JP4858376B2 - Vibration control device for hybrid vehicle - Google Patents

Description

The present invention relates to a control apparatus of a hybrid vehicle equipped with engine and motor / generator.

In the hybrid vehicle equipped with a motor / generator is an engine and the electric motor is an internal combustion engine involves periodic torque variation due to cycles associated with explosive combustion in the engine. When this torque fluctuation excites the resonance of the drive system including the power source, there is a problem that a large vibration is generated and the ride comfort performance is impaired.

For this reason, a device as described in Patent Document 1 is conventionally known as a vibration damping device that suppresses or prevents periodic torque fluctuations caused by the explosion combustion cycle of the engine by the output torque of the motor / generator.
The vibration damping device disclosed in Patent Document 1 controls the output torque of a motor / generator and controls the engine explosion in a hybrid vehicle in which the rotation shaft of the motor / generator is connected to the engine output shaft via an elastic buffer mechanism. It is intended to cancel the periodic torque fluctuation caused by the combustion cycle.
JP 2001-136605 A

However, the vibration damping device described in Patent Document 1 still has the problems described below. Conventionally, in order to reduce or avoid vibration input from the engine to the vehicle body, a power source is attached to the vehicle body frame with a mount made of an elastic material, and the resonance of the power system, the mount, and the mount system that the vehicle body constitutes However, there is still a problem that a large vibration is generated by exciting the resonance of the mount system. Further, in the vibration damping device described in Patent Document 1, the amplitude and phase of the control torque must be stored in advance in the controller as a multidimensional map, and there is a problem that a large amount of memory is required. .
It is an object of the present invention to effectively prevent drive system vibration and mount system vibration.

The vibration control apparatus for a hybrid vehicle according to the present invention for the purpose, as described in claim 1, comprising the engine and the motor / generator, an output shaft of the engine via a torsional damper to dampen the rotational direction force In a hybrid vehicle that is coupled to the rotation shaft of the motor / generator and supports the power source on the vehicle body side via an elastically deformable mount member,
Motor / generator torque command means for generating torque obtained by multiplying the engine torque variation of the engine by a predetermined coefficient from the motor / generator;
The explosion frequency corresponding to the engine torque fluctuation determined from the number of engine explosions per one rotation of the engine output shaft and the engine rotation speed is the relative rotation between the engine output shaft and the motor / generator rotation shaft with respect to the torsion damper means. When the torsional vibration system of the power source and the displacement vibration system that is the relative displacement of the power source with respect to the vehicle body with respect to the mount member substantially match the resonance frequency, the excitation force that excites the resonance of these vibration systems is Predetermined coefficient calculation means for obtaining the predetermined coefficient in accordance with the engine speed so as to be minimized;
It is characterized by comprising.

According to the vibration control device of the present invention, the torsional vibration system of the power source in which the explosion frequency corresponding to the engine torque fluctuation is a relative rotation between the engine output shaft and the motor / generator rotation shaft with respect to the torsional damper means, When the resonance frequency of a displacement vibration system that is a relative displacement of the power source with respect to the vehicle body with respect to the mount member substantially matches, the excitation force that excites the resonance of the vibration system is minimized according to the engine speed. In order to obtain the predetermined coefficient,
The vibration of the drive system and the vibration of the mount system can be suitably reduced. Further, when the predetermined coefficient with respect to the engine rotation speed is stored as a table, the drive control memory can be saved by making it simpler than the conventional multidimensional map.

Hereinafter, embodiments of the present invention will be described in detail based on examples shown in the drawings.
FIG. 1 shows a power train of a front engine / rear wheel drive hybrid vehicle equipped with a vibration control apparatus according to an embodiment of the present invention, wherein 1 is an engine and 2 is a drive wheel (rear wheel).
In the power train of the hybrid vehicle shown in FIG. 1, the automatic transmission 3 is arranged in tandem at the rear of the engine 1 in the longitudinal direction of the vehicle in the same manner as a normal rear wheel drive vehicle, and the engine 1 (crankshaft 1a) is rotated. A motor / generator 5 is provided in combination with the shaft 4 that transmits to the input shaft 3a of the automatic transmission 3. The engine 1 is a reciprocating type internal combustion engine such as a gasoline engine or a diesel engine, and has four cycles and six cylinders.

The motor / generator 5 functions as a motor or functions as a generator (generator), and is disposed between the engine 1 and the automatic transmission 3.
More specifically, a torsional damper 6 having a damper function for buffering torque fluctuations is interposed between the motor / generator 5 and the engine 1, more specifically, between the shaft 4 and the engine crankshaft 1a. The torque fluctuation between 1 and the motor / generator 5 is reduced.
The torsional damper 6 is formed by inserting a torsion spring that is axially twisted and absorbs torque fluctuation between two disks that are coaxial and face each other, and the disks are connected by this torsion spring.
Alternatively, the torsional damper 6 may be, for example, a well-known torsional damper or the like, in which a plurality of springs are arranged in the circumferential direction between two disks that are coaxial and face each other, and the disks are connected by these springs. These springs expand and contract in the circumferential direction of the rotary shafts 1a and 4 to alleviate fluctuations in the transmission torque.
The torsion damper 6 may have a clutch function for connecting and disconnecting between the shaft 4 and the engine crankshaft 1a.

More specifically, a clutch 7 is inserted between the motor / generator 5 and the automatic transmission 3 and, more specifically, between the shaft 4 and the transmission input shaft 3a, and the motor / generator 5 and the automatic transmission 3 are separated by this clutch 7. Join as possible.
The clutch 7 has a transmission torque capacity that can be changed continuously or stepwise. For example, a wet multi-plate that can change the transmission torque capacity by continuously controlling the clutch hydraulic oil flow rate and the clutch hydraulic pressure with a proportional solenoid. Although it is constituted by a clutch, it is preferable that a torsional damper is incorporated so as to alleviate torque fluctuation to the automatic transmission 3.

The automatic transmission 3 is the same as that described in pages C-9 to C-22 on the "Skyline New Car (CV35) Manual" issued by Nissan Motor Co., Ltd. in January 2003. By selectively engaging and releasing the shift friction elements (clutch, brake, etc.), the transmission path (shift stage) is determined by the combination of engagement and release of these shift friction elements.
Therefore, the automatic transmission 3 shifts the rotation from the input shaft 3a at a gear ratio corresponding to the selected shift speed and outputs it to the output shaft 3b.
This output rotation is distributed and transmitted to the left and right rear wheels 2 via the left and right drive shafts 20 by the differential gear device 8, and used for traveling of the vehicle.

  The power train support structure of the hybrid vehicle shown in FIG. 1 will be described. The drive transmission path from the torsion damper 6 to the automatic transmission 3 of the hybrid vehicle of the present embodiment is extended from the engine 1 as shown in FIG. Housed in case 11. The integrated structure of the engine 1 and the case 11 is attached to the vehicle body 10 via an engine mount 9 (also referred to as an insulator). The engine mount 9 is made of an elastic material such as anti-vibration rubber, and prevents engine vibration from being transmitted to the vehicle body 10 side.

The control of the engine 1, the motor / generator 5, and the clutch 7 constituting the power train of the hybrid vehicle shown in FIG. 1 will be described.
When the fuel injection device 1n attached to the engine 1 receives a fuel injection command from the driving force controller 16, the fuel injection device 1n injects fuel at a timing and an amount corresponding to the command toward each cylinder of the engine 1.

  The motor / generator 5 is connected to the inverter 17 via a power cable, and the inverter 17 is connected to the battery 19 via a power cable. The inverter 17 drives the shaft 4 with the target motor / generator torque by controlling the current supplied to the motor / generator 5 in accordance with the torque command from the driving force controller 16.

  A clutch actuator 18 that engages and disengages the clutch 7 sets the transmission torque capacity of the clutch 7 being engaged to a target value in accordance with an output command from the driving force controller 16.

In order to determine the above commands to be input to the fuel injection device 1n, the inverter 17, and the clutch actuator 18, the driving force controller 16 includes a signal from the crank angle sensor 12 that detects the crank angle of the engine 1,
A signal from the rotational speed sensor 13 for detecting the engine rotational speed (rotation angle);
A signal from the accelerator sensor 14 that detects the accelerator opening of the accelerator operator operated by the driver;
A signal from the brake sensor 15 for detecting the brake operation amount of the brake operator operated by the driver,
Other detection signals (such as the state of the battery 19) necessary for controlling the power train are input.

  The driving force controller 16 calculates the target engine torque, the target motor / generator torque, and the target clutch transmission torque capacity that can realize the driving force of the vehicle desired by the driver from the above input information and other input information, respectively. To do.

The driving force controller 16 calculates the target engine torque, the target motor / generator torque, and the target clutch transmission torque capacity as described above.
The calculation described below is executed to perform engine torque fluctuation correction control for correcting the torque fluctuation of the engine 1 targeted by the present invention, and the torque fluctuation generated in the crankshaft 1a when the engine 1 is started and stopped. The correction value corresponding to the minute is corrected to be added to the target motor / generator torque of the motor / generator 5.

  FIG. 2 is an explanatory diagram showing a vibration model formed by the power train of the hybrid vehicle shown in FIG.

  In this embodiment, the rotating part 1r in the engine 1, the rotor part 5r of the motor / generator 5 and the engine case part 1s are mass bodies, and the torsion damper 6 and the engine mount 9 are spring elements. The mass body of the engine rotating portion 1r includes the engine crankshaft 1a, and the mass body of the rotor 5r of the motor / generator 5 includes the shaft 4. The mass body of the engine case portion 1s is a non-rotating portion obtained by removing a rotating shaft operating portion such as a crankshaft from the engine 1 itself, and includes a non-rotating portion outside the engine 1, a stator portion of the motor / generator 5, and an extension. An installation case 11.

  In the vibration model shown in FIG. 2, since the vehicle body 10 supports the engine case portion 1 s through an elastically deformable engine mount 9, the engine case portion 1 s is a displacement vibration system in which the engine case portion 1 s is rotationally displaced around the rotation axis of the engine 1. There exists a roll vibration system. When the resonance frequency of the roll vibration system substantially coincides with the explosion frequency corresponding to the engine torque fluctuation, the roll resonance of the engine case portion 1s is generated and a large vehicle body vibration is generated, so that riding comfort performance is impaired. The explosion frequency corresponding to the engine torque fluctuation is a periodic torque fluctuation caused by the explosion combustion cycle of the engine 1. Specifically, it is determined from the number of engine explosions per one rotation related to the engine crankshaft 1a and the engine rotation speed. In the 4-cycle 6-cylinder engine 1, explosion occurs 6 times per 2 revolutions, so the number of engine explosions per revolution is 3 times. Therefore, the explosion frequency becomes a rotation third-order frequency.

  In the vibration model shown in FIG. 2, the engine crankshaft 1 a is coupled to the rotor shaft 4 of the motor / generator 5 via the torsional damper 6 that buffers the rotational direction force. There is a torsional vibration system that is relatively rotated. When the resonance frequency of the torsional vibration system substantially coincides with the above-described explosion frequency, a large rotational vibration is generated, so that the riding comfort performance is also impaired.

  Normally, the above-described resonance frequency of the roll vibration system and the resonance frequency of the torsional vibration system are designed to be removed from the normal rotational speed region of the engine 1. Specifically, the spring constants of the engine mount 9 and the torsion damper 6 are designed so that the frequency is equal to or lower than the rotation basic order related to the idling rotation speed of the engine 1. For this reason, the above-mentioned problem of ride comfort performance does not become apparent.

  However, when the engine is started to start the operation of the engine 1 and when the engine is stopped to stop the operation of the engine 1, the rotation basic order component related to the engine rotation speed passes through the resonance frequencies of the roll vibration system and the torsional vibration system. The ride performance will be impaired. For example, when the engine is started, when torque is input to the engine 1 for cranking, air is compressed and expanded by the rotation of the crankshaft 1a, so that the pressure in the cylinder fluctuates. Then, the cylinder internal pressure is converted into torque according to the change of the crank angle due to the geometric shape of the crank and connecting rod, and torque fluctuation is generated, which becomes an excitation force for the roll vibration system and the torsional vibration system.

  In this embodiment, for the purpose of eliminating such roll vibration and rotational fluctuation of the engine 1, correction control for engine torque fluctuation by the motor / generator 5 described below is performed.

ここでθ_cはエンジン回転部1rの回転角、θ_mはロータ部5rの回転角、θ_eはエンジンケース部１ｓの回転角、Icはエンジン回転部１rの慣性モーメント、Imはロータ部5rの慣性モーメント、Ieはエンジンケース部１ｓの慣性モーメント、Kcはねじりダンパ６のばね定数、Ccはねじりダンパ６の減衰定数、Keはエンジンマウント９のばね定数、Ceはエンジンマウント９の減衰定数、Teはエンジントルク、Tmはモータ／ジェネレータトルクである。 First, the equation of motion of the vibration system is established from the mass body and the spring elements that constitute the vibration model shown in FIG.

Here, θ _c is the rotation angle of the engine rotation portion 1r, θ _m is the rotation angle of the rotor portion 5r, θ _e is the rotation angle of the engine case portion 1s, Ic is the moment of inertia of the engine rotation portion 1r, and Im is the rotation angle of the rotor portion 5r. Moment of inertia, Ie is the moment of inertia of the engine case 1s, Kc is the spring constant of the torsional damper 6, Cc is the damping constant of the torsional damper 6, Ke is the spring constant of the engine mount 9, Ce is the damping constant of the engine mount 9, Te Is the engine torque and Tm is the motor / generator torque.

The engine case portion inertia moment Ie also includes the stator portion of the motor / generator 5 and the engine extension case 11. The displacement volume of the engine case section 1s from the rotation radius of the rotation angle theta _e and the engine case section 1s of the engine case portion 1s is obtained.

The formula (1) is summarized as follows.

ここでFnの中央はロール振動系が共振するロール振動モードの共振周波数、右側はねじり振動系が共振するねじり振動モードの共振周波数を表す。また、φの左列は剛体系、中列はロール振動モードの共振、右列はねじり振動モードの共振を表す。これら共振周波数[Hz]を回転３次の回転速度[rpm]に単位変換するには２０倍すればよいことから、ロール共振周波数12.79[Hz] は256[rpm]になり、ねじり共振周波数28.09[Hz]は562 [rpm]となる。 From the equation (2), the resonance frequency Fn and mode vector φ of this vibration system are as follows by performing eigenvalue analysis of −ω ² M + K = 0.

Here, the center of Fn represents the resonance frequency of the roll vibration mode in which the roll vibration system resonates, and the right side represents the resonance frequency of the torsional vibration mode in which the torsion vibration system resonates. The left column of φ represents the rigid system, the middle column represents the resonance in the roll vibration mode, and the right column represents the resonance in the torsional vibration mode. Since these resonance frequencies [Hz] need only be multiplied by 20 to convert the unit to the rotation tertiary rotation speed [rpm], the roll resonance frequency 12.79 [Hz] becomes 256 [rpm], and the torsional resonance frequency 28.09 [ Hz] is 562 [rpm].

(2) multiplying the inner product a transposed matrix ^t phi mode vector phi both sides of the equation, rigid body system, rewritten as follows using the mode displacement ξ of rolling vibration mode and torsional vibration mode.

Here, since M _r and K _r are diagonal matrices, when the damping is lower than the left side of the equation (5), it can be considered as a superposition of the one-degree-of-freedom vibration system of each vibration mode. ^T φT on the right side is the excitation force for these vibration modes. The components of this ^t .phi.T is the sum of the excitation force by excitation force and the motor / generator torque by the engine torque for one vibration mode, hereinafter referred to as modal forces. When the engine torque T _e is applied, if generating a motor torque T _m as the modal force becomes 0, the resonance of the vibration system does not occur worsening vibration not excited.

（６）式を解き、このようなトルクをエンジンケース部１ｓのロール振動モード（T_m1）、回転軸のねじり振動モード（T_m2）それぞれについて求めると、

となる。

By solving the equation (6) and obtaining such torque for each of the roll vibration mode (T _m1 ) of the engine case 1 s and the torsional vibration mode (T _m2 ) of the rotating shaft,

It becomes.

エンジン１の運転中は、（８）式の係数Ｋを、エンジン回転速度に応じて適宜定めるとよい。つまり、ロール振動系の加振力を０にするときは、Ｋ＝−１と定める。またねじり振動系の加振力を０にするときは、Ｋ＝０．７３と定める。 Therefore, the relationship between the motor / generator torque _Tm and the engine torque _Te is generalized as

During operation of the engine 1, the coefficient K in the equation (8) may be appropriately determined according to the engine speed. That is, when the excitation force of the roll vibration system is set to 0, K = −1. Further, when the excitation force of the torsional vibration system is set to 0, K = 0.73 is set.

However, according to the above equation (6), since T _m1 and T _m2 are opposite in sign, if the coefficient K is determined so as to make the excitation force zero for either vibration system resonance, Exciting force in the vibration system resonance of is worsened.
Therefore, in the present invention, the modal force is controlled so that the excitation force in the vibration mode becomes 0 at the engine rotation speed at which the third-order explosion frequency coincides with the resonance frequency at which the vibration greatly deteriorates, and the engine torque fluctuation component is controlled. Is multiplied by a coefficient K from the motor / generator 5.
Further, at an engine rotational speed at which the explosion frequency does not match the resonance frequency, a coefficient K is determined so that the modal force is minimized, and a torque obtained by multiplying the engine torque fluctuation by the coefficient K is input from the motor / generator 5.
Specifically, the engine torque Te is estimated from the detected engine speed and crank angle with reference to a map stored in the drive controller 16, and a coefficient K multiplied by the engine torque Te is driven as a coefficient table for the engine speed. The coefficient K is stored in advance in the controller 16 and the coefficient K is obtained by referring to the coefficient table according to the engine speed.

  The coefficient table described above is shown in FIG. According to FIG. 3A, the coefficient K is set to -1 at about 256 rpm at which the roll resonance of the engine case portion deteriorates, and the coefficient K is set to 0.73 at about 562 rpm at which the torsional resonance of the rotating shaft deteriorates. FIG. 3B shows a modal force, which is an excitation force for the vibration mode, based on the motor / generator torque output from the motor / generator based on the coefficient table of FIG. According to this embodiment, the coefficient table sets the modal force of each vibration mode to 0 in each vibration mode, thereby preventing excitation of resonance due to engine torque fluctuation.

  Regarding the rotational speed other than the resonance frequency, at a rotational speed lower than the engine case roll resonance (less than 256 [rpm]), the coefficient K is set to a constant value −1 so as to reduce the engine main body roll resonance. At a rotational speed higher than the rotational axis torsional resonance (exceeding 562 [rpm]), the coefficient K is set to a constant value 0.73 so as to reduce the rotational axis torsional resonance. During these periods (from 256 [rpm] to 562 [rpm]), the coefficient K is continuously determined according to the engine speed, and a constant value −1 and a constant value 0.73 are linearly displayed on the table. 3 is determined.

FIG. 4 shows the effect of vibration reduction by the coefficient table of FIG.
In FIG. 4, the solid line is an explanatory view showing the vibration reduction effect by performing the engine torque fluctuation correction control based on the coefficient table of FIG. Further, for comparison, a case where the engine torque fluctuation correction control of this embodiment is not performed is indicated by a broken line. FIG. 4A shows the engine speed when the engine 1 is started and then stopped. 4 (b) shows roll vibration of the engine 1. FIG. 4C shows the rotational fluctuation [rpm / sec] of the engine 1.

  According to the present embodiment shown by a solid line in FIG. 4B, both the engine start time (instantaneous 0.5 [sec]) and the stop time (instantaneous 2.0 [sec]) are conspicuous compared with the case where control is not performed (dashed line). The engine roll vibration reduction effect is obtained.

  Further, according to the present embodiment shown by solid lines in FIGS. 4A and 4C, when the engine is not started (instantaneous 0.5 [sec]) and stopped (instantaneous 2.0 [sec]), the control is not performed (broken line). In comparison, a remarkable effect of reducing fluctuations in engine speed is obtained.

  FIG. 5 is a time chart showing a state when the engine torque fluctuation correction control of this embodiment is performed in a specific traveling scene.

  When the vehicle speed, the engine rotation speed, the fuel injection amount, and the motor / generator torque are zero, the driver releases the brake pedal at the instant t1 when the clutch 7 is released, and the engine 1 is allowed to start, the motor / A torque command is given to the generator 5 to start the engine 1. After the instant t1, the engine speed and the motor / generator speed (not shown) increase.

  From this instant t1 to the subsequent instant t2, engine torque fluctuation correction control for starting the engine 1 is performed. At the time of engine start, while cranking torque is input from the motor / generator 5 to the engine 1 until the engine 1 operates independently, a correction value corresponding to the torque fluctuation generated in the engine crankshaft 1a is used for cranking. By adding to the torque basic value, the torque command value applied to the motor / generator 5 is added and corrected.

  Preferably, the coefficient table shown in FIG. 3A is stored separately when the engine is started and when the engine is stopped, which will be described later, and the coefficient K when the engine is started is set as follows. That is, paying attention to the modal force that is the sum of the engine torque and the motor / generator torque obtained by multiplying the engine torque by the predetermined coefficient K, the engine rotation speed 256 [rpm] and the torsion at which the resonance of the roll vibration system occurs Of the engine rotation speed 562 [rpm] at which vibration resonance occurs, the higher value of the engine rotation speed 562 [rpm] and the absolute value of the modal force at the time of engine start from the instant t1 to the instant t2 is the engine described later. The coefficient K is set to be smaller than the absolute value of the modal force at the time of stopping. By setting in this way, vibration system resonance can be suitably reduced at the time of engine start when the engine speed increases.

  When the engine rotational speed reaches a predetermined value (for example, idling rotational speed 1000 [rpm]) at the instant t2, the engine torque fluctuation correction control is terminated and fuel injection is started. After the instant t2, the motor / generator torque becomes zero.

  When the driver depresses the accelerator pedal at the subsequent instant t3, a torque command value to be given to the engine 1 and the motor / generator 5 is calculated based on the accelerator opening. Further, the clutch 7 is engaged after the instant t3, and the vehicle travels at an accelerated speed using the engine torque and the motor / generator torque. The vehicle speed increases from the instant t3 to the subsequent instant t4.

  The vehicle travels at a constant speed from the instant t4 to the subsequent instant t5. During this time, the engine 1 is operated with high efficiency, and the motor / generator 5 generates electric power so as to store the surplus output of the engine 1 in the battery 19. The motor / generator torque becomes negative after the instant t4.

  When the driver releases the accelerator pedal at the instant t5, the accelerator opening becomes zero after the instant t5. Then, fuel injection is stopped after the instant t5, and regenerative torque is input to the motor / generator 5 from the wheel 2 side. That is, the vehicle is decelerated by applying a braking torque from the motor / generator 5 side to the wheel 2 side. The battery 19 is charged using this deceleration energy as electric energy.

  When the vehicle speed reaches a predetermined value (for example, 10 [km / h]) at the subsequent instant t6, the clutch 7 is released. Then, the engine 1 stops at the following instant t7. Also during the instant t6 to t7, the engine torque fluctuation correction control is performed in the same way as the instant t1 to t2 described above. Then, at the following instant t8, the vehicle speed becomes zero.

  Preferably, the coefficient table shown in FIG. 3A is stored separately when the engine is started and when the engine is stopped, which will be described later, and the coefficient K when the engine is started is set as follows. That is, paying attention to the modal force that is the sum of the engine torque and the motor / generator torque obtained by multiplying the engine torque by the predetermined coefficient K, the engine rotation speed 256 [rpm] and the torsion at which the resonance of the roll vibration system occurs The absolute value of the modal force at the time of engine stop from the instant t6 to the instant t7 at the engine revolution speed 256 [rpm] on the lower side of the engine revolution speed 562 [rpm] at which vibration resonance occurs is the engine mentioned above. The coefficient K is set so as to be smaller than the absolute value of the modal force at the start. By setting in this way, vibration system resonance can be suitably reduced when the engine is stopped when the engine rotation speed decreases.

  Next, another embodiment of the present invention will be described.

  The basic configuration of this embodiment is the same as that shown in FIGS. 1 and 2, and the coefficient K at the resonance frequency is obtained according to the equations (1) to (8). The coefficient K is obtained by weighting the displacement of the vibration system and the displacement of the roll vibration system.

R_i:重み
そして、図６（ａ）の重み付二乗和が最小となる係数Ｋをテーブルとして予め記憶しておく。 The weighting method is the same as the previous embodiment in that it is -1 at about 256 rpm where the engine body roll resonance deteriorates, and 0.73 at about 562 rpm where the rotation shaft torsion resonance deteriorates. The coefficient table as shown in FIG. 6A is set so that the weighted square sum J expressed by the following equation is minimized for the mode displacements of the two vibration modes obtained from the equations (5).

R _i : Weight The coefficient K that minimizes the weighted square sum in FIG. 6A is stored in advance as a table.

  FIG. 6B shows a modal force when the coefficient K is obtained with reference to FIG. 6A and torque is output from the motor / generator 5 according to the obtained coefficient K. In FIG. 6B, the solid line indicates the modal force related to the resonance of the system in which the engine case 1s vibrates in the roll direction. A broken line indicates a modal force related to resonance of a system in which the rotation axis is torsionally vibrated in the rotation direction. According to the embodiment shown in FIG. 6, roll vibration and torsional vibration can be reduced as a whole regardless of the engine speed.

  Next, another embodiment of the present invention will be described.

  As a method of determining the weight in this embodiment, a constant weight is given to the displacement of the torsional vibration system relative to the displacement of the roll vibration system regardless of the engine speed. A coefficient K related to the engine speed is stored in advance as a table from a certain weight value, and the coefficient K is obtained by referring to this table. FIG. 7 is a table showing the coefficients K when the weight value of the torsional vibration system is 1, 0.1, and 0.01.

By changing the weight value, it becomes possible to change the vibration characteristic, and the optimum value is determined by performing a simulation or experiment by changing the vibration characteristic.
Here, when one of the weights is increased, the change in the coefficient K increases around the rotation speed 256 [rpm] corresponding to around the resonance frequency in the vibration mode with a small weight value as shown in FIG. When the coefficient K changes periodically, such as when the coefficient K changes suddenly with time during ascending or rotating down, the original engine torque fluctuation correction control frequency changes. And shall be. That is, when the coefficient K periodically fluctuates at the frequency Fk, the engine torque Te at the frequency Fe is amplitude-modulated at the frequency Fk. Therefore, the frequency of the motor / generator torque Tm is Fe + Fk and Fe-Fk. turn into.

α：エンジン回転速度の回転上昇速度
ΔN：係数Ｋが0となる回転速度差
である。図７には重み値0.1の回転速度差ΔNを例示する。 Then, even if the resonance can be reduced in one vibration system, the resonance of the other vibration system which is not going to be reduced before and after the engine rotation speed corresponding to the resonance frequency of one vibration system is excited, The vibration of the vibration system may deteriorate. That is, this vibration deterioration occurs at frequencies around the resonance frequency of one vibration system that reduces resonance. The fluctuation frequency Fk of the coefficient K described above can be estimated from the difference ΔN in engine speed at which the coefficient K becomes 0 before and after the engine speed corresponding to one resonance frequency,

α: Rotational speed increase of engine speed ΔN: Rotational speed difference at which coefficient K becomes zero. FIG. 7 illustrates a rotational speed difference ΔN with a weight value of 0.1.

  Therefore, the weight value of the other vibration system is restricted to a predetermined value or less so that Fe + Fk and Fe-Fk coincide with the resonance frequency of the other vibration system and the vibration of the other vibration system does not deteriorate. More preferably, the weight value is regulated so that Fe + Fk and Fe-Fk do not fall between 1 / √2 times and √2 times the resonance frequency of the other vibration system.

  By restricting the weight value in this way, even if the resonance can be reduced in one vibration system, it is possible to avoid the disadvantage that the vibration of the other vibration system is deteriorated.

  FIG. 8 is a graph showing the maximum displacement amounts of the torsional vibration system and the roll vibration system when starting the engine when the weight value (torsional resonance weight) of the torsional vibration system of the rotating shaft is changed. Indicates the engine speed fluctuation of the torsional vibration system. FIG. 8B shows the amplitude of the roll vibration system. As shown in FIG. 8, as the weight value increases, the engine roll amplitude of the roll vibration system increases to a degree that cannot be ignored. For this reason, in this embodiment, the weight value of the torsional vibration system is restricted to 0.1 or less.

  The vibration reduction effect of this example is shown by a solid line in FIG. Further, for comparison, a case where the engine torque fluctuation correction control of this embodiment is not performed is indicated by a broken line. FIG. 9A shows the engine speed when the engine 1 is started and then stopped. 9 (b) shows the roll vibration of the engine 1 (engine case portion 1s). FIG. 9C shows the rotational fluctuation [rpm / sec] of the engine 1.

  According to the present embodiment shown by a solid line in FIG. 9B, both the engine start time (instantaneous 0.5 [sec]) and the stop time (instantaneous 2.0 [sec]) are conspicuous compared with the case where control is not performed (dashed line). The engine roll vibration reduction effect is obtained.

  Further, according to the present embodiment shown by solid lines in FIGS. 9A and 9C, when the engine is not started (instantaneous 0.5 [sec]) and stopped (instantaneous 2.0 [sec]), the control is not performed (broken line). In comparison, a remarkable effect of reducing fluctuations in engine speed is obtained.

  Further, in this embodiment, as shown in FIG. 8 (a), the weight value of the torsional vibration system is restricted to 0.1 or less. Therefore, the coefficient K is applied to the first engine speed 256 [rpm] at which roll vibration system resonance occurs. The motor / generator torque upper and lower frequencies obtained by adding and subtracting the fluctuation frequency of the engine do not match the engine speed 562 [rpm] at which torsional vibration resonance occurs. Therefore, when the resonance is reduced in the roll vibration system, it is possible to avoid deterioration of the vibration of the torsional vibration system.

  Next, another embodiment of the present invention will be described.

  FIG. 10 shows a power train of a front engine / rear wheel drive hybrid vehicle including the vibration control apparatus according to this embodiment. Since the basic configuration of the power train in FIG. 10 is the same as that of the power train in FIG. 1 described above, description of the common parts will be omitted, and different parts will be described with new reference numerals.

  In the embodiment shown in FIG. 10, instead of the automatic transmission 3, the motor / generator 5 and the clutch 7 shown in FIG. 1, the first motor / generator 21, the planetary gear set 23 and the second motor / A generator 24 is mounted. The first motor / generator 21 is mainly used for power generation and engine start. The second motor / generator 24 mainly performs vehicle driving and regeneration.

  As shown in FIG. 10, the output shaft of the torsional damper 6 is coupled to the carrier of the planetary gear set 23. The rotation shaft 22 of the first motor / generator 21 is coupled to the sun gear of the planetary gear set 23. The ring gear of the planetary gear set 23 is coupled to one end of the shaft 25. The shaft 25 also serves as the shaft of the second motor / generator 24. The other end of the shaft 25 protrudes from the case 11 and is connected to the differential gear device 8.

  The first motor / generator 21 is connected to the first inverter 17 via a power cable, and the first inverter 17 is connected to the battery 19 via a power cable. The first inverter 17 drives the shaft 22 with the target first motor / generator torque by controlling the current supplied to the first motor / generator 21 in accordance with the torque command from the driving force controller 16.

  The second motor / generator 24 is connected to the second inverter 26 by a power cable, and the second inverter 26 is connected to the battery 19 by a power cable. The second inverter 26 drives the shaft 25 with the target second motor / generator torque by controlling the current supplied to the second motor / generator 24 in accordance with the torque command from the driving force controller 16.

The hybrid system of the present embodiment is an electric transmission that continuously controls the reduction ratio from engine rotation to axle 25 rotation by controlling the rotation speed of the first motor / generator 21, and when the engine is stopped. Even so, by appropriately controlling the rotational speed of the first motor / generator 21, the vehicle can be driven by the driving force of the second motor / generator 24. The engine 1 to the axle 25 do not include a clutch or the like. , Always in a mechanically coupled state. Therefore, at the time of engine start and engine stop, if the engine rotational speed passes through the torsional resonance frequency of the rotating shaft by the torsional damper system and the torsional vibration of the rotating shaft deteriorates, this causes the rotation of the tire of the wheel 2 As a result, the vehicle longitudinal vibration deteriorates.

Similarly, when the engine is started and when the engine is stopped, the engine rotational speed also passes through the roll resonance frequency of the engine body by the engine mount system, so that the engine roll resonance deteriorates and this is transmitted to the vehicle body via the engine mount 9. It is transmitted and worsens the body vibration.
In order to prevent such deterioration of vibration, in this embodiment, correction control for engine torque fluctuations by the first motor / generator 21 and the second motor / generator 24 is performed when the engine is started and when the engine is stopped.

  FIG. 11 is an explanatory diagram showing a vibration model formed by the power train of the hybrid vehicle shown in FIG.

In the present embodiment, the rotating part 1r in the engine 1, the rotor part 21r of the first motor / generator 21, the rotor part 24r of the second motor / generator 24, and the engine case part 1s are mass bodies, and a torsional damper. 6 and the engine mount 9 serve as spring elements. The mass body of the engine rotating portion 1r includes the engine crankshaft 1a, and the mass body of the rotor 21r of the first motor / generator 21 includes the shaft 22. The mass body of the engine case portion 1 s includes a non-rotating portion outside the engine 1, a stator portion of the first motor / generator 21, a stator portion of the second motor / generator 24, and the extending case 11.
Furthermore, when considering the vibration model of the power train, the tires of the drive shaft 20 and the wheel 2 are evaluated by the torsion spring constant Kt and the damping constant Ct. In FIG. 11, 31 is a load.

ここで、
ｒ₁は遊星歯車組２３のサンギヤ半径、ｒ₄は遊星歯車組２３のリングギヤ（インターナルギヤ）半径、ｒ_tは車輪２のタイヤ半径、αはディファレンシャルギヤ装置8のファイナルギヤ比である。 The engine torque fluctuation correction control for the first motor / generator 21 and the second motor / generator 24 will be described.
In the present embodiment, as shown in FIG. 11, a vibration system composed of the inertia moment Ie of the engine case 1s, the spring Ke and the damping Ce of the engine mount, the inertia moment Ic of the engine rotating portion 1r, and the inertia moment Ii of the planetary carrier. , The inertia moment Im1 of the rotor 21r of the first motor / generator, the inertia moment Im2 of the rotor 24r of the second motor / generator 24, the vehicle mass Mv, the spring Kc and damping Cc of the torsion damper, the wheel tire 2 and the drive shaft 20 The torsion spring Kt and the damping system Ct are configured so that the engine torque Te is input to the engine rotating portion 1r, the reaction force is input to the engine case portion 1s, and the first motor / generator torque Tm1 is input to the rotor 21r. In addition, the second motor / generator torque Tm2 is input to the rotor 24r and the reaction force is Input to the engine body.
The equation of motion of this vibration system includes the rotation angle θe of the engine case portion 1s, the rotation angle θc of the engine rotation portion 1r, the planetary carrier rotation angle θi, the rotation angle θm1 of the rotor portion 21r, the rotation angle θm2 of the rotor portion 24r, and the vehicle longitudinal displacement. as described by x _v, erasing the θm1 by the relationship of the gear ratio of the planetary gear from these represented by the following equation of motion.

here,
r ₁ is the sun gear radius, r ₄ of the planetary gear set 23 is a ring gear (internal gear) radius of the planetary gear set 23, the r _t tire radius of the wheel 2, alpha is a final gear ratio of the differential gear unit 8.

ここで最初の2つの振動モードは回転軸の回転に関する剛体系、3番目の振動モードはエンジンケース部１ｓのロール共振、4、5番目の振動モードは回転軸ねじり共振を示している。ここで、振動として問題となるのは3,4,5番目のエンジンケース部１ｓのロール共振及び2つの回転軸ねじり共振であり、これら3つの振動モードに関して、第1および第2の2つのモータ／ジェネレータ２１，２４を制御することにより低減する。 By performing the eigenvalue analysis of Equation 11 above in the same manner as Equation 2 in the first embodiment, the resonance frequency Fn and mode vector φ of this vibration system are as follows.
In this embodiment, the resonance frequency Fn and mode vector φ of this vibration system are as follows.

Here, the first two vibration modes indicate a rigid system related to the rotation of the rotating shaft, the third vibration mode indicates the roll resonance of the engine case portion 1s, and the fourth and fifth vibration modes indicate the rotation axis torsional resonance. Here, the problems of vibration are the roll resonance and the two rotational shaft torsional resonances of the third, fourth, and fifth engine case sections 1s. The first and second motors are related to these three vibration modes. / Reduce by controlling generators 21,24.

  Specifically, as in the embodiment of FIG. 2 described above, at the engine rotation speed at which the rotation third-order explosion frequency matches the resonance frequency of the vibration system, the control torque is set so that the modal force of the resonance becomes zero. Generated from the first and second motor / generators 21, 24 to avoid resonance excitation. That is, the engine torque Te is estimated from the detected rotational speed of the engine 1 and the crank angle of the engine crankshaft 1 a according to the map stored in the driving force controller 16. Then, the coefficients K1 and K2 for the first and second motor / generators 21 and 24 are stored in the driving force controller 16 as a coefficient table with respect to the engine speed, and the coefficient K1, with reference to the coefficient table according to the engine speed, Determine K2. The coefficients K1 and K2 are determined so that the modal force of each vibration system becomes 0 at about 256 rpm at which the engine case roll resonance deteriorates and at about 300 rpm and 618 rpm at which the rotational shaft torsion resonance deteriorates. Excitation of resonance due to engine torque fluctuation is prevented by the control torque.

  Regarding the engine rotational speed other than resonance, the coefficient table as shown in FIG. 12 is determined so that the weighted sum of squares of these three vibration modes is minimized.

  This control is performed when the engine is started and when the engine is stopped. Since the vibration modes excited are different between when the engine is started and when the engine is stopped, different coefficient tables are used when the engine is started and when the engine is stopped as shown in FIG. I am doing so. That is, at the initial stage of rotation when the engine is started, the amount of intake air of the engine is small, so that the torque fluctuation is small, and the low-frequency vibration mode that passes at an early timing is not so excited. In addition, since the rotational speed is high, low-frequency vibrations are not excited so much and the contribution of the relatively high-frequency vibration mode is relatively large. For this reason, the coefficient table shown in FIG.

  On the other hand, when the engine is stopped, there is a sufficient amount of intake air even at a low speed, and the rotation lowering speed is slower than the rotation increasing speed at the time of starting the engine. The contribution of the high frequency vibration mode is increased. For this reason, the coefficient table shown in FIG.

  According to such a present Example, regarding the weight value at the time of determining a coefficient table, the weight value of the vibration mode of a low frequency becomes large at the time of engine starting. Thereby, a modal force as shown in FIG. 13 can be obtained. In FIG. 13, (a) shows the modal force of engine roll resonance. Moreover, (b) shows the modal force of the rotational axis torsional resonance with the lower frequency. Furthermore, (c) shows the modal force of the rotational axis torsional resonance with the higher frequency.

  According to this embodiment, in the engine roll resonance (FIG. 13 (a)) and the low-frequency rotary shaft torsion resonance (FIG. 13 (b)), the modal force is more or less before and after the resonance rotational speed when the engine is stopped. The effect that a small range becomes wide can be acquired. Further, in the high-frequency rotating shaft torsional resonance with a high frequency (FIG. 13C), it is possible to obtain an effect that the modal force becomes smaller before and after the resonance rotational speed when the engine is started.

  Since the second motor / generator 24 also drives the vehicle, the magnitude of torque that can be used for torque fluctuation correction control varies depending on the torque required for driving the vehicle. As the torque of the second motor / generator 24 changes, the optimum torque of the first motor / generator 21 also changes. Therefore, the coefficient table of the first motor / generator 21 and the second motor / generator 24 depends on the vehicle driving torque. Be changed. That is, these coefficients are stored in the driving force controller 16 as a two-dimensional map regarding the engine speed and the vehicle driving torque of the second motor / generator 24.

  FIG. 14 shows the effect of vibration reduction by performing the engine torque fluctuation correction control that corrects the engine torque fluctuation with the motor / generator torque based on the coefficient tables of FIGS. 12 (a) and 12 (b).

In FIG. 14, the solid line indicates the vibration reduction effect by performing the above-described engine torque fluctuation correction control. Further, for comparison, a case where the engine torque fluctuation correction control of this embodiment is not performed is indicated by a broken line. FIG. 14A shows the engine speed when the engine 1 is started and then stopped. 14 (b) represents the amplitude of roll vibration of the engine 1. FIG. 14C shows vehicle longitudinal vibration [m / sec ² ] of a vehicle equipped with the power train of this embodiment.

  According to the present embodiment shown by a solid line in FIG. 4A, both the engine start time (instantaneous 0.5 [sec]) and the stop time (instantaneous 2.0 [sec]) are conspicuous compared with the case where control is not performed (dashed line). The effect of reducing the engine speed is obtained.

  Further, according to the present embodiment shown by a solid line in FIG. 14B, compared to the case where the control is not performed (dashed line) both at the time of engine start (instantaneous 0.5 [sec]) and at the time of stop (instantaneous 2.0 [sec]), A remarkable engine roll vibration reduction effect is obtained.

  Further, according to the present embodiment shown by a solid line in FIG. 14 (c), both when the engine is started (instantaneous 0.5 [sec]) and when stopped (instantaneous 2.0 [sec]), compared to the case where control is not performed (dashed line), A significant vehicle longitudinal vibration reduction effect is obtained.

  By the way, according to each of the above-described embodiments, the driving force controller 16 generates the torque obtained by multiplying the engine torque variation of the engine 1 by the predetermined coefficient K from the motor / generator 5 or the like as shown in the equation (8). At this time, the driving force controller 16 determines that the explosion frequency corresponding to the engine torque fluctuation determined from the number of engine explosions per one rotation of the engine crankshaft 1a and the engine rotation speed is coupled to the engine crankshaft 1a and the motor / The resonance frequency of the torsional vibration system of the output shaft that rotates relative to the generator rotation shaft 4 or the like substantially matches the resonance frequency of the roll vibration system that is the relative displacement of the engine case 1s relative to the vehicle body 10 with respect to the engine mount 9. When they substantially coincide with each other, a predetermined coefficient K is obtained according to the engine speed so that the excitation force for exciting the resonance of these vibration systems is minimized. As a result, the torsional vibration of the output shafts 1a, 4 and 3a and the roll vibration of the engine case 1s according to the hybrid vehicle can be reduced as shown in FIGS. Further, when the predetermined coefficient K with respect to the engine rotation speed is stored as a simple table illustrated in FIGS. 3, 6, and 12, the drive control memory can be saved.

In each of the above-described embodiments, the engine rotation speed is equivalent to the resonance frequency of the torsional vibration system (562 [rpm] or 300 rpm) using the coefficient tables illustrated in FIGS. [rpm], 618 [rpm]) and the engine rotational speed (256 [rpm]) corresponding to the resonance frequency of the roll vibration system, which is the displacement of the engine 1, the coefficient according to the engine rotational speed From finding K continuously,
It is possible to prevent the deterioration of vibration even during a transient operation in which the engine speed changes.

  In the above-described embodiment, as illustrated in the coefficient table of FIG. 7, the coefficient K may be obtained by weighting the displacement of the torsional vibration system and the displacement of the roll vibration system for each engine rotation speed. It is possible to change the vibration characteristics and set an optimum coefficient table.

Further, in the above-described embodiment, the coefficient table of FIG. 6 is set so that the weighted square sum of the displacement of the torsional vibration system and the displacement of the roll vibration system is minimized based on the equation (9), and the coefficient is set for each engine speed. From seeking K,
Both resonance of the torsional vibration system related to the torsion damper 6 and resonance of the roll vibration system related to the engine mount 9 can be reduced.

Further, in each of the above-described embodiments, as shown in the time chart of FIG. 5, when the engine is started to start the operation of the engine 1 (instantaneous t1 to t2) and when the engine is stopped to finish the operation (instantaneous t6 to t7). ) Is multiplied by the coefficient K from the motor / generator 5 or the like,
The resonance of the torsional vibration system and the resonance of the roll vibration system that are most likely to occur when the engine is running at a low speed, such as when the engine is started and when the engine is stopped, can be appropriately reduced.

  Preferably, FIG. 12 (a) and FIG. 12 (b) may be set separately so that the coefficient K when the engine is started and the coefficient K when the engine is stopped are different.

Specifically, based on the equation (6), the modal force that is the sum of the excitation force due to the engine torque of each vibration system and the excitation force due to the motor / generator torque obtained by multiplying the engine torque by a coefficient K. (7) The coefficient K is set so that the absolute value of the modal force is minimized at the resonance frequency of the vibration system. As shown in FIGS. 13 (a) and 13 (b), the engine rotation speed 256 [rpm] on the lower side of the engine rotation speed 256 [rpm] and the engine rotation speed 256 [rpm] at which the resonance of the vibration system occurs occurs. The coefficient K is set as shown in FIG. 12 so that the absolute value of the modal force at the time of stop becomes smaller than the absolute value of the modal force at the time of engine start.
As a result, a low-frequency vibration mode that is particularly likely to occur when the engine is stopped can be reduced. In addition, a high-frequency vibration mode that is particularly likely to occur when the engine is started can be reduced.

  By the way, if the coefficient K obtained by referring to the coefficient table changes periodically due to a sudden change in the engine speed or the like, there is a concern that the vibration of the torsional vibration system deteriorates while the resonance of the roll vibration system is reduced. Is done. Specifically, as shown by the solid line in FIG. 7, when the weight of the torsional vibration system with respect to the roll vibration system is 1, when the engine speed increases and decreases around the roll resonance frequency 256 [rpm], the coefficient K is about − This is a case where the period periodically varies between 0.8 and 0.7. In this case, since the engine torque of the explosion frequency corresponding to the roll resonance frequency 256 [rpm] is amplitude-modulated, there is a concern that the resonance of the roll vibration that becomes the other vibration system is excited.

Therefore, in this embodiment, the weighting is restricted to a predetermined range. Specifically, the motor / generator torque upper frequency Fn + Fk obtained by adding or subtracting the fluctuation frequency Fk of the coefficient K calculated by the equation (10) to the resonance frequency Fn related to the engine rotational speed 256 [rpm] at which the resonance of the roll vibration system occurs. In addition, the weighting is restricted to 0.1 or less so that the lower frequency Fn−Fk does not coincide with the resonance frequency of the torsional vibration system (FIG. 8).
By regulating the weighting in this way, even if the resonance can be reduced in one vibration system, it is possible to avoid the disadvantage that the vibration of the other vibration system deteriorates.

  More preferably, the weighting is regulated so that the motor / generator torque upper frequency Fn + Fk and the lower frequency Fn−Fk are 1 / √2 times or less of the resonance frequency of the torsional vibration system and √2 times or more. Good. Thereby, even if resonance can be reduced in one vibration system, it is possible to reliably avoid the disadvantage that the vibration of the other vibration system deteriorates.

The embodiment shown in FIG. 10 includes a first motor / generator 21 and a second motor / generator 24. As shown in FIG. 12, coefficients K1 and K2 are obtained for each of the motor / generators 21 and 24.
Even in a hybrid vehicle having a plurality of motors / generators, torsional vibration and roll vibration can be reduced.

Preferably, the coefficient table used for determining the coefficients K1 and K2 of the plurality of motor / generators 21 and 24 is a two-dimensional map that differs depending on the magnitude of the vehicle driving torque.
As a result, even if the magnitude of the torque that can be used for the engine torque fluctuation correction control changes due to the torque required for driving the wheels 2, the modal force of the vibration system can be appropriately minimized.

  The above description is merely an example of the present invention, and the present invention can be variously modified without departing from the spirit of the present invention. For example, the present invention is not limited to a six-cylinder engine and can be widely applied from a single cylinder to a multi-cylinder engine.

1 is a shift control system diagram showing a power train of a hybrid vehicle including a vibration control device according to an embodiment of the present invention together with its control system. It is explanatory drawing which shows the vibration model which the same power train comprises. (A) is a coefficient table used in the embodiment, and (b) is a characteristic diagram showing a modal force corresponding to the coefficient table. It is explanatory drawing which shows the vibration reduction effect by the engine torque variation correction control of the same Example, (a) shows an engine rotational speed, (b) shows an engine roll amplitude, (c) shows engine rotational speed. Showing fluctuations. It is a time chart which shows a mode when performing engine torque variation correction control of a present Example. (A) is a coefficient table used by another Example, (b) is a characteristic view which shows the modal force corresponding to the same coefficient table. It is a coefficient table used in another Example. It is a graph which shows the displacement amount in the predetermined | prescribed engine rotational speed when changing the weight value (torsional resonance weight) of a torsional vibration system, (a) is an engine rotational speed fluctuation | variation of a torsional vibration system, (b) is. The amplitude of the roll vibration system is shown. It is explanatory drawing which shows the vibration reduction effect by the engine torque variation correction control of the same Example, (a) shows an engine rotational speed, (b) shows an engine roll amplitude, (c) shows engine rotational speed. Showing fluctuations. It is a shift control system figure which shows the power train of the hybrid vehicle provided with the vibration control apparatus which becomes another Example of this invention with the control system. It is explanatory drawing which shows the vibration model which the same power train comprises. FIG. 4 is a coefficient table used in the embodiment, (a) shows a coefficient table used when the engine is started, and (b) shows a coefficient table used when the engine is stopped. It is a characteristic view which shows the modal force corresponding to the same coefficient table, (a) shows the modal force of engine roll resonance. Moreover, (b) shows the modal force of the rotational axis torsional resonance with the lower frequency. Furthermore, (c) shows the modal force of the rotational axis torsional resonance with the higher frequency. It is explanatory drawing which shows the vibration reduction effect by the engine torque variation correction control of the same Example, (a) shows an engine rotational speed, (b) shows an engine roll amplitude, (c) shows vehicle longitudinal vibration. Show.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine 1a Engine crankshaft 1r Engine rotation part 1s Engine case part 2 Wheel 3 Automatic transmission 4 Rotor shaft 5 Motor / generator 5r Rotor part 6 Torsional damper 7 Clutch 8 Differential gear device
11 Engine extension case
12 Crank angle sensor
13 Rotational speed sensor
14 Accelerator sensor
15 Brake sensor
16 Driving force controller
20 Drive shaft
21 1st motor / generator
22 Rotor shaft of the first motor / generator
23 Planetary gear set
24 Second motor / generator
25 Output shaft

Claims (12)

Comprising a engine and the motor / generator, through the torsion damper to dampen the rotational direction force by combining the rotation axis of the output shaft and the motor / generator of the engine, the power via an elastically deformable mounting member In a hybrid vehicle with the source supported on the vehicle body side,
Motor / generator torque command means for generating torque obtained by multiplying the engine torque variation of the engine by a predetermined coefficient from the motor / generator;
The explosion frequency corresponding to the engine torque fluctuation determined from the number of engine explosions per one rotation of the engine output shaft and the engine rotation speed is the relative rotation between the engine output shaft and the motor / generator rotation shaft with respect to the torsion damper means. When the torsional vibration system of the power source and the displacement vibration system that is the relative displacement of the power source with respect to the vehicle body with respect to the mount member substantially match the resonance frequency, the excitation force that excites the resonance of these vibration systems is Predetermined coefficient calculation means for obtaining the predetermined coefficient in accordance with the engine speed so as to be minimized;
A vibration control device for a hybrid vehicle, comprising: The vibration control device for a hybrid vehicle according to claim 1,
The predetermined coefficient calculating means is configured to change the engine speed when the engine speed changes between an engine speed corresponding to a resonance frequency of the torsional vibration system and an engine speed corresponding to the resonance frequency of the displacement vibration system. A vibration control apparatus for a hybrid vehicle, wherein the predetermined coefficient is continuously obtained according to a rotational speed. The vibration control device for a hybrid vehicle according to claim 2,
The hybrid vehicle vibration control apparatus according to claim 1, wherein the predetermined coefficient calculation means obtains the predetermined coefficient by weighting the displacement of the torsional vibration system and the displacement of the displacement vibration system for each engine speed. The vibration control device for a hybrid vehicle according to claim 3,
The predetermined coefficient calculating means obtains the predetermined coefficient so that a weighted sum of squares of the displacement of the torsional vibration system and the displacement of the displacement vibration system is minimized for each engine speed. Vibration control device. In the vibration control device for a hybrid vehicle according to any one of claims 1 to 4,
The motor / generator torque command means generates torque multiplied by the predetermined coefficient from the motor / generator at least one of when the engine is started to start the engine and when the engine is stopped to end the operation of the engine. A vibration control device for a hybrid vehicle characterized by the above. In the hybrid vehicle vibration control device according to claim 5,
The vibration control apparatus for a hybrid vehicle, wherein the predetermined coefficient when the engine is started differs from the predetermined coefficient when the engine is stopped. The vibration control apparatus for a hybrid vehicle according to claim 6,
The predetermined coefficient calculation means includes an excitation force that excites resonance by the engine torque in one of the vibration systems, and an excitation force that excites resonance by a motor / generator torque obtained by multiplying the engine torque by the predetermined coefficient. For the modal force that is the sum of the above, the predetermined coefficient is set so that the absolute value of the modal force is minimized at the resonance frequency of the vibration system,
The absolute value of the modal force when the engine is stopped is smaller than the absolute value of the modal force when the engine is started at a lower engine rotational speed among two or more engine rotational speeds at which the vibration system resonance occurs. The vibration control apparatus for a hybrid vehicle, wherein the predetermined coefficient is set to be The vibration control apparatus for a hybrid vehicle according to claim 6 or 7,
The predetermined coefficient calculation means includes an excitation force that excites resonance by the engine torque in one of the vibration systems, and an excitation force that excites resonance by a motor / generator torque obtained by multiplying the engine torque by the predetermined coefficient. For the modal force that is the sum of the above, the predetermined coefficient is set so that the absolute value of the modal force is minimized at the resonance frequency of the vibration system,
The absolute value of the modal force at the time of starting the engine is smaller than the absolute value of the modal force at the time of stopping the engine at a higher engine speed out of two or more engine speeds at which resonance of the vibration system occurs. The vibration control apparatus for a hybrid vehicle, wherein the predetermined coefficient is set to be The vibration control device for a hybrid vehicle according to claim 3,
The predetermined coefficient calculation means regulates the weighting to a predetermined range,
When the weighted predetermined coefficient continuously changes in order to increase or decrease the engine rotation speed across the first engine rotation speed related to the engine rotation speed at which resonance of one of the vibration systems occurs ,
The motor / generator torque upper frequency and lower frequency obtained by adding / subtracting a predetermined coefficient fluctuation frequency equal to the periodic change of the predetermined coefficient to the resonance frequency related to the first engine rotation speed,
The vibration control device for a hybrid vehicle, wherein the weighting is restricted to a predetermined range so as not to coincide with a resonance frequency of the other vibration system. The vibration control apparatus for a hybrid vehicle according to claim 9,
The predetermined coefficient calculation means is configured so that the motor / generator torque upper frequency and lower frequency are equal to or less than 1 / √2 times the resonance frequency of the other vibration system. The vibration control apparatus for a hybrid vehicle, wherein the weighting is regulated so that the resonance frequency becomes √2 times or more. The vibration control device for a hybrid vehicle according to claim 1,
A hybrid vehicle vibration control device comprising a plurality of the motor / generators, wherein the predetermined coefficient calculating means obtains the predetermined coefficient for each motor / generator. The vibration control device for a hybrid vehicle according to claim 11,
The vibration control apparatus for a hybrid vehicle, wherein the predetermined coefficient of the plurality of motors / generators varies depending on the magnitude of the driving force.

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