) for the controller R (as indicated by the broken line in FIG. 4).
    The identified system parameters SP (ω κ , J E , c, d, J D ), or at least one of them, can also be used for an observer to measure immeasurable quantities of the test set-up, for example an internal effective torque T E of the internal combustion engine 2 , appreciate.
    Likewise, the identified system parameters SP (ω κ , J E , c, d, J D ), or at least one of them, can be used to determine changes in the test setup PA, for example due to aging, damage, configuration changes, etc. For this purpose, the system parameters SP (JE, c, d, JD) can be re-determined at regular intervals and the change over time of the system parameters SP (JE, c, d, JD) can be monitored. Here, -17 · / 26
    AV-3965 AT, the test run can be adapted or interrupted if an unusual or undesirable change is detected.
    Last but not least, the test setup PA for carrying out the test test itself can also be changed in order to change the at least one identified system parameter SP in order to represent a desired dynamic behavior on the test bench 1. This can be used, for example, to adapt the dynamics of the test setup PA on the test bench 1 to the dynamics of a vehicle in which the torque generator DE of the test setup PA is to be used.
    A test stand 1 with a test setup PA, for example with a dual-mass oscillator as in FIG. 1, is indicated in FIG. Measured values MW for the required input signals u (t) and output signals y (t) with dynamic excitation are recorded on the test setup PA with suitable measurement sensors MS, which of course also includes the time-discrete detection of the signals u (k), y (k). The required system parameters SP, for example J E , c, d, J D , of the test setup PA are ascertained from the measured values MW in an evaluation unit 12 (hardware and / or software). At least one identified system parameter SP is used in a test bench control unit 11 (hardware and / or software) for carrying out a test run on the test bench 1 with the test setup PA. The evaluation unit 12 and the test bench control unit 11 can of course also be integrated in a common unit. The test setup control unit 11 excites the test setup PA. The at least one identified system parameter SP can be used on the test stand 1 to carry out the test run as described.
    Although the invention was described above using the example of a two-mass test setup, it is obvious that the invention can also be applied to any multi-mass test setup, for example in the case of a drive train test bench in which the torque generator DE uses a combination of shafts, couplings, shaft couplings and / or gearboxes with a torque sink DS is connected, can be expanded. To carry out a test run, a test object is always mechanically connected to a torque generator DE, for example an internal combustion engine 2, an electric motor, but also a combination of an internal combustion engine 2 and an electric motor with a torque sink DS, for example a load machine 3. The connection is made with at least one coupling element KE, for example with a test bench shaft 4, as shown in FIG. 1 for the dual mass test bench. However, the coupling can also be carried out with a coupling element KE, which comprises various mechanical couplings, which can also include further mass inertia (multi-mass test setup), such as in the case of a drive train as a test specimen on test bench 1.
    , -18 · / 26
    AV-3965 AT
    For example, a three-mass test setup is shown in FIG. This contains three mass inertias J1, J2, J3, which are connected by coupling elements with a torsional stiffness c1, c2 and a rotary damping d1, d2. J1 is for example a torque generator DE, e.g. an internal combustion engine 2, and J2, J3 torque sinks DS, for example 5 a torque measuring flange, a dual mass flywheel or a clutch and a load machine 3. Also a configuration with two torque generators DE, an internal combustion engine 2 and an electric motor, and a torque sink DS, a load machine 3, is of course conceivable. The system parameters SP (J1, J2, J3, c1, c2, d1, d2) can be identified according to the invention in an analogous manner to that described above.
    Depending on which measurement variables are available, the following configurations can be mapped, for example:
    Known parameters Measured signals (u (t), y (t)) Identified system parameters SP - ω1, ω2, ω3, Tsh1, Tsh2 J1, J2, c1, c2 J1, J2 Tsh2, ω2 c1, d1 J1, J2 Tsh2, ω2, T2 c1, d1 - ω2, ω3, Tsh2 c2, d2 - ω1, ω2, ω3, T2, Tsh2 (J1 + J2), c2, d2 J2 ω1, ω2, Tsh2 J1, c1
    / 26
    AV-3965 AT
    权利要求:
    Claims (13)
    [1]
    claims
    1. Method for carrying out a dynamic test run for a test setup (PA) on a test bench (1), the test setup (PA) comprising at least one torque generator (DE), which is mechanically connected to the test bench (1) by means of at least one coupling element (KE) at least one torque sink (DS) is connected, and wherein the torque generator (DE), the coupling element (KE) and the torque sink (DS) are described with the dynamic behavior-characterizing system parameters (SP), characterized in that the test setup (PA) on Test bench (1) is dynamically excited by applying a dynamic input signal (u (t)) to the test setup (PA) and thereby measuring values (MW) of the input signal (u (t)) of the test setup (PA) and a resulting output signal ( y (t)) of the test setup (PA) are detected, that a frequency response (non-parametric identification method) results from the recorded input signal (u (t)) and output signal (y (t)) G (Q k )) of the dynamic behavior of the test setup (PA) between the output signal (y (t)) and the input signal (u (t)) is determined that from the frequency response (G (Q k )) a model structure of a parametric Model that maps the input signal (u (t)) to the output signal (y (t)) is derived from the fact that at least one system parameter (SP) of the test setup (PA) is determined based on the model structure and a parametric identification method, and that at least an identified system parameter (SP) is used to carry out the test run.
    [2]
    2. The method according to claim 1, characterized in that the at least one system parameter (SP) for a design of a controller (R, R D ) is used for at least one component of the test setup (PA).
    [3]
    3. The method according to claim 1, characterized in that the at least one system parameter (SP) is used for a design of a filter (F) that either a setpoint (SW) for a controller (R, R D ) for at least one component of the Filters test setup (PA) or for a control deviation (w) fed to a controller (R, R D ) for at least one component of the test setup (PA).
    [4]
    4. The method according to claim 1, characterized in that a change in the at least one system parameter (SP) is monitored over time.
    [5]
    5. The method according to claim 1, characterized in that the at least one system parameter (SP) is used to adapt the dynamic behavior of the test setup (PA) to a desired dynamic behavior.
    21/26
    AV-3965 AT
    [6]
    6. The method according to any one of claims 1 to 5, characterized in that measurement noise of the input signal (u (t)) and / or of the output signal (y (t)) is taken into account in the non-parametric identification method.
    [7]
    7. The method according to any one of claims 1 to 6, characterized in that the method of local polynomials is used as the non-parametric identification method, the local frequency response (G (Q k )) for all frequencies around a local frequency (Q k ) approximated by a polynomial.
    [8]
    8. The method according to claim 7, characterized in that with the nonparametric identification method additionally a variance (σ 2 (k)) of the measurement noise of the output signal (y (t)) and / or a variance (σ 2 (k)) of the measurement noise of the Input signal (u (t)) and / or a covariance (o ^ j (k)) of the noise between the input and output can be determined.
    [9]
    9. The method according to any one of claims 1 to 8, characterized in that a resonance frequencies (ω κ ) and / or a repayment frequencies (ω Ε ) is determined from the frequency response (G (Q k )) as a system parameter (SP).
    [10]
    10. The method according to any one of claims 1 to 9, characterized in that a parametric identification method is used in the time domain, preferably a prediction error method (PEM), or in the frequency domain, preferably a maximum likelihood estimator method.
    [11]
    11. The method according to any one of claims 1 to 10, characterized in that with the parametric identification method model parameters (θ) of the parametric model are determined and by comparing the parametric model with a physical system model with the at least one system parameter (SP), the at least one System parameters (SP) is determined from the model parameters (θ).
    [12]
    12. The method according to any one of claims 1 to 10, characterized in that the parametric model is broken down into partial models with respective model parameters (θ) and by comparing at least one partial model with a physical partial model with the at least one system parameter (SP), the at least one System parameters (SP) is determined from the model parameters (θ) of the sub-model.
    [13]
    13. Test bench for performing a dynamic test run for a test setup (PA) with at least one torque generator (DE), which is mechanically connected to at least one torque sink (DS) by means of at least one coupling element (KE), and wherein the torque generator (DE), Coupling element (KE) and the torque sink (DS) with the system parameters characterizing the dynamic behavior (SP)
    22/26
    AV-3965 AT, characterized in that a test bench control unit (11) is provided on the test bench (1), which dynamically excites the test setup (PA) on the test bench (1) by specifying an input signal for the test setup (PA), and on Test bench (1) measuring sensors (MS) are provided, which measure measured values (MW) of the input signal (u (t)) and a resultant output signal (y (t)) of the test setup (PA), that an evaluation unit (12) is provided is a frequency response (G (Q k )) of the dynamic behavior of the test setup (PA) between the output signal (y () from the recorded input signal (u (t)) and output signal (y (t)) using a non-parametric identification method. t)) and the input signal (u (t)) and from the frequency response (G (Q k )) a model structure 10 of a parametric model that maps the input signal (u (t)) to the output signal (y (t)) , derives and based on the model structure and a parametric identification method de determines at least one system parameter (SP) of the test setup (PA) and that the test bench control unit (11) uses the at least one identified system parameter (SP) to carry out the test run.
    23/26
    AVL List GmbH
    1.3
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    引用文献:
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    EP1452848A1|2003-02-28|2004-09-01|AVL List GmbH|Method for controlling an engine testing bed|
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    JP4591176B2|2005-04-20|2010-12-01|株式会社明電舎|Engine inertia moment measurement device|
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    US16/958,353| US20210063277A1|2017-12-29|2018-12-28|Test Bench And Method For Performing A Dynamic Test Run For A Test Setup|
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