JP2008070119A - Method and device for controlling engine bench system - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To solve a problem wherein axial torque control in an engine bench system requires a lot of time to determine a weighting factor. <P>SOLUTION: A control unit to compute a dynamometer torque control signal T2 for an inverter is set based on the following equation. T2=(Ki/s)*(T12r-T12)-(Kp+s*Kd)/(a2*s*s+a1*s+1)*T12. If a connecting shaft has a nonlinear spring characteristic, a dynamometer torque control signal T2 is computed after determining the connecting shaft rigidity to a resonance frequency of the nonlinear spring characteristic based on an axial torque value. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to an engine bench system control method and apparatus, and more particularly to a control method and apparatus for directly obtaining a control gain.

  In an engine bench system configured to connect a test engine and a dynamometer as a power absorber by a coupling shaft and control the dynamometer by an inverter, the calculation unit calculates a deviation between the torque setting value and the detected torque value. Calculations are executed, and speed and torque are controlled based on the calculation results. In an engine bench system that is PID controlled, there is a risk that resonance of the coupling shaft may occur due to pulsating torque generated by the engine. Therefore, Patent Document 1 discloses a shaft torque control method using the μ design method, which is one of the robust control design theories, as a method for controlling the shaft torque of an engine bench system that requires resonance suppression.

In Patent Document 1, the controller uses a μ design method for the transfer function from the torque command value of the dynamometer to the dynamometer torque and the transfer function from the engine torque or power torque to the engine speed, shaft torque, or dynamometer speed. And having a transfer function that uses a structured singular value μ as a condition of robust stability and robust performance for each perturbation. The disturbance input in the μ design method is a shaft torque or dynamometer speed detection signal, the controlled variable is a torque deviation and an engine load torque command value, and the perturbation term is an engine moment of inertia and a shaft spring constant.
JP 2003-149085 A

  The thing of patent document 1 has the advantage that measurement of various characteristics of an engine can be obtained stably and at high speed because the shaft torque or engine load torque control is performed by a controller configured based on the μ design method. ing.

  However, this document requires roughly two operations: (1) modeling of the controlled object and (2) determination of the weighting function. Of these, the determination of the weighting function requires trial and error and takes time. It has such a problem. Depending on the control specifications of the system, there is a specification that does not require the μ design method, and there is a demand for a shaft torque control method having a resonance suppression effect in a PID control system for such specifications.

  An object of the present invention is to provide a control method and apparatus capable of directly calculating a control gain without adjusting a weight function.

Claim 1 of the present invention connects an engine and a dynamometer with a coupling shaft, detects the axial torque of the coupling shaft, the angular velocity of the engine and the angular velocity of the dynamometer, and detects each detected value and shaft torque command value. In an engine bench system configured to control a dynamometer via an inverter based on the calculation of
The controller for calculating the dynamometer torque control signal T2 for the inverter is based on the following equation.

T2 = (Ki / s) * (T12r-T12)-(Kp + s * Kd) / (a2 * s * s + a1 * s + 1) * T12
However, Ki: integral coefficient, T12r: shaft torque command value, T12: shaft torque detected value, Kp: proportional coefficient, Kd: differential coefficient, a1, a2: filter coefficient Claim 2 of the present invention is nonlinear to the coupling shaft. The engine bench system having a spring characteristic is characterized in that the dynamometer torque control signal T2 is calculated after obtaining the coupled shaft rigidity with respect to the resonance frequency of the nonlinear spring characteristic based on the shaft torque value.

According to a third aspect of the present invention, the engine and the dynamometer are connected by a coupling shaft, the shaft torque of the coupling shaft, the angular velocity of the engine, and the dynamometer angular velocity are detected. In an engine bench system configured to control a dynamometer via an inverter based on the calculation of
The controller for controlling the inverter is calculated based on the detected shaft torque value and a table creating means for representing the torsion torque-resonance frequency, and a coupled shaft spring stiffness calculating section for the resonance frequency obtained by the table creating means. A weight coefficient calculating unit for calculating each of the integral coefficient, proportional coefficient, differential coefficient, a2 and a1 coefficients using the spring stiffness value of the coupled shaft, and the difference signal between the shaft torque command value and the shaft torque value. An integration element unit that calculates using an integration coefficient calculated by a coefficient calculation unit; a proportional / differentiation element unit that calculates a proportional / differential element using the coefficient calculated by the weight coefficient calculation unit for the shaft torque value; A means for calculating a difference signal between the output of the integral element part and the proportional / differential element part is provided.

As described above, according to the present invention, it is possible to directly determine the control gain in the control of the engine bench system. Therefore, the resonance suppression effect can be easily obtained without adjusting the weighting coefficient as in the prior art. It is possible to obtain the control gain.
In addition, it has the effect of suppressing each resonance point even for a nonlinear engine bench system in which the resonance frequency changes.

Prior to a specific description of the present invention, the underlying technology will be described.
In general, the mechanical system configuration of an engine bench system is expressed as a multi-inertia system model having two or more inertia systems, but the present invention is directed to an engine bench system that can be approximated to a two-inertia system. FIG. 1 shows the two-inertia mechanical system model.

  The equation of motion of the 2-inertia machine model is as follows: J1: Engine moment of inertia, J2: Dynamometer moment of inertia, K12: Coupling shaft spring stiffness, T12: Coupling shaft torsion torque (shaft torque), w1: Engine angular velocity, w2: Dynamometer When the angular velocity, T2: dynamometer torque, and the Laplace operator as s, equations (1) to (3) are obtained.

Here, by solving equations (1) to (3) for w1, w2 and T12, (4)
It becomes an expression.

  In the present invention, the command value of the shaft torque T12 is expressed as T12r, and the dynamometer torque T2 is controlled based on the calculation of equation (5).

  Here, when Equations (4) and (5) are combined and solved for T12 and T2, Equation (6) is obtained.

Here, Ki is an integral coefficient, Kp is a proportional coefficient, Kd differential coefficient, and a1 and a2 are filter coefficients. The denominator polynomial in equation (6) is a quintic equation, and the denominator polynomial is divided by the constant term J1 * J2 * Ki, and the first to fifth coefficients of the equation with the constant term set to 1 are Ki, Kp, Kd, a2 and a1 are independent. Therefore, if the coefficients are compared so that the denominator polynomial of equation (6) becomes 1 + c1 * s + c2 * s ^ 2 + c3 * s ^ 3 + c4 * s ^ 4 + c5 * ^ 5, equations (7) to (11) are obtained. The coefficients Ki, Kp, Kd, a2, and a1 may be determined.

The first embodiment of the present invention is a two-inertia approximation engine bench system that is modeled as shown in FIG. 1 and expressed as equations (1) to (3). First, the resonance frequency wc (rad / s) is obtained from Equations (1) to (3) as shown in Equation 12.
K12 = J1 * J2 * wc ^ 2 / (J1 + J2) (12)
Next, a fifth order polynomial having an appropriate damping characteristic is selected. For example, as a binomial coefficient type in which all poles have a damping coefficient of 1,
1 / (s + 1) 5 = 1 / ((1+ (5 * s) + (10 * s ^ 2) + (10 * s ^ 3) + (5 * s ^ 4) + s ^ 5))
And the characteristic without the vibration element, s is replaced with s / wc, and the coefficient is c1
c5.

  In the binomial counting type, c1 = 5 / wc, c2 = 10 / wc, c3 = 10 / wc ^ 3, c4 = 5 / wc ^ 4, and c5 = 1 / wc ^ 5. For the obtained coefficients c1 to c5, parameters Ki, Kp, Kd, a2, and a1 of the PID type controller expressed by equation (5) are determined according to equations (7) to (11).

  FIG. 4 shows a characteristic diagram of a conventional engine bench system. This figure is a Bode diagram from T2 to T12 when J1 = 0.2, K12 = 1500, and J2 = 0.7 based on design values based on past actual machines. In this case, the engine bench system has a resonance frequency of about 15 Hz. The steady-state gain (gain in the low range) from the dynamometer torque command T2 to the shaft torque detection T12 is also about -13 dB.

  FIG. 3 shows a characteristic diagram of the engine bench system measured based on the first embodiment of the present invention. That is, the shaft torque is detected from the shaft torque command value T12r when the gains of Ki, Kp, Kd, a2, and a1 calculated based on the equations (7) to (11) are directly calculated according to the equation (5). It is a Bode diagram of T12.

  As apparent from the comparison between FIG. 3 and FIG. 4, in FIG. 4, the resonance frequency is near 15 Hz, and the steady (low frequency) gain near 0 Hz is about −13 dB. On the other hand, in the case of the present invention shown in FIG. 3, there is no resonance frequency as shown in FIG. 4, and the gain in the low band is 0 dB.

  Therefore, according to the first embodiment, the control gain can be directly determined by the PID type controller of the engine bench system. Therefore, the resonance suppression effect can be easily achieved without adjusting the weighting coefficient as in the prior art. It is possible to obtain the control gain of the control gain.

  FIG. 2 shows a configuration diagram of an axial torque control system controller in the case where the coupling shaft portion has a nonlinear spring such as a clutch. Reference numeral 1 denotes an engine as a test machine, 2 denotes a dynamometer, and the engine 1 and the dynamometer 2 are directly connected by a coupling shaft 3. Reference numeral 4 denotes a torsion torque detector, and reference numeral 5 denotes a table creation means for creating a table T12f representing the relationship between torsion torque and resonance frequency. Reference numeral 6 denotes a combined shaft spring stiffness calculating section, and 7 is a weighting coefficient calculating section that calculates the coefficients Ki, Kp, Kd, a2, and a1 based on the equations (7) to (11) described above. 8 and 11 are adders, 9 is an integral element, and 10 is a proportional / differential element.

In the case of a non-linear spring in which the spring stiffness value changes depending on the torsion angle, the table T12f is created in advance by some means. This table is created by the table creation means 5, for example, when the characteristics of the nonlinear spring are known, it is obtained by calculation, and when the characteristics are unknown, the torsion torque has a certain value. The table T12f is created by means such as actually measuring the resonance frequency during
FIG. 5 is a characteristic diagram created by such means. (A) is a torsion angle-torsion torque characteristic diagram, (b) is a torsion torque-clutch rigidity (spring constant) characteristic diagram, and (c) torsion torque-resonance. Examples of frequency characteristic diagrams are respectively shown.

  The coupled shaft spring stiffness calculation unit 6 obtains the coupled shaft spring stiffness K12 from the equation (12) with respect to the resonance frequency wc (rad / s) obtained by the table creation means 5.

  The weight coefficient calculation unit 7 calculates the coefficients Ki, Kp, Kd, a2, and a1 based on the equations (7) to (11), and the calculated values are calculated when the integral element unit 9 and the proportional / differential element unit 10 are calculated. Are output as weighting factors. That is, the deviation value between the shaft torque command value T12r obtained by the adder 8 and the torsion torque detection value T12 is input to the integral element unit 9 and is integrated using the coefficient Ki calculated by the weight coefficient calculator 7. The operation is executed.

  Further, the torsional torque detection value T12 is input to the proportional / differential element unit 10, and the proportional / differential calculation is executed using each coefficient calculated by the weighting coefficient calculating unit 7. The output of the integral element unit 9 and the output of the proportional / differential element unit 10 are respectively output to the adder unit 11, and the adder unit 11 obtains the deviation between them. This deviation value becomes the torque command T2 of the dynamometer 2 according to the equation (5).

  According to the second embodiment, even in a system in which the resonance frequency of the Bode diagram shown in FIG. 4 varies depending on the magnitude of the axial torque, the gain of the equation (5) constantly adapted to the resonance frequency as shown in FIG. Since the parameters Ki, Kp, Kd, a2, and a1 are selected, each resonance point is effectively suppressed as shown in FIG. 3 even for a nonlinear engine bench system in which the resonance frequency changes.

2 is a two-inertia model diagram of an engine bench system to which the present invention is applied. FIG. The block diagram of the shaft torque control system controller of this invention. The Bode diagram of axis torque detection from the axis torque command by the present invention. Board diagram of shaft torque detection from conventional shaft torque command. FIG. 6 is a characteristic diagram of the table creation means, where (a) is a torsion angle-torsion torque characteristic, (b) is a torsion torque-clutch rigidity characteristic, and (c) is a torsion torque-resonance frequency characteristic.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 ... Engine 2 ... Dynamometer 3 ... Connection shaft 4 ... Torsion torque detection part 5 ... Table preparation means 6 ... Connection shaft spring rigidity calculation part 7 ... Weight coefficient calculation part 8, 11 ... Addition part 9 ... Integration element part 10 ... Proportional・ Derivative element

Claims (3)

The engine and dynamometer are connected by a coupling shaft, and the shaft torque of the coupling shaft, the angular velocity of the engine and the dynamometer angular velocity are detected, and an inverter is connected based on the calculation of each detected measurement value and shaft torque command value. In an engine bench system configured to control a dynamometer,
The controller for calculating the dynamometer torque control signal T2 for the inverter is based on the following equation, and is a control method for an engine bench system.
T2 = (Ki / s) * (T12r-T12)-(Kp + s * Kd) / (a2 * s * s + a1 * s + 1) * T12
However, Ki: integral coefficient, T12r: shaft torque command value, T12: shaft torque detection value, Kp: proportional coefficient, Kd: differential coefficient, a1, a2: filter coefficient In the engine bench system having a non-linear spring characteristic on the coupling shaft, the dynamometer torque control signal T2 is calculated after obtaining a coupling shaft stiffness with respect to a resonance frequency of the non-linear spring characteristic based on the shaft torque value. Item 4. A method for controlling an engine bench system according to Item 1. The engine and dynamometer are connected by a coupling shaft, and the shaft torque of the coupling shaft, the angular velocity of the engine and the dynamometer angular velocity are detected, and an inverter is connected based on the calculation of each detected measurement value and shaft torque command value. In an engine bench system configured to control a dynamometer,
The controller for controlling the inverter is calculated based on the detected shaft torque value and a table creating means for representing the torsion torque-resonance frequency, and a coupled shaft spring stiffness calculating section for the resonance frequency obtained by the table creating means. A weight coefficient calculating unit for calculating each of the integral coefficient, proportional coefficient, differential coefficient, a2 and a1 coefficients using the spring stiffness value of the coupled shaft, and the difference signal between the shaft torque command value and the shaft torque value. An integration element unit that calculates using an integration coefficient calculated by a coefficient calculation unit; a proportional / differentiation element unit that calculates a proportional / differential element using the coefficient calculated by the weight coefficient calculation unit for the shaft torque value; An engine bench system control device comprising means for calculating a difference signal between an output of an integral element and a proportional / differential element.

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