• 专利附录
  • 专利说明
  • 权利要求
  • 类似技术
  • 同族专利
  • 引用文献
  • 法律状态
  • 优先权
    • 专利摘要:
    Method for regulating and regulating a drive train test bench (1) having a drive machine (2) and an output machine (3, 3 '), wherein a torque applied by the input drive machine (2) is controlled, a control variable (Mx) for the control the torque of the input drive machine (2) for damping oscillations between the input drive machine (2) and the output machine (3, 3 ') as a function of a current speed (nist) of the input drive machine (2) is modified from a default value (Msetpoint) by the reference variable (Mx) is determined from a tire model (10) of a virtual tire.
    公开号:AT515110A4
    申请号:T50011/2014
    申请日:2014-01-09
    公开日:2015-06-15
    发明作者:Robert Dipl Ing Dr Bauer；Bernd Dipl Ing Pressl；Martin Mag Baschnegger；Martin Wipfler
    申请人:Seibt Kristl & Co Gmbh；
    IPC主号:
    专利说明:

    The invention relates to a method for controlling and a control device of a powertrain test stand with a Eintriebsmaschine and an output machine, wherein an applied by the Eintriebsmaschine torque is controlled with an inner loop, and with an outer loop for controlling the reference variable of the inner loop in dependence current speed of Eintriebsmaschine, so that a reference variable for the control of the torque of Eintriebsmaschine for damping vibrations between the input drive machine and the driven machine is modified in response to a current speed of Eintriebsmaschine against a default value.
    In a powertrain test stand, the component to be tested (also referred to below as test specimen) is usually not forcibly connected to the environment as at its later place of use. For example, a powertrain on the test rig is connected via relatively stiff shafts to a drive-in machine and one or more load machines rather than being in contact with the road via tires. This results in the test stand usually weakly damped (and thus strongly pronounced) resonant frequencies that does not find the test specimen at its actual site. If these resonant frequencies are excited by one of the machines, the resulting vibrations can massively influence the test result or even lead to the destruction of the test object and / or the test stand.
    Therefore, measures for damping these resonant frequencies on the test bench are necessary. In the following the known methods are enumerated.
    By using softer, more damped connecting waves, the resonance frequencies can be lowered and, in addition, more attenuated. If the resonance frequencies are no longer excited during normal operation, the problem is thus solved. The disadvantage is the low-pass effect of the soft connecting shaft. In addition, by friction, a significant power in the shaft can be implemented (the shaft is hot and possibly destroyed).
    Apart from such a passive damping methods for active damping are already known in which the output machine an additional torque is applied, which corresponds to a wave attenuation. For example, the differential angular velocity of the shaft, i. the difference between the measured rotational speeds on the driven machine and the input drive machine are used (cf., for example, DE 38 08 524 C2). The disadvantage of using the two current speeds on the output and on the input is that these speeds may only be noisy and with considerable delay, e.g. due to bus transmission time and / or filtering of the signal, are available. In the worst case, the intrusion of such a distorted damping torque of the powertrain test bench can be quite unstable.
    The need for high accuracy and dynamics in the use of the two speeds has already been recognized in EP 1 333 268 A2. There, it is proposed to estimate the differential angular velocity with the aid of the measured shaft torque: The measured shaft torque is differentiated, weighted with a correction factor and applied to the torque setpoint of the output machine as a correction value. However, the differentiation of a measured variable has the disadvantage that the ever-present measurement noise is significantly increased. Although you could filter the differentiated moment with a low-pass filter, but then this method becomes unstable at higher resonance frequencies, whereby the use in practice is severely limited. Furthermore, in large test benches with play and many distributed masses (for example in drive trains with matching gears), only the first, machine-near shaft is damped in this way; the remaining waves remain undamped.
    In addition, in WO 2013/126940 Al a method is shown in which the measured shaft torque is not differentiated, but as such, either directly or after a low-pass filtering the torque setpoint of the output machine is switched. This surprising measure for the first time causes a very good damping, especially at high resonance frequencies. However, if the masses present on the test bench are large or the natural frequency to be damped is small, the achievable damping effect is low.
    Other, relatively more complicated, active damping techniques include alternative means of using measured torque (e.g., DE 102 47 347 A and US 4,468,958 A) or attempting to predict the resonance behavior of the test bench (e.g., AT 010 301 U2 and US 8,006,548 B2).
    In connection with the most realistic possible simulation or simulation of the behavior of a vehicle on a road, it is also already known that
    Load machines of a powertrain test rig, i. the downforce machines using a tire model. Either the rotational speed (see AT 508 031 B1) or the torque (see EP 1 037 030 B1) of the driven machines can be regulated accordingly. The aim of both methods is to load the output of the drive train more realistic and not to dampen the vibrations and in particular any natural frequencies of the specimen on the test bench. Such a realistic load is obviously achieved when the tire model is used to control the output (but not the input).
    It is an object of the present invention to provide a method of the initially mentioned kind or a device for carrying out this method which avoids or avoids the oscillation problems described above in a simple and efficient manner or at least reduces them to an acceptable level. In this case, problems with delayed and offset measured values should be avoided and damping of relatively low-frequency oscillations should succeed. The control should also be insensitive to measurement noise and also achieve the desired attenuation in transient tests - especially in the first examination of a test specimen.
    To achieve the object, the invention provides a method as stated above, which is characterized in that the reference variable for the regulation of the torque of the input drive machine is determined from a tire model of a virtual tire. Correspondingly, the invention provides in the control device as stated at the outset that the outer control loop for controlling the reference variable of the inner control loop as a function of a current speed of Eintriebsmaschine has a tire model of a virtual tire. By the in the regulation of the drive, i. the powertrain test engine driveline, tire model used will not achieve a more realistic drivetrain load; it is only the damping effect of such a control advantageous (also) used on the input. Naturally, the actual rotational speed of the input drive machine (on which the virtual tire is arranged) - or an equivalent measured variable - enters the tire model, but a measurement of several rotational speeds on the drive train is not required, so that problems resulting from the relative time Behavior of several measured values result, can be avoided.
    It is particularly advantageous if the tire model produces a, preferably static, relationship between the reference variable and a slippage of the virtual tire or has such a relationship. The slip of the tire, which is a prerequisite for an energy transfer, and a corresponding slip model are particularly well suited for the present method. With a static relationship between the slip and the command variable, i. the (modified) torque setpoint, influences of other -dynamic - measurands can be completely avoided.
    In a particularly simple and therefore preferred tire model, the reference variable is essentially calculated according to the formula MX = Fz-rd-Yn-D-sin (C-arctan (B * sj) or the outer control loop for determining the reference variable is substantially according to this formula where Fz is a rider force, rdyn is a roll radius, B, C and D are constant tire parameters, and s
    Slip of the virtual tire is. In this model, the leader only shows linear behavior in a small range and the amount of abrupt changes is limited.
    When a virtual lane velocity entering the tire model, in particular the slip of the virtual tire, is determined from an inverse tire model, or when the outer lane velocity determining loop is arranged from an inverse tire model, the control of the input machine may be more advantageous Way (continue) by specifying a desired torque. The inverse tire model can be chosen independently of the tire model. In particular, the inverse tire model does not have to correspond to the mathematical inverse of the tire model, but can be derived, for example, from a simplified tire model.
    In this context, it is advantageous if the virtual roadway speed is determined from the default value for the reference variable and a target rotational speed of the input machine, which target rotational speed is preferably proportional to a desired rotational speed of the output machine. In a corresponding manner, the outer control loop is preferably set up for determining the virtual road speed from a default value for the reference variable and a target rotational speed of the input drive machine. Due to the exclusive use of given static quantities, possible feedbacks of dynamic measured variables - in addition to the current rotational speed of the input drive machine - can be prevented in the tire model.
    The derivation of the inverse tire model is particularly simple if the inverse tire model establishes a linear relationship between the virtual roadway speed and the setpoint rotational speed of the input machine. The linear relationship can be used here, without having to accept disadvantages with respect to the damping effect of the tire model, since the tire model (as opposed to the inverse tire model) is at least partially non-linear
    Can have connection.
    The invention will be explained below with reference to particularly preferred embodiments, to which it should not be limited, and with reference to the drawings. The drawings show in detail:
    Fig. 1 shows schematically a powertrain test stand with differential without adjustment gear;
    Figure 2 schematically shows a powertrain test bench with differential with matching gears on the input and output;
    FIG. 3 shows an input torque curve during a load change on a test bench according to the prior art for comparison; FIG.
    Fig. 4 shows schematically the principle of a control of the drive with virtual tire;
    FIG. 5 shows a friction-value slip characteristic according to a typical tire model; FIG.
    FIG. 6 shows an input torque curve during a load change under the present control method for comparison; FIG. and
    Fig. 7 is a graph of friction coefficient-slip characteristics according to three different tire models in comparison.
    FIGS. 1 and 2 each show a powertrain test bench 1 with a drive-in machine 2 and two output machines 3, 3 ', which are connected to a drive train 4 to be tested. The drive train 4 has a rear differential 5 on. In such a powertrain test bench 1, only parts of a drive train 4 are usually loaded with the aid of input and output machines 3, 3 'in accordance with predetermined rotational speed and torque characteristics. Due to the translation iD of the differential 5 applies to the angular velocities of the machines 2, 3, 3 '
    where ωΕ the angular velocity of the input machine 2, and ωΑι or ωΑ2 the angular velocities of the two
    Output machines 3, 3 'are. Furthermore applies - apart from friction losses in the differential 5 - for the torques
    In order to be able to load the differential 5 with larger torques than the machines 2, 3, 3 'could supply, often also fitting gear 6, 7, 7' as shown in Fig. 2 is used. The input gear 2 associated with the matching gear 6 has a translation iE and the output machines 3, 3 'associated matching gear 7, 7' each have a translation iA on. The matching gears 6, 7, 7 'are used so that the torques on the specimen side, i. on the part of the powertrain 4 to be tested are larger than on the machine side. For the angular velocities ωΕ, ωΑΕ, ωΑ2 of the machines 2, 3, 3 'is therefore valid
    where coGE is the angular velocity on the input-side adjusting gear 6 and coGAi or (oGA2 are the angular velocities on the two output-side adjusting gears 7, 7 '.
    wherein - analogous to the angular velocities - MGE, the torque at the input side fitting gear 6 and MGAi or MGA2 the torques at the two output side matching gears 7, 7 'are. Usually, the speeds nA1, nA2 of the output and the torque ME are given when driving on the test bench 1 for testing. Accordingly, the output machines 3, 3 'are usually speed-controlled and the input machine 2 is operated with torque control. Without further measures, the structure on the test rig 1 oscillates, in particular, with the first natural frequency (inertia of the input machine 2 via stiffnesses of the shaft connections against inertia of the driven machines 3, 3 '). This problem comes with
    Matching gears 6, 7, 7 'still exacerbated, as the examinee 4 through the translations iE, iM, iA2 sees much greater machine inertia. Furthermore, the inertia and games of the adjustment gear 6, 7, 7 'themselves also have a negative effect on the vibration behavior.
    Fig. 3 shows an example of a possible Eintriebs torque curve during a load change (good excitation of the first natural frequency by the hard stops after driving through the game). The dashed line 8 represents the default value ME, SOu of the torque of the input machine 2 changing with the time t (which is plotted on the abscissa). The solid line 9 represents the time profile of the actual acting torque ME. The extreme weakly damped, low-frequency vibration of the torque ME is of course undesirable and should be reduced.
    According to the present method for controlling the torque ME of the input drive machine 2, a wheel with tires that runs on a virtual roadway is virtually mounted on the drive-in machine 2. The road speed is chosen so that, together with the tire slip stationary exactly the desired torque ME, i. according to a default value ME, should (hereinafter also briefly called Msoll), stops. The principle of the control is shown schematically in FIG. As shown in FIG. 4, the torque control command value Mx of the input driving machine 2 indicated on the right as the output value is obtained from the tire model represented by the block 10. Apart from constant model parameters, which are not shown here, the current speed nE, the drive-in machine 2 and the virtual roadway speed v enter the tire model 10. The current input speed nlst is measured on the input machine 2. The virtual road speed v is obtained from the inverse tire model represented by the second block 11. The inverse tire model 11 generally does not correspond to the inverse of the tire model 10, but is e.g. a simplified tire model that is easier to invert. In addition to its own constant model parameters, a setpoint speed nE, SOu (or nsou) of the input drive machine 2 and the default value Msetpoint for the torque ME of the input drive machine 2 enter the inverse tire model 11.
    Each of the tire models underlying tire model 10 and inverse tire model 11 may be any tire model, i. a model which is suitable for approximating the transmission behavior of a wheel with a tire and / or has comparable damping properties. The following is an example of a simple tire model based on the so-called "magic formula". by Pacejka (see Pacejka H., "Tire and Vehicle Dynamics", 2nd edition, Butterworth-Heinemann, Oxford, 2007). However, one could also assume another tire model with similar qualitative characteristics, such as the model "TMsimple". von Hirschberg (see Hirschberg W., "TMsimple Application Manual", 2007), the model "TMeasy". from Rill (s.
    Rill G., "Simulation of Motor Vehicles", Vieweg-Verlag, Regensburg, 2007), the "Hohenheim Tire Model". or the tire model "TameTire" developed by the Michelin company, advantageously using the simplest possible version for the present method.
    The starting point in the model demonstrated here is a wheel with the angular velocity ωκ and with the dynamic rolling radius rdyn, which is assumed to be constant here, as well as a roadway velocity v. Alternatively, the dynamic rolling radius rdyn may be changed slightly depending on the road speed v (or the corresponding vehicle speed), e.g. according to a polynomial approach, however, the possible influence of such a change on the damping effect has been found to be negligible. With the above sizes, a slip of the virtual tire as
    are specified, the limit speed vlow prevents a division by zero. There are other definitions of slippage in literature as well; The one shown here is the preferred variant of Pacejka. (An alternative will also be explained below.) With this slip s, the simple magic formula " for the longitudinal movement of the coefficient of friction μ
    calculated with the three tire parameters B, C and D.
    Fig. 5 shows a coefficient of friction characteristic 12 for a typical tire. With the contact force Fz between the virtual tire and the virtual roadway, the longitudinal force acting in the tire noise can now
    and finally the torque due to the slip s
    be calculated. Usually, in a tire model, the rolling friction is taken into account. In order to simplify the present regulation, the modeling of the rolling friction is dispensed with here. For the uprising Fz could use a body model; for the sake of simplicity, however, one chooses a constant contact force FZ; 0. With these simplifications and when equation (5) is used in equation (8), one obtains with
    the relationship between torque Mx, angular velocity ωκ and roadway velocity v at a certain constant contact force FZ; 0- For the angular velocity ωκ the converted actual rotational speed nE, is (or short nist) of the input driving machine 2 is used in the control according to FIG.
    If one wants to calculate the appropriate road speed v for a given default value Msou of the torque ME and a current rotational speed nist, this is possible, but very expensive. Due to the non-linear coefficient of friction characteristic (see Fig. 5), there may also be several or no solutions at all. A simplification that makes sense in this context therefore consists of a linear characteristic instead of the non-linear coefficient of friction slip characteristic
    with slope k, the slope being best chosen as well as at the origin of the non-linear coefficient of friction characteristic; it then applies:
    Thus, equation (9) simplifies too
    and you can solve for the road speed to get to a simplified inverse tire model:
    In order to determine the road speed v from the inverse tire model according to equation (14), the setpoint value Msetpoint for the torque Mx and the converted target rotational speed nE; SOu of the input driving machine 2 for the angular velocity ωκ on nrfDoo ··· 7 + · ·
    The target rotational speed nE, SOu of the input driving machine 2 is calculated from the target rotational speeds nA1, SOll and nA2, soii of the output machines 3, 3 'taking into account the applicable gear ratios iD, iE. In the case of the powertrain test bench 1 with matching gears 6, 7, 7 'according to FIG. 2, the input engine 2 is obtained approximately for the desired rotational speed nE;
    FIG. 6 shows the dramatic improvement, compared to FIG. 3, of the input torque curve 9 'during a load change, i. a variation 8 of the default value Msou.
    In this case, the reference variable Mx for the torque control of the input drive machine 2 was dynamically determined on the basis of the tire model according to equation (9) with the roadway speed according to the inverse tire model according to equation (14).
    As an alternative to the static relationship of the slip s to the angular velocity oR and the road speed v given in equation (5), specifically, the damping effect of the virtual tire can be improved by a more detailed description taking into account the dynamics of the force buildup. For example, in the literature, the so-called "contact point model" is found. (single contact point transient tire model, see Pacejka, 2007), in which a point of contact between the tire and the road surface experiences a certain deflection u from the rest position due to the differences in speed, which corresponds to the differential equation
    with the slip speed
    and the maximum deflection σκ (relaxation length) is described. The slip s could now in principle with
    be calculated. In order to achieve a better damped behavior even at low speeds, Pacejka calculates
    With
    and the slip stiffness cFk suggested.
    The present method or the present control device is in no way limited to a specific tire model. As an alternative to the model of Pacejka (the "magic formula") given in equation (6), for example, the tire model "TMsimple" can be used. von Hirschberg with the basic formula
    where K, B and A are model parameters, or the tire model "TMeasy". by Rill with the basic formula
    where sM, sG, fM, μ0 ', μΜ and μ0 are (partially dependent) model parameters.
    The coefficients of friction-slip curves resulting from the tire models cited are very similar, i. they have a very similar course, at least in the area of interest. The courses of the three tire models given here are shown in the diagram shown in FIG. 7, wherein for reasons of space only the first quadrant of the coordinate system is shown. The course in the third quadrant is analogous in each case (see Fig. 5). The solid line in Fig. 7 represents the " magic formula " from Pacejka (see equation (6)), the dashed line represents the "TMsimple" model (see equation (22)) and the dotted line represents the "TMeasy" model (see equation (23)).
    Of course, depending on the choice of the model and the model parameters, even more divergent gradients can be provided. FIG. 7 demonstrates the gualitative behavior of the reference variable generally achieved by tire models, in particular with a continuous course of the friction-value slip characteristic, which initially rises approximately linearly starting from the zero point, flattens out and finally monotonically drops with still increasing slip.
    权利要求:
    Claims (12)
    [1]
    1. A method for controlling a powertrain test stand (1) with a Eintriebsmaschine (2) and an output machine (3, 3 '), wherein one of the Eintriebsmaschine (2) applied torque is controlled, wherein a reference variable (Mx) for the Regulation of the torque of the input drive machine (2) for damping vibrations between the input drive machine (2) and the output machine (3, 3 ') in dependence on a current speed (nist) of the input drive machine (2) compared to a default value (Msou) is modified characterized in that the reference variable (Mx) is determined from a tire model (10) of a virtual tire.
    [2]
    2. The method according to claim 1, characterized in that the tire model (10) establishes a, preferably static, relationship between the reference variable (Mx) and a slip of the virtual tire.
    [3]
    3. The method according to claim 1 or 2, characterized in that the reference variable (Mx) is calculated substantially according to the formula Mx = Fz-rdYTi-D-sin (c-arctan (Bs)), wherein Fz a Aufstandskraft, rdyn a Roll radius, B, C and D are constant tire parameters and s is a slip of the virtual tire.
    [4]
    4. The method according to any one of claims 1 to 3, characterized in that a virtual road speed (v), which enters the tire model (10), in particular in the slip of the virtual tire, is determined from an inverse tire model (11).
    [5]
    5. The method according to claim 4, characterized in that the virtual road speed (v) from the default value (Msoll) for the reference variable (Mx) and a target rotational speed (nson) of the input machine (2) is determined, which target rotational speed ( nsou) is preferably proportional to a desired rotational speed of the output machine (3, 3 ').
    [6]
    6. The method according to claim 4 or 5, characterized in that the inverse tire model (11) establishes a linear relationship between the virtual roadway speed (v) and the target speed (nsoli) of the input drive machine (2).
    [7]
    7. Control device of a powertrain test stand (1) with a drive-in machine (2) and an output machine (3, 3 '), for carrying out the method according to one of claims 1 to 6, with an inner control loop for controlling one of the input drive machine (2 ) applied torque and with an outer control loop for controlling the reference variable (Mx) of the inner loop in dependence on a current speed (nist) of the input drive machine (2), characterized in that the outer control loop has a tire model (10) of a virtual tire.
    [8]
    8. Control device according to claim 7, characterized in that the tire model (10) has a, preferably static, relationship between the reference variable (Mx) and a slip of the virtual tire.
    [9]
    9. Control device according to claim 7 or 8, characterized in that the outer control loop for determining the reference variable (Mx) is essentially set up according to the formula Mx = Fz-rdYTi-D-sin (c-arctane (Bs)), wherein Fz a rearing force, rdyn a rolling radius, B, C and D constant tire parameters and s is a slip of the virtual tire.
    [10]
    10. Control device according to one of claims 7 to 9, characterized in that the outer control loop for determining a virtual road speed (v), which enters the tire model (10), in particular in the slip of the virtual tire, from an inverse tire model ( 11) is set up.
    [11]
    11. Control device according to claim 10, characterized in that the outer control loop for determining the virtual roadway speed (v) is set up from a default value (Msetpoint) for the reference variable (Mx) and a setpoint rotational speed (nson) of the input drive machine (2), wherein the setpoint speed (nsetpoint) is preferably proportional to a setpoint speed of the output machine (3, 3 ').
    [12]
    12. Control device according to claim 10 or 11, characterized in that the inverse tire model (11) has a linear relationship between the virtual roadway speed (v) and the target rotational speed (nsetpoint) of the input drive machine (2).
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    引用文献:
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    法律状态:
    优先权:
    申请号 | 申请日 | 专利标题
    ATA50011/2014A|AT515110B1|2014-01-09|2014-01-09|Method and device for controlling a powertrain test stand|ATA50011/2014A| AT515110B1|2014-01-09|2014-01-09|Method and device for controlling a powertrain test stand|
    US15/110,382| US9651452B2|2014-01-09|2015-01-08|Method and device for controlling a powertrain test stand|
    EP15703724.3A| EP3092471B1|2014-01-09|2015-01-08|Method and device for controlling a powertrain test stand|
    PCT/AT2015/050007| WO2015103658A1|2014-01-09|2015-01-08|Method and device for controlling a powertrain test stand|
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