Electric vehicle control method and electric vehicle control device

Abstract

To provide an electric vehicle control method and an electric vehicle control device capable of reducing a sense of discomfort felt by an occupant due to variations in the dead zone interval lengths of driving systems in an electric vehicle including a plurality of driving motors.SOLUTION: An electric vehicle control method includes: basic torque distribution processing for determining basic torque command values for respective driving motors; and vibration damping processing for correcting each basic torque command value to suppress the vibration of a driving force transmission system, thereby obtaining a corrected torque command value. In the vibration damping processing, a damping coefficient in a model response related to driving force transmission in each driving system is set as a parameter which takes a reference damping coefficient when the driving system is in an interval other than a dead zone interval, and which takes a dead zone damping coefficient different from the reference damping coefficient when the driving system is in the dead zone interval. It is estimated whether each individual driving system is in the dead zone interval. In the driving system in the interval other than the dead zone interval, the amount of correction for the basic torque command value is set on the basis of the reference damping coefficient. In the driving system in the dead zone interval, the amount of correction for the basic torque command value is set on the basis of the dead zone damping coefficient.SELECTED DRAWING: Figure 11

Description

[Technical Field]

[0001] The present invention relates to an electric vehicle control method and an electric vehicle control device. [Background technology]

[0002] Patent Document 1 discloses a control method for suppressing vibrations in the drive force transmission system of an electric vehicle that can be driven using torque from an electric motor, taking into account the transmission characteristics of the power transmission mechanism connected between the output shaft of the drive motor and the drive wheels. In particular, this control method estimates a dead zone in which drive motor torque is not transmitted to the drive shaft torque of the electric vehicle, and suppresses shocks caused by gear backlash by performing processing to speed up the response of the drive shaft torque in the dead zone. [Prior art documents] [Patent documents]

[0003] [Patent Document 1] International Publication No. 2017 / 183231 Summary of the Invention [Problem to be solved by the invention]

[0004] The inventors focused on the fact that in an electric vehicle equipped with multiple drivetrains, each with its own drive motor, which drive different drive wheels, the timing at which gear backlash is eliminated varies for each drivetrain due to differences in the mechanical characteristics and drive force distribution of each drivetrain, which can result in variations in the length of the dead zone. For this reason, even if a conventional electric vehicle control method is applied to an electric vehicle equipped with multiple drivetrains, the torque response performance of the electric vehicle in the dead zone varies depending on the driving state, which can cause discomfort to occupants.

[0005] The present invention aims to provide an electric vehicle control method and an electric vehicle control device that can reduce the discomfort felt by occupants due to variations in the length of the dead zone sections of each drive system in an electric vehicle equipped with multiple drive motors. [Means for solving the problem]

[0006] One aspect of the present invention is a method for controlling the driving force of each drivetrain in an electric vehicle equipped with multiple drivetrains, each equipped with a drive motor. The method includes a basic torque allocation process for determining a basic torque command value for each drivetrain based on a total required driving force for the electric vehicle and a driving force allocation to each drivetrain. A vibration damping process for correcting each basic torque command value to determine a corrected torque command value for suppressing vibration in the driving force transmission system and a driving force control process for controlling the driving force generated by each drivemotor based on the corrected torque command value. The vibration damping process sets different damping coefficients for a reference response related to driving force transmission in each drivetrain as parameters: a reference damping coefficient that is a damping coefficient when the drivetrain is in a range other than a dead-band range; and a dead-band damping coefficient that is a damping coefficient when the drivetrain is in the dead-band range. Each drivetrain is individually estimated to be in the dead-band range, and a correction amount for the basic torque command value is set based on the reference damping coefficient for drivetrains outside the dead-band range, and a correction amount for the basic torque command value is set based on the dead-band damping coefficient for drivetrains in the dead-band range. When the combination of mechanical characteristics of each drivetrain changes, the dead-band damping coefficient is set in accordance with the combination of mechanical characteristics. Alternatively, when the drive force distribution of each drivetrain changes, the dead-band damping coefficient is set in accordance with the drive force distribution. [Effects of the Invention]

[0007] According to the present invention, it is possible to provide an electric vehicle control method and an electric vehicle control device that can reduce the discomfort felt by occupants due to variations in the length of the dead zone sections of each drive system in an electric vehicle equipped with multiple drive motors. [Brief explanation of the drawings]

[0008] [Figure 1] FIG. 1 is a block diagram illustrating the main configuration of an electric vehicle. [Figure 2]FIG. 2 is a flowchart showing the flow of processing in the electric vehicle control method. [Figure 3] FIG. 3 is a diagram showing an example of an accelerator opening-torque table. [Figure 4] FIG. 4 is a block diagram illustrating the front and rear torque distribution process. [Figure 5] FIG. 5 is a block diagram illustrating the vibration suppression process according to the first embodiment. [Figure 6] FIG. 6 is a block diagram illustrating the F / F compensator of the front drive train according to the first embodiment. [Figure 7] FIG. 7 is a block diagram illustrating the F / F compensator of the rear drivetrain according to the first embodiment. [Figure 8] FIG. 8 is a diagram showing a model of a driving force transmission system of an electric vehicle (4WD). [Figure 9] FIG. 9 is a diagram illustrating characteristics such as torque response of an electric vehicle according to a combination of mechanical characteristics (gear ratios) of each drivetrain. [Figure 10] FIG. 10 is a speed ratio-dead zone damping coefficient table showing the settings of the dead zone damping coefficient according to the combination of the mechanical characteristics (speed ratio) of each drive system. [Figure 11] FIG. 11 is an explanatory diagram showing the relationship between the front torque correction unit and the setting of the dead-band damping coefficient. [Figure 12] FIG. 12 is a block diagram illustrating the F / B compensator of the front drive train. [Figure 13] FIG. 13 is a block diagram illustrating the F / B compensator of the rear drive train. [Figure 14] FIG. 14 is a time chart illustrating the control results of the first embodiment. [Figure 15] FIG. 15 is a diagram illustrating the torque response characteristics of an electric vehicle according to the driving force distribution. [Figure 16] FIG. 16 is a diagram showing a driving force distribution ratio-dead zone damping coefficient table. [Figure 17] FIG. 17 is an explanatory diagram showing the relationship between the front torque correction unit and the driving force distribution ratio-dead zone damping coefficient table. [Figure 18] FIG. 18 is a time chart illustrating the control results of the second embodiment. [Figure 19] FIG. 19 is a block diagram illustrating a vibration suppression process according to a modified example. [Figure 20] FIG. 20 is a block diagram illustrating a modified F / F compensator. [Figure 21] FIG. 21 is a diagram showing a modified example of the configuration of an electric vehicle. [Figure 22] FIG. 22 is a diagram showing a modified example of the configuration of an electric vehicle. [Figure 23] FIG. 23 is a diagram showing a modified example of the configuration of an electric vehicle. DETAILED DESCRIPTION OF THE INVENTION

[0009] Hereinafter, an embodiment of the present invention will be described with reference to the drawings.

[0010] [First embodiment] FIG. 1 is a block diagram illustrating the main configuration of an electric vehicle 100. In this embodiment, the electric vehicle 100 refers to a vehicle such as an electric vehicle (EV) or a hybrid vehicle (HEV) equipped with one or more electric motors (drive motors) as a driving source. In particular, the electric vehicle 100 of this embodiment includes a front drivetrain S including a drive motor 4 (hereinafter referred to as a front motor 4f or simply as a motor 4f) that applies driving force to the front drive wheels 9fR and 9fL. f and a rear drive system S including a drive motor 4 (hereinafter referred to as a rear motor 4r or simply as a motor 4r) that applies drive force to the rear drive wheels 9rR and 9rL. r It is equipped with:

[0011] For the sake of simplicity, the following description will be limited to the front drivetrain S. f and rear drive system S r For each component in the f ,S r In other words, unless otherwise specified, the following description will refer to the front drivetrain Sf and rear drive system S r This is common to each component of the

[0012] The battery 1 is configured as an on-board secondary battery that can supply (discharge) driving power when the motors 4f, 4r are in power running mode and can receive (charge) regenerated power when the motors 4f, 4r are in regenerative running mode.

[0013] Motor controller 2 has vehicle speed V and accelerator opening A PO , rotor phase α of motors 4f and 4r f ,α r , the motor current i which flows through the motors 4f and 4r f ,i r Signals indicating various vehicle variables such as the above are input as digital signals. The motor controller 2 generates PWM signals for controlling the motors 4f, 4r based on the input signals. The motor controller 2 also generates drive signals for the inverters 3f, 3r in response to the generated PWM signals.

[0014] The inverters 3f, 3r convert the DC current supplied from the battery 1 into AC by turning on / off two switching elements (e.g., power semiconductor elements such as IGBTs and MOS-FETs) provided for each phase, and pass the desired current through the motors 4f, 4r.

[0015] The motors 4f and 4r generate driving force by AC current supplied from the inverters 3f and 3r, and the driving force is transmitted to the transmissions 5f and 5r and the drive shaft DS f ,DS r The motors 4f and 4r transmit driving force to the drive wheels 9fR, 9fL, 9rR, and 9rL via a driving force transmission system consisting of the motors 4f and 4r. Furthermore, the motors 4f and 4r recover, as electrical energy, kinetic energy based on the regenerative braking force received from the drive wheels 9fR, 9fL, 9rR, and 9rL when the vehicle is running. In this case, the inverter 3 converts the AC current generated during regenerative operation of the motor 4 into DC current and supplies it to the battery 1. In particular, the front motor 4f supplies driving force to the front drive wheels 9fR and 9fL, and the rear motor 4r supplies driving force to the rear drive wheels 9rR and 9rL.

[0016] The current sensor 20 detects the motor current i f ,i r (Especially the front three-phase AC current i uf ,i vf ,i wf and rear three-phase AC current i ur ,i vr ,i wr ) is detected. However, the three-phase AC current i u ,i v ,i w Since the sum of these is 0, the currents of any two phases may be detected and the current of the remaining phase may be calculated.

[0017] The rotation sensor 21 is, for example, a resolver or an encoder, and detects the rotor phase α of the motors 4f and 4r. f ,α r Detect.

[0018] Fig. 2 is a flowchart showing each process in the electric vehicle control method programmed to be executed by the motor controller 2. Note that the processes relating to steps S201 to S205 shown in Fig. 2 are repeatedly executed at a predetermined calculation cycle while the vehicle system is running.

[0019] In step S201, the motor controller 2 executes input processing. Specifically, the motor controller 2 receives signals indicating the various vehicle conditions described above as vehicle information. In particular, the signals indicating the vehicle conditions include the vehicle speed V (km / h), the accelerator opening A PO (%), rotor phase α f ,α r (rad), the motor rotation speed N mf ,N mr (rpm), motor current i f ,i r , and the DC voltage value V of battery 1 dc (V) and others.

[0020] The vehicle speed V (km / h) is acquired by communication from another controller such as a meter or a brake controller. Note that the motor controller 2 may obtain the vehicle speed v (m / s) by multiplying the rotational angular velocity of the motor 4 (for example, either the front mechanical angular velocity or the rear mechanical angular velocity) by the tire dynamic radius r and dividing by the gear ratio of the final gear, and then multiplying this by 3600 / 1000 for unit conversion to calculate the vehicle speed V (km / h).

[0021] Accelerator opening A PO (%) is obtained from an accelerator opening sensor (not shown) or is obtained by communication from another controller (not shown) such as a vehicle controller.

[0022] Rotor phase α f ,α r (rad) is acquired from the rotation sensor 21. The motor controller 2 calculates the rotor phase α f ,α r Differentiate the rotor angular velocity ω of each motor 4f, 4r ef ,ω er (electrical angular velocity). The motor controller 2 also calculates the rotor angular velocity ω ef ,ω er is divided by the number of pole pairs p of the motors 4f and 4r to obtain the motor rotation angular velocity ω mf ,ω mr Furthermore, the motor controller 2 calculates the motor rotation angular velocity ω mf ,ω mr The motor rotation speed N is calculated by multiplying by the unit conversion coefficient (60 / 2π). mf ,N mr Calculate (rpm).

[0023] Motor current i f ,i r (A) is acquired from the current sensor 20. Also, the DC voltage value V dc (V) is detected by a voltage sensor (not shown) provided on the DC power supply line between the battery 1 and the inverter 3. The power supply voltage value obtained by a battery controller (not shown) is referred to as the DC voltage value V dc(V) may be obtained.

[0024] In step S202 (basic torque distribution process), the motor controller 2 executes basic torque distribution process. PO , vehicle speed V, and motor rotation speed N m (In this embodiment, for example, the front motor rotation speed N mf ), the accelerator opening-torque table shown in FIG. 3 is referenced to determine the basic combined torque command value T m * Set.

[0025] Furthermore, the motor controller 2 executes the front / rear torque distribution process shown in Fig. 4. Specifically, the motor controller 2 calculates the basic total torque command value T m * Front distribution gain K f (0≦K f ≦1), the front drivetrain S f The front basic torque command value T corresponds to the target value of the driving force to be distributed to the mf * The motor controller 2 calculates the basic combined torque command value T m * Rear distribution gain (1-K f ) by multiplying the rear drivetrain S r The rear basic torque command value T corresponds to the target value of the driving force to be distributed to mr * Ask for.

[0026] Next, in step S203 (vibration suppression process), the motor controller 2 executes the vibration suppression process. Specifically, the motor controller 2 executes the vibration suppression process by calculating the front basic torque command value T mf * Vibration in the drivetrain (front drive shaft DS f By performing correction to suppress the torsional vibration of the front wheel, the front final torque command value T mff *On the other hand, the motor controller 2 calculates the rear basic torque command value T mr * Vibration in the drivetrain (rear drive shaft DS r By performing vibration compensation calculations to suppress the torsional vibration of the rear wheel, the rear final torque command value T mrf * Details of the vibration suppression process will be described later.

[0027] In step S204, the motor controller 2 executes a current command value calculation process. Specifically, the motor controller 2 calculates the final torque command value T mff * ,T mrf * , each motor rotation angular velocity ω mf ,ω mr , and DC voltage value V dc Based on this, a predetermined table is referenced to determine the d-axis current target values ​​i of the front motor 4f and the rear motor 4r. df * ,i dr * and q-axis current target value i qf * ,i qr * Ask for.

[0028] In step S205, the motor controller 2 executes a current control calculation process. Specifically, the motor controller 2 calculates the d-axis current i df and q-axis current i qf are the d-axis current target values ​​i df * and q-axis current target value i qf * and the d-axis current i of the rear motor 4r dr and q-axis current i qr are the d-axis current target values ​​i dr * and q-axis current target value i qr * Control is performed to match the

[0029] More specifically, the motor controller 2 controls the d-axis current target value i df * ,i dr * and q-axis current target value i qf * ,i qr * The PWM signals are generated based on the above-mentioned values, and the switching elements of the front inverter 3f and the rear inverter 3r are opened and closed by the PWM signals thus obtained, thereby driving the front motor 4f and the rear motor 4r with desired torque.

[0030] The vibration suppression process S203 will be described in detail below.

[0031] 5 is a block diagram illustrating the vibration suppression process of this embodiment. In particular, the vibration suppression process S203 of this embodiment is f The configuration that suppresses vibrations in the drivetrain transmission system (top of Figure 5), and the rear drivetrain S r In the following, for the sake of simplicity, the front drivetrain S f and rear drive system S r Each component in the above will be comprehensively described with its corresponding reference numeral.

[0032] As shown in the figure, the vibration suppression process S203 includes F / F compensators S501 and S503 and F / B compensators S502 and S504.

[0033] 6 and 7 are block diagrams illustrating the F / F compensators S501 and S503. As shown in the figures, the F / F compensators S501 and S503 have vehicle models S601 and S701 and torque correction units S602 and S702.

[0034] The vehicle models S601 and S701 are used to calculate the first torque command value T mf1 * ,Tmr1 * is used as an input, and a dead zone model that simulates the driving force characteristics of the electric vehicle 100 is used to calculate the drive shaft DS f ,DS r The estimated torsional angular velocity ω df ^,ω dr ^ and estimated twist angle θ df ^,θ dr ^ is calculated. Here, the estimated torsional angular velocity ω df ^,ω dr The torsion angle estimated value θ is fed back to the torque correction units S602 and S702 and functions as an index for determining the amount of correction to suppress vibration in the torque correction units S602 and S702. df ^,θ dr ^ is fed back to the torque correction units S602 and S702, and each drive system S f ,S r This serves as an index for determining whether the signal is in the dead zone.

[0035] In addition, the drive system S f ,S r The dead zone is the period when the backlash of the gears is reduced, and the output torque of the motors 4f and 4r (hereinafter referred to as "motor torque T mf ,T mr ") but the drive shaft DS f ,DS r torque (hereinafter referred to as "drive shaft torque T df ,T dr In particular, the drivetrain S f ,S r The dead zone of the front drive system S f and rear drive system S r Each of these is specified separately. f ,S r are in the dead zone (a state in which there is no torque response of the electric vehicle 100), will be referred to as the "dead zone of the electric vehicle 100" as appropriate.

[0036] The vehicle models S601 and S701 are configured to calculate the first torque command value T mf1 * ,T mr1* Based on this, the motor rotation angular velocity ω mf ,ω mr (hereinafter referred to as the motor rotation angular velocity estimate ω mf ^,ω mr ^") is calculated and output to the F / B compensators S502 and S504.

[0037] The torque correction units S602 and S702 correct the basic torque command value T mf * ,T mr * , estimated twist angle θ df ^,θ dr ^, and the estimated torsional angular velocity ω df ^,ω dr ^ is input, and the first torque command value T mf1 * ,T mr1 * Specifically, the torque correction units S602 and S702 calculate the torsional angular velocity estimated value ω df ^,ω dr ^ is multiplied by a predetermined feedback gain to obtain the vibration compensation torque ΔT mf ,ΔT mr Calculate the basic torque command value T mf * ,T mr * Vibration compensation torque ΔT mf ,ΔT mr is subtracted to obtain the first torque command value T mf1 * ,T mr1 * Calculate the following.

[0038] In particular, the torque correction units S602 and S702 correct the torsion angle estimated value θ df ^,θ dr Depending on the reference result of ^, the normal gain k described later is used as the feedback gain. f1 ,k r1 or dead band gain k f2 ,k r2 Vibration compensation torque ΔT mf ,ΔT mrMore specifically, the torque correction units S602 and S702 calculate the torsion angle estimated value θ df ^,θ dr If ^ is not zero (drive system S f ,S r is estimated to be in a normal zone other than the dead zone), the estimated torsional angular velocity ω df ^,ω dr ^Normal gain k f1 ,k r1 Multiplied by the vibration compensation torque ΔT mf ,ΔT mr On the other hand, the torque correction units S602 and S702 calculate the torsion angle estimated value θ df ^,θ dr If ^ is zero (drive system S f ,S r is estimated to be in the dead zone), the estimated torsional angular velocity ω df ^,ω dr ^ the dead band gain k f2 ,k r2 Multiplied by the vibration compensation torque ΔT mf ,ΔT mr Ask for.

[0039] The vehicle models S601 and S701 shown in FIGS. 6 and 7 will be described in detail.

[0040] 8 is a diagram showing a model of the driving force transmission system of electric vehicle 100. The definitions of each parameter, including those already explained, are shown below.

[0041] J mf ,J mr : Motor inertia J wf ,J wr : Drive wheel inertia (per axle) K df ,K dr : Torsional rigidity of the drive shaft K tf ,K tr : Coefficient of friction between the tire and road surface N f ,N r :Overall gear ratio r f,r r : Tire load radius ω mf ,ω mr : Motor rotation angular velocity ω mf ^,ω mr ^: Estimated motor rotation angular velocity θ mf ,θ mr : Motor rotation angle ω wf ,ω wr : Drive wheel angular velocity θ wf ,θ wr : Drive wheel angle T mf ,T mr :Motor torque T df ,T dr : Drive shaft torque F f ,F r : Driving force (for 2 axes) θ df ,θ dr :Twist angle of the drive shaft ω df ,ω dr : Torsional angular velocity of the drive shaft V: Vehicle speed M: Vehicle weight

[0042] From FIG. 8, the equations of motion of the 4WD electric vehicle 100 are expressed by the following equations (1) to (11).

[0043]

number

[0044] By Laplace transforming the above equations (1) to (11), the front motor torque T mf from the front motor rotation angular velocity ω mf The transfer characteristics up to are expressed by the following equations (12) and (13): a3, a2, a1, a0, b3, b2, b1, and b0 in equation (13) are each expressed by equation (14).

[0045]

number

[0046]

number

[0047] In addition, the front motor torque T mf Front drive shaft torque T df The transfer characteristics up to are expressed by the following equation (15): where c1 and c2 in equation (15) are expressed by equation (16).

[0048]

number

[0049] From equations (3), (6), (8), and (10), the front motor rotational angular velocity ω mf from the front drive wheel angular velocity ω wf The transfer characteristic up to is expressed by the following equation (17).

[0050]

number

[0051] From equations (12), (13), and (17), the front motor torque Tmf is used to calculate the front drive wheel angular velocity ω wf The transfer characteristic up to is expressed by the following equation (18).

[0052]

number

[0053] From equations (15) and (18), the front drive shaft torque T df from the front drive wheel angular velocity ω wf The transfer characteristic up to is expressed by the following equation (19).

[0054]

number

[0055] Here, by transforming equation (1), the following equation (20) is obtained.

[0056]

number

[0057] Therefore, by equations (19) and (20), the front drive shaft DS f Torsional angular velocity ω df (Front torsional angular velocity ω df ) can be expressed by the following equation (21).

[0058]

number

[0059] However, H in formula (21) wf (s) is determined by the following equation (22).

[0060]

number

[0061] Furthermore, v1, v0, w1, and w0 in equation (22) are determined by the following equation (23).

[0062]

number

[0063] Moreover, equation (15) can be transformed into the following equation (24).

[0064]

number

[0065] Here, "ζ" in equation (24) p" is the front drive shaft DS f The damping coefficient of the torque transmission system, ω p " is the front drive shaft DS f is the natural vibration frequency of the torque transmission system.

[0066] Furthermore, since the pole α and zero c0 / c1 of the transfer function expressed by equation (24) can be considered to be almost the same, the following equation (25) can be obtained by performing pole-zero cancellation. However, g in equation (25) tf is determined by equation (26).

[0067]

number

[0068] Here, the front drivetrain S shown in Figure 6 f According to the control logic of the torque correction unit S602, the first torque command value T mf1 * is the front final torque command value T mff * The front final torque command value T mff * , the front basic torque command value T mf * and normal gain k f1 is expressed by the following equation (27).

[0069]

number

[0070] Then, the front final torque command value T mff * can be substituted as shown in the following equation (28) using equations (6) and (10).

[0071]

number

[0072] And T mf =Tmff * By substituting equation (28) into equation (24), equation (29) can be obtained.

[0073]

number

[0074] Here, the front basic torque command value T mf * Front drive shaft torque T df The reference response up to is expressed by the following equation (30).

[0075]

number

[0076] The normal gain k for which equation (29) and equation (30) match f1 The condition is expressed by the following equation (31): f1 is the front basic torque command value T mf * Front motor torque T according to mf is the front drive shaft torque T df It is determined from the viewpoint of realizing the standard response in the normal section transmitted as

[0077]

number

[0078] In addition, the dead band gain k f2 is determined by the following equation (32). γf " is a front drive system S f damping coefficient of the reference response in the dead zone section of the γf ").

[0079]

number

[0080] That is, the front dead band gain k f2 is the front basic torque command value T mf * Front motor torque T according to mf is the front drive shaft torque T df In particular, the front deadband damping coefficient ζ is determined from the viewpoint of realizing a standard response in the deadband section where the damping coefficient is not sufficiently transmitted as γf is made smaller than the damping coefficient in the reference response of the normal section (hereinafter also referred to as the reference damping coefficient, which is "1" in this embodiment), the dead band gain k f2 is the normal gain k f1 As a result, the front drivetrain S f When is in the dead zone, the front torsional angular velocity ω df and dead band gain k f2 Vibration compensation torque ΔT is determined by the product of mf (i.e., front basic torque command value T mf * Since the reduction correction amount for the torque is smaller, the torque response is faster than in the normal section.

[0081] Furthermore, in this embodiment, equations (1) to (23) are applied to calculate the front drivetrain S f A dead zone model is constructed that simulates the gear backlash characteristics in the front drive shaft torque T df is expressed by the following equation (33): deadf " is a front motor 4f to front drive shaft DS f This represents the overall gear backlash amount up to

[0082]

number

[0083] Also, rear drive system S r Normal gain k r1 and dead band gain kr2 Regarding the front drive system S f Normal gain k f1 and dead band gain k f2 Using the same calculation logic as above, it can be calculated according to the following equations (34) and (35). γr ” is the damping coefficient of the reference response in the rear dead zone (hereinafter referred to as the “rear dead zone damping coefficient ζ γr ").

[0084]

number

[0085] Furthermore, rear drive system S r The rear drive shaft torque T is calculated by taking into account the dead zone model that simulates the gear backlash characteristics in dr Also, front drive system S f Using the same calculation logic, it can be expressed as the following equation (36).

[0086]

number

[0087] The front drivetrain S shown in the above equations (33) and (36) f and rear drive system S r The dead zone model is the dead zone block B shown in Fig. 6. df and the dead zone block B shown in Fig. 7 dr are realized by

[0088] Next, the dead band gain k f2 ,k r2 The specific method for determining this will be explained below.

[0089] 9A and 9B are diagrams illustrating the torque response characteristics of the electric vehicle 100 according to the combination of the mechanical characteristics (gear ratios) of each drive train. m *, and Fig. 9(B) shows the front final torque command value T mff * 9(C) shows the rear final torque command value T mrf * 9(D) shows the longitudinal acceleration G of the electric vehicle 100. In FIG. 9, a solid line indicates the case where the front derailleur 5f is in "first gear" and the rear derailleur 5r is in "first gear," and a dashed line indicates the case where the front derailleur 5f is in "second gear" and the rear derailleur 5r is in "second gear." Also, a dashed line indicates the case where the front derailleur 5f is in "second gear" and the rear derailleur 5r is in "first gear," and a chain-dotted line indicates the case where the front derailleur 5f is in "first gear" and the rear derailleur 5r is in "second gear."

[0090] 9, in a scene where the vehicle is gradually transitioning from a deceleration state (longitudinal acceleration G<0) to an acceleration state (longitudinal acceleration G>0), the torque response of the electric vehicle 100 according to this embodiment enters a dead zone at approximately time t1. However, the timing at which the dead zone is exited varies depending on the combination of the gear ratios of the transmissions 5f, 5r.

[0091] In the torque response of the electric vehicle 100 according to this embodiment, the gear ratio of the front transmission 5f is dominant, and when the gear ratio of the front transmission 5f is "first gear," the dead zone is exited at approximately time t2, and when the gear ratio of the front transmission 5f is "second gear," the dead zone is exited at time t3, which is later than time t2. For this reason, the stagnation (stay) time of the torque response of the electric vehicle 100 in the dead zone switches between approximately "t2-t1" and "t3-t1."

[0092] Of course, in other electric vehicles, the duration of time spent in the dead band section may vary more finely depending on the combination of gear ratios, for example, in a manner different from the example shown in Figure 9. Also, in the electric vehicle 100 according to this embodiment, in addition to the above-described change in the gear ratio of the front transmission 5f, when comparing the cases where the gear ratio of the rear transmission 5r is "first gear" and "second gear," the torque response is faster when the gear ratio of the rear transmission 5r is "first gear" than when it is "second gear."

[0093] That is, the stagnation time in the dead zone of the torque response varies greatly depending on the combination of the gear ratios of the transmissions 5f, 5r, which are mechanical characteristics of the transmissions 5f, 5r. This variation changes every time the combination of the gear ratios of the transmissions 5f, 5r changes, which makes a driver who performs a certain accelerator operation feel uncomfortable when operating the electric vehicle 100.

[0094] Therefore, in this embodiment, a dead band gain k that changes depending on the combination of the gear ratios (mechanical characteristics) of the transmissions 5f and 5r is used so that a constant torque response can be obtained regardless of the combination of the gear ratios, which are the mechanical characteristics of the transmissions 5f and 5r. f2 ,k r2 In particular, in this embodiment, the combination that provides the fastest torque response among the combinations of the speed ratios of the transmissions 5f and 5r is used as the reference, and the dead band gain k is set so as to provide that fastest torque response. f2 ,k r2 In the example of Fig. 9, as shown by the solid line, the combination in which the gear ratio of the front transmission 5f is "1st gear" and the gear ratio of the rear transmission 5r is "1st gear" is the combination that provides the fastest torque response among the combinations of the gear ratios of the transmissions 5f, 5r.

[0095] Hereafter, the dead band gain k f2 ,k r2 The details of the setting of the dead band gain k will be explained. f2 ,k r2 is expressed as the deadband damping coefficient ζ γf ,ζ γr This is set via the config.

[0096] Figure 10 shows each drive system S f ,S r Dead band damping coefficient ζ according to the combination of mechanical characteristics (gear ratio) γf ,ζ γr 10 is a speed ratio-dead band damping coefficient table showing the settings of

[0097] As mentioned above, the combination of gear ratios that provides the quickest torque response in the electric vehicle 100 is when the gear ratio of the front transmission 5f is "first gear" and the gear ratio of the rear transmission 5r is "first gear." Therefore, as shown in FIG. 10, the front dead-band damping coefficient ζ γf and rear dead zone damping coefficient ζ γr is set to the reference damping coefficient when the gear ratio of the front derailleur 5f is "1st gear" and the gear ratio of the rear derailleur 5r is "1st gear." In this embodiment, the reference damping coefficient is "1," and the front dead-band damping coefficient ζ γf and rear dead zone damping coefficient ζ γr However, the front dead band damping coefficient ζ γf and rear deadband damping coefficient ζ γr A different reference damping coefficient can be defined for each of the

[0098] Front deadband damping coefficient ζ γf is a value smaller than the reference damping coefficient (here, 1>ζ) when the front derailleur 5f is in "second gear" and the stagnation time in the dead band section is long due to the setting of the gear ratio of the front derailleur 5f. γf ) is set to the front deadband damping coefficient ζ γf When is set to a value smaller than the reference damping coefficient, the front deadband gain k f2 As a result, the front basic torque command value T mf * The correction amount for the front drive system S is reduced. f Therefore, the stagnation time in the dead zone of the torque response of electric vehicle 100 is reduced, and the stagnation time approaches the stagnation time obtained by the combination of gear ratios that provides the fastest torque response.

[0099] Similarly, the rear deadband damping coefficient ζ γr is a value smaller than the reference damping coefficient (here, 1>ζ) when the rear derailleur 5r is in "second gear" and the stagnation time in the dead band section is long due to the setting of the gear ratio of the rear derailleur 5r. γr ) is set to the rear deadband damping coefficient ζ γrWhen is set to a value smaller than the reference damping coefficient, the rear deadband gain k r2 As a result, the rear basic torque command value T mr * The correction amount for the rear drive system S is reduced. r Therefore, the stagnation time in the dead zone of the torque response of electric vehicle 100 is reduced, and the stagnation time approaches the stagnation time obtained by the combination of gear ratios that provides the fastest torque response.

[0100] Therefore, specifically, as shown in FIG. 10, when the speed ratio of the front transmission 5f is "first speed" and the speed ratio of the rear transmission 5r is "first speed", the front dead-band damping coefficient ζ γf and rear dead zone damping coefficient ζ γr are both set to the reference damping coefficient "1." When the speed ratio of the front transmission 5f is "second gear" and the speed ratio of the rear transmission 5r is "second gear," the front dead-band damping coefficient ζ γf and rear dead zone damping coefficient ζ γr are set to values ​​smaller than the reference damping coefficient. When the speed ratio of the front derailleur 5f is "first speed" and the speed ratio of the rear derailleur 5r is "second speed", the front dead-band damping coefficient ζ γf is set to the reference damping coefficient, and the rear deadband damping coefficient ζ γr is set to a value smaller than the reference damping coefficient. Conversely, when the speed ratio of the front derailleur 5f is "second speed" and the speed ratio of the rear derailleur 5r is "first speed," the front dead-band damping coefficient ζ γf is set to a value smaller than the reference damping coefficient, and the rear deadband damping coefficient ζ γr is set to the reference damping coefficient.

[0101] 11 is an explanatory diagram showing the relationship between the front torque correction section S602 and the setting of the dead-band damping coefficient. As shown in FIG. 11, in this embodiment, the front torque correction section S602 sets the front dead-band gain k f2 is the front dead-band damping coefficient ζ set by the gear ratio-dead-band damping coefficient table based on equation (32). γfIt is set variably depending on the

[0102] Specifically, when the speed ratio of the front transmission 5f is "first gear", the front dead band damping coefficient ζ γf is set to the reference damping coefficient, so the front dead band gain k f2 is the front normal gain k f1 On the other hand, when the speed ratio of the front derailleur 5f is "second speed" and the stagnation time in the dead-band section is long due to the mechanical characteristics of the front derailleur 5f, the front dead-band damping coefficient ζ γf is set to a value smaller than the reference damping coefficient. This allows the front dead band gain k f2 is the normal front gain k f1 As a result, the torque response of the electric vehicle 100 is improved.

[0103] Here, the relationship between the front torque correction section S602 and the setting of the dead-zone damping coefficient has been described, but the relationship between the rear torque correction section S702 and the setting of the dead-zone damping coefficient is also the same as above. That is, the rear torque correction section S702 calculates the rear dead-zone gain k r2 is the rear dead band damping coefficient ζ set by the gear ratio-dead band damping coefficient table based on equation (35). γr It is set variably depending on the

[0104] As described above, in this embodiment, the dead-band damping coefficient ζ is adjusted in accordance with the combination of the gear ratios (mechanical characteristics) of the transmissions 5f and 5r. γf ,ζ γr By setting each drive system S f ,S r Deadband gain k f2 ,k r2 is the normal gain k f1 ,k r1 Therefore, in a scene where the stagnation time in the dead zone section is long depending on the combination of the gear ratios of the transmissions 5f and 5r, the drive system S f ,S r Vibration compensation torque ΔT at m(=k f ×ω d ) becomes smaller. As a result, the torque response performance of the electric vehicle 100 in the dead-band section, i.e., the ability of the actual driving force to follow the total required driving force, is improved, and variations in the torque response characteristics are reduced or suppressed regardless of the combination of the gear ratios (mechanical characteristics) of the transmissions 5f, 5r. In addition, the dead-band gain k f2 ,k r2 Therefore, the torque response characteristic of the electric vehicle 100 in the dead zone section is made to approach the highest torque response characteristic, and the timing of exiting the dead zone section becomes substantially constant.

[0105] Next, the F / B compensators S502 and S504 will be described in detail.

[0106] 12 and 13 are block diagrams showing the configurations of the F / B compensators S502 and S504. In particular, Fig. 12 shows the configuration of the front F / B compensator S502, and Fig. 13 shows the configuration of the rear F / B compensator S504.

[0107] As shown in the figure, the F / B compensators S502 and S504 have gain sections S5021 and S5041, filter sections S5022 and S5042, and filter sections S5023 and S5043.

[0108] The gain κ used in the gain sections S5021 and S5041 is arranged to adjust the stability margin (gain margin, phase margin) of the feedback control system, and is set to a value of 1 or less.

[0109] The filter sections S5022 and S5042 are connected to the motor torque T mf ,T mr from the motor rotation angular velocity ω mf ,ω mr Transfer characteristics G p (s). The transfer characteristic Gp(s) is determined by, for example, the above equation (13).

[0110] The filter sections S5023 and S5043 have a transfer characteristic G p H(s) / G consists of the inverse system of (s) and the bandpass filter H(s) p The bandpass filter H(s) has attenuation characteristics that are approximately the same on the low-pass side and the high-pass side, and has a torsional resonance frequency f p is set to be near the center of the passband on a logarithmic scale.

[0111] For example, when the band-pass filter H(s) is configured with a first-order high-pass filter and a first-order low-pass filter, the band-pass filter H(s) is configured as shown in the following equation (37): L =1 / (2πf HC ), f HC =k f p , τ H =1 / (2πf LC ), f LC =f p / k. Note that the frequency f p Drivetrain S f ,S r is the torsional resonance frequency of the filter, and k is an arbitrary value that constitutes a bandpass filter.

[0112]

number

[0113] As a result, the F / B compensators S502 and S504 use the first torque command value T calculated by the vehicle models S601 and S701 of the F / F compensators S501 and S503. mf1 * ,T mr1 * The motor rotation angular velocity estimate ω based on mf ^,ω mr ^ and the second torque command value T before being multiplied by the gain κ mf2 * ,T mr2 * The transfer characteristic G p (s) and the estimated motor rotation angular velocity ω mf1^,ω mr1 ^ and are added to obtain the final motor rotation angular velocity estimate ω mff ^,ω mrf Furthermore, the F / B compensators S502 and S504 calculate the final motor rotation angular velocity estimate value ω mff ^,ω mrf ^ and the motor rotation angular velocity detection value ω obtained by the rotation sensor 21 mf-d ,ω mr-d Calculate the deviation from the filter H(s) / G p (s) and multiplying it by the gain κ to obtain the second torque command value T mf2 * ,T mr2 * Ask for.

[0114] As shown in FIG. 5, in the vibration suppression process S203, the final torque command value T mff * ,T mrf * is output from the F / F compensators S501 and S503 as the first torque command value T mf1 * ,T mr1 * and the second torque command value T output from the F / B compensators S502 and S504. mf2 * ,T mr2 * The operation is performed as the sum of

[0115] In this way, the final torque command value T mff * ,T mrf * is calculated, the first torque command value T after vibration compensation (feedforward compensation) by the F / F compensators S501 and S503 is calculated. mf1 * ,T mr1 * Further, the torque command value T mff * ,T mrf * As a result, each drive system S f ,Sr Therefore, vibrations in the driving force transmission system can be more reliably suppressed.

[0116] <Control Results According to the First Embodiment> 14 is a comparison diagram between the control results of the first embodiment (first example) and the control results of a comparative example (first comparative example). FIGS. 14(A) to 14(D) show the control results of the comparative example (first comparative example), and FIGS. 14(E) to 14(H) show the control results of the first embodiment (first example). In addition, from top to bottom in the figure, the basic total torque command value T m * , front final torque command value T mff * , rear final torque command value T mrf * , and longitudinal acceleration G, respectively. The solid line indicates the control result when the gear ratio of the front transmission 5f is "first gear" and the gear ratio of the rear transmission 5r is "first gear," and the dashed line indicates the control result when the gear ratio of the front transmission 5f is "second gear" and the gear ratio of the rear transmission 5r is "second gear." The dashed dotted line indicates the control result when the gear ratio of the front transmission 5f is "second gear" and the gear ratio of the rear transmission 5r is "first gear," and the dashed two dotted line indicates the control result when the gear ratio of the front transmission 5f is "first gear" and the gear ratio of the rear transmission 5r is "second gear."

[0117] In addition, in FIG. 14, when the vehicle is decelerating due to the regenerative torque, the basic combined torque command value T m * The figure shows the control results in a scene where the vehicle accelerates by increasing the driving force at a gentle rate. f ,S r and distributes the driving force to the transmission 5f, 5r, and generates a vibration compensation torque ΔT according to the reference damping coefficient, regardless of the combination of the speed ratios of the transmissions 5f, 5r. mf ,ΔT mr The feedback gain (dead band gain k f2 ,k r2 On the other hand, in the first embodiment, each drive system S f ,Sr The driving force distribution is the same as that of the first comparative example, and the dead zone gain k f2 ,k r2 Set.

[0118] As shown in Figures 14(A) to (D), in the first comparative example, after the longitudinal acceleration G becomes 0 at time t1 and the vehicle enters the dead zone, the time spent in the dead zone (time t2 to t3) varies greatly depending on the combination of gear ratios of the transmissions 5f, 5r.

[0119] In contrast, as shown in Figures 14(E) to (H), referring to the control results of the first embodiment, regardless of the combination of the gear ratios of the transmissions 5f, 5r, the timing of the torque response approaches the fastest torque response when both the gear ratios of the transmissions 5f, 5r are in "first gear." As a result, the variation in the stagnation time (times t2 to t3) in the dead-band section is eliminated, and the timing of exiting the dead-band section is substantially constant regardless of the combination of the gear ratios of the transmissions 5f, 5r. This is because the dead-band section is estimated for each of the front and rear drivetrains, and the dead-band damping coefficient ζ γf ,ζ γr Through the setting of df ,ω dr This is because the feedback gain multiplied by is reduced.

[0120] In the first embodiment, each drive system S f ,S r The dead band damping coefficient ζ γf ,ζ γr and through this, the dead band gain k f2 ,k r2 However, the torque response of the electric vehicle 100 is determined by the respective drivetrain S f ,S r The mechanical characteristics of each drive system S f ,S r Specifically, in addition to the combination of gear ratios, each drive system S f ,S r inertia, differential gear ratio, drive shaft DSf ,DS r The combination of the inertia and torsional rigidity of each of the drive trains S may also cause a difference in the stagnation time in the dead zone in the torque response of the electric vehicle 100. f ,S r The dead band damping coefficient ζ is set according to one or more combinations of vehicle conditions based on the mechanical characteristics of the γf ,ζ γr may be set.

[0121] In the first embodiment, the dead band damping coefficient ζ is set to the combination of gear ratios that provides the fastest torque response, so that the fastest torque response is maintained regardless of the combination of gear ratios. γf ,ζ γr However, the dead band damping coefficient ζ γf ,ζ γr is not limited to this setting method, and may be set so as to at least reduce variations in torque response. For example, the combination of speed ratios that results in the slowest torque response is used as a reference, and the dead-band damping coefficient ζ is set to match this. γf ,ζ γr However, as in the first embodiment, the dead band damping coefficient ζ may be set to match a combination of speed ratios that provides the fastest possible torque response. γf ,ζ γr It is of course desirable to set

[0122] [Second embodiment] In the first embodiment, each drive system S f ,S r The dead zone damping coefficient ζ is determined according to the mechanical characteristics of γf ,ζ γr By setting the dead band gain k f2 ,k r2 and the basic torque command value T mf * ,T mr * However, the dead band damping coefficient ζ γf ,ζ γrThe method of setting is not limited to this. In the second embodiment, an example will be described in which the dead-band damping coefficient is set in accordance with the drive force distribution of each drive system. Specifically, in the second embodiment, the dead-band damping coefficient is set in accordance with the drive system S that has a faster torque response of the electric vehicle 100 when all the drive force is distributed. f ,S r The dead band gain k changes depending on the driving force distribution, while f2 ,k r2 The following is established.

[0123] 15 is a diagram illustrating the torque response characteristics of the electric vehicle 100 according to the driving force distribution. In particular, FIG. 15 shows the torque response characteristics of the electric vehicle 100 according to the front distribution gain K f When set to 1 (front drive system S f The solid lines show the time-dependent changes in each parameter (when all driving force is distributed to the front wheels), and the front distribution gain K f The broken lines show the time-dependent changes of each parameter when the basic combined torque command value T m * , and Fig. 15(B) shows the front final torque command value T mff * , and Fig. 15(C) shows the rear final torque command value T mrf * 15(A) and 15(B) show the longitudinal acceleration G of the electric vehicle 100, respectively.

[0124] First, the front drivetrain S f In the scene where the entire driving force is distributed to the front drive system S to transition from a deceleration state (front-rear acceleration G<0) to an acceleration state (front-rear acceleration G>0), f At time t1, the vehicle enters the dead zone, and at time t2, the vehicle exits the dead zone and generates a torque response. r When the entire driving force is distributed to the front drivetrain S, the dead zone is entered at the same time t1, but the torque response occurs at time t3, which is later than time t2. f When all driving force is distributed to (K f =1) in the front drivetrain S fThe stagnation time (t2-t1) in the dead zone of the rear drivetrain S r When all driving force is distributed to (K f =0) in the rear drivetrain S r This is shorter than the stagnation time (t3-t1) in the dead zone of the front drivetrain S f and rear drive system S r Therefore, in this case, the torque response in the dead zone is faster when the full driving force is distributed to the front drivetrain S. f Based on the dead band gain k f2 ,k r2 The following is established.

[0125] Hereinafter, the dead zone damping coefficient ζ γf ,ζ γr The dead band gain k is set by setting f2 ,k r2 The details of the mode of determining the above will be described below.

[0126] Fig. 16 shows a driving force distribution ratio-dead zone damping coefficient table, and Fig. 17 shows the relationship between the front torque correction section S602 and the driving force distribution ratio-dead zone damping coefficient table.

[0127] As shown in Fig. 16, the front dead zone damping coefficient ζ γf Front drivetrain S f When all driving force is distributed to (K f = 1), which is the same as the reference damping coefficient, 1, which is the damping coefficient of the standard response in the normal section. Therefore, as shown in FIG. 17, the front dead band gain k f2 is the normal front gain k f1 (Equation (31)). In other words, in this case, the front drivetrain S f Therefore, the electric vehicle 100 is in a state where it is driven solely by the front drivetrain S f This ensures torque response performance equivalent to that achieved when only the engine is installed.

[0128] In addition, the front drive system Sf and rear drive system S r The driving force is distributed to both the front drivetrain and the rear drivetrain. f The area where the driving force distribution is relatively large (0.5 <K f <1), the front dead band damping coefficient ζ γf is the front distribution gain K f Even if the driving force is reduced, f = 1) to maintain the torque response performance of the front distribution gain K f On the other hand, the rear drivetrain S r Rear deadband damping coefficient ζ γr is maintained at the reference damping coefficient. f This can reduce variations in torque response characteristics caused by changes in the drive force distribution, thereby further improving the torque response performance of electric vehicle 100.

[0129] Furthermore, the front drive system S f Rear drivetrain S r The area where the driving force is large (0 <K f <0.5), the front deadband damping coefficient ζ γf is maintained at the reference damping coefficient. Meanwhile, the rear deadband damping coefficient ζ γr Rear drivetrain S r The fastest torque response is achieved when all driving force is distributed (K f =0) as the base point, 0.5 <K f Gain k for front deadband in the region <1 f2 The same decrease profile as the front distribution gain K f Increase in rear drive system S r In other words, in this case, the rear drivetrain S r The torque response performance of the is independent of the driving force distribution, and is f = 0). This keeps the rear drivetrain S r This can reduce variations in torque response characteristics caused by changes in the drive force distribution, thereby further improving the torque response performance of electric vehicle 100.

[0130] The above explanation applies to the rear drivetrain S when all driving force is distributed. r The torque response in the dead zone of the front drivetrain S f When the rear drivetrain S is higher than r Based on the dead band gain k f2 ,k r2 This is also applicable to cases where

[0131] As described above, in the second embodiment, the drivetrain S f ,S r Deadband gain k f2 ,k r2 is the normal gain k r1 ,k r1 Therefore, the drivetrain S f ,S r Vibration compensation torque ΔT m (=k f ×ω d ) becomes smaller, and the torque response performance (the ability of the actual driving force to follow the total required driving force) of the electric vehicle 100 in the dead zone section improves. In addition, the front drivetrain S, which shows a faster torque response when the entire driving force is distributed, f Based on the dead band gain k f2 ,k r2 is determined, each drive system S f ,S r Regardless of the difference in the mechanical characteristics of the front drivetrain S, the torque response characteristics of the electric vehicle 100 in the dead zone are f The characteristics can be made closer to those when only the drive source is used, and the timing at which the dead zone is passed can be made substantially constant.

[0132] <Control Results According to the Second Embodiment> FIG. 18 is a comparison diagram between the control results of the second embodiment (Second Example) and the control results of a comparative example (Second Comparative Example). FIGS. 18(A) to 18(D) show the control results of the comparative example (Second Comparative Example), and FIGS. 18(E) to 18(H) show the control results of the second embodiment (Second Example). In addition, from top to bottom in the diagram, the basic combined torque command value T m * , front final torque command value T mff * , rear final torque command value T mrf * , and longitudinal acceleration G, respectively. In each figure, the solid line, dashed line, and dashed line represent the front distribution gain K f The control results are shown for values ​​of 0.5, 0.7, and 1.0.

[0133] In addition, in FIG. 16, when the vehicle is decelerating due to the regenerative torque, the basic combined torque command value T m * The figure shows the control results in a scene where the vehicle accelerates by increasing the torque at a gentle slope. In the second comparative example, the mechanical characteristics of the front and rear drivetrains are differentiated, and the vibration compensation torque ΔT is adjusted according to the reference damping coefficient, regardless of the drive force distribution. mf ,ΔT mr Each feedback gain (dead band gain k f2 ,k r2 On the other hand, in the second example, the difference in the mechanical characteristics of the front and rear drive trains is set to be the same as in the second comparative example, and the dead zone gains (particularly the front dead zone gain k f2 ) to set the

[0134] As shown in Figures 18(A) to 18(D), in the second comparative example, after the longitudinal acceleration G becomes 0 at time t1 and the vehicle enters the dead zone, the same drive system (front drive system S in the figure) f Even with the same drive force distribution, the time spent in the dead zone (time t2 to t3) varies greatly depending on the drive force distribution. Furthermore, even with the same drive force distribution, the time spent in the dead zone differs between the front and rear drivetrains due to differences in mechanical characteristics.

[0135] In contrast, as shown in Figures 18(E) to (H), referring to the control results (solid lines) of the second embodiment, the front drivetrain S f Regardless of the driving force distribution, the torque response timing is the timing when the full driving force is distributed (K f = 1.0) (time t2). As a result, the variation in the stagnation time in the dead zone (times t2 to t3) is eliminated, and the timing of exiting the dead zone is essentially constant, regardless of the drive force distribution. This is because the dead zone is estimated for each front and rear drive system, and the dead zone damping coefficient ζ γf ,ζ γr Through the setting of df ,ω dr This is due to the fact that the feedback gain multiplied by the is reduced. In particular, the front drivetrain S, which has a short dead zone due to its mechanical characteristics, f Based on the dead band damping coefficient ζ γf ,ζ γr and through this, the dead band gain k f2 ,k r2 By setting the above, the torque response timing (time t2) when the entire driving force is distributed is also advanced, so the time that the electric vehicle 100 remains in the dead zone is further shortened.

[0136] The second embodiment can be implemented in combination with the first embodiment. For example, the dead-band damping coefficient ζ can be adjusted depending on the combination of the gear ratios of the transmissions 5f and 5r. γf ,ζ γr (See Figure 10) and set the set value as the reference (upper limit), and then set the front distribution gain K f Depending on the deadband damping coefficient ζ γf ,ζ γr In this case, each drive system S f ,S r Regardless of the mechanical characteristics of each drive system S f ,S r This makes it possible to reduce or suppress variations in the time during which the torque response of electric vehicle 100 stagnates in the dead zone, regardless of the driving force distribution.

[0137] [First Modification] The vibration suppression process S203 in the first and second embodiments can be modified to the configuration of a first modified example described below.

[0138] 19 is a block diagram illustrating the vibration suppression process S203 of the first modified example. In particular, in the first modified example, an F / F compensator S1201 is used instead of the F / F compensators S501 and S503 of the first embodiment.

[0139] 20 is a block diagram illustrating the configuration of the F / F compensator S1201. As shown in the figure, the F / F compensator S1201 is f and rear drive system S r The vehicle model S1301 is composed of each dead zone model that simulates the driving force characteristics of the front and rear wheels, and the front basic torque command value T mf * and the estimated front torsional angular velocity ω df ^ to front first torque command value T mf1 * and a front torque correction unit S1302 that calculates the rear basic torque command value T mr * and the estimated rear torsional angular velocity ω dr ^ to rear first torque command value T mr1 * and a rear torque correction unit S1303 that calculates the rear torque.

[0140] In addition, the front drive system S f Normal gain k f1 and dead band gain k f2 can be determined by the formulas (31) and (32) in the same manner as in the first embodiment. r Normal gain k r1 and dead band gain k r2 can also be determined by equations (34) and (35), respectively, in the same way as in the first embodiment.

[0141] Furthermore, each drive system S f , S r In the motor 4f, 4r to the drive shaft DSf ,DS r The backlash characteristics of the gears up to can also be determined from equations (33) and (36), respectively, in the same way as in the first embodiment.

[0142] Returning to FIG. 19, the vibration suppression process S203 includes F / B compensators S1202 and S1203 instead of the F / B compensators S502 and S504 of the first embodiment.

[0143] Here, the F / B compensator S1202 calculates the front motor rotational angular velocity estimate ω mf ^ to the front motor rotation angular velocity detection value ω mf-d Subtract and apply bandpass filter H f (s) and front vehicle model G pf By multiplying the inverse model of (s), the second front torque command value T mf2 * Furthermore, the F / B compensator S1202 calculates the front first torque command value T mf1 * The second front torque command value T mf2 * By adding mff * Ask for.

[0144] On the other hand, the F / B compensator S1203 calculates the estimated rear motor rotational angular velocity ω mr ^ to the rear motor rotation angular velocity detection value ω mr-d Subtract and apply bandpass filter H r (s) and vehicle model G pr By multiplying the inverse model of (s), the second rear torque command value T mr2 * Furthermore, the F / B compensator S1203 and the rear first torque command value T mr1 * The second rear torque command value T mr2 * By adding mrf * Ask for.

[0145] The electric vehicle control method according to the first modification also achieves the same effects as those of the first and second embodiments.

[0146] The control logic relating to the electric vehicle control method described in each of the above embodiments and the first modified example can also be applied to each vehicle having the system configuration shown in each of Figures 21 to 23 by making appropriate necessary modifications.

[0147] [Second Modification] Specifically, the electric vehicle 200 shown in FIG. 21 does not have a drive system at the front, but has two first rear drive systems S rR and second rear drive system S rL It is equipped with:

[0148] 1st rear drive system S rR is the first rear drive shaft DS rR The first rear motor 4rR drives the first rear drive wheel 9rR via the first rear motor 4rR, and various sensors and actuators for controlling the first rear motor 4rR. rL DS is the second rear drive shaft rL The second rear motor 4rL drives the second rear drive wheel 9rL via a second rear motor 4rL, and various sensors and actuators for controlling the second rear motor 4rL.

[0149] The electric vehicle 200 of this modification is, for example, a vehicle having a front drivetrain S f and rear drive system S r Each parameter of the first rear drive system S rR and second rear drive system S rL By making modifications such as replacing the parameters with those related to the above and setting a suitable vehicle model, the electric vehicle control method according to the present invention can be executed.

[0150] [Third Modification] The electric vehicle 300 shown in FIG. 22 has a front drivetrain S f , 1st rear drive system S rR , and second rear drivetrain S rLThat is, in this electric vehicle 300, the motor 4 is made up of three motors: a front motor 4f that drives the front drive wheels 9fR and 9fL, a first rear motor 4rR that drives the first rear drive wheel 9rR, and a second rear motor 4rL that drives the second rear drive wheel 9rL.

[0151] In the electric vehicle 300 of this modification, for example, the rear drivetrain S r The parameters set in the first rear drivetrain S rR and second rear drive system S rL By allocating the power to the electric vehicle, the electric vehicle control method according to the present invention can be carried out.

[0152] [Fourth Modification] The electric vehicle 400 shown in FIG. 23 has a front drivetrain S f However, the first front drive shaft DS fR A first front drive system S is provided with various sensors and actuators for controlling a first front motor 4fR that drives a first front drive wheel 9fR via a fR and the second front drive shaft DS fL A second front drive system S is provided with various sensors and actuators for controlling a second front motor 4fL that drives a second front drive wheel 9fL via a fL Also, the rear drive system S r Also the first rear drive system S rR and second rear drive system S rL Therefore, the electric vehicle 400 has four motors 4: a first front motor 4fR, a second front motor 4fL, a first rear motor 4rR, and a second rear motor 4rL.

[0153] In the electric vehicle 400 of this modification, for example, the front drivetrain S f Each parameter of the first front drive system S fR and second front drive system S fL While the rear drivetrain Sr Each parameter of the first rear drive system S rR and second rear drive system S rL By appropriately allocating the power to the electric vehicle, the electric vehicle control method according to the present invention can be carried out.

[0154] The configurations of the above-described embodiments and modifications, and the resulting effects will be described below.

[0155] The electric vehicle control method according to each of the above embodiments and each of the modifications is a method for controlling a plurality of drive systems (S f ,S r In the electric vehicle 100 equipped with each drive system (S f ,S r ) driving force (T mf ,T mr In this electric vehicle control method, the total required driving force (T m * ) and each drive system (S f ,S r ) to the driving force distribution (K f ,1-K f ) based on the basic torque command value (T mf * ,T mr * ) and the basic torque command value (T mf * ,T mr * ) to the driving force transmission system (S f ,S r ) to suppress the vibration and obtain the corrected torque command value (T mf1 * ,T mr1 * , or T mff * ,T mrf * ) and the correction torque command value (T mf1 * ,T mr1 * , or Tmff * ,T mrf * ) based on the driving force (T mf ,T mr and a driving force control process (S204, S205) for controlling the driving force.

[0156] Then, in the vibration suppression process S203, each drive system (S f ,S r The damping coefficient of the reference response related to the driving force transmission in the dead zone is the reference damping coefficient ("1"), which is the damping coefficient when the vehicle is in a zone other than the dead zone, and the dead zone damping coefficient ζ, which is the damping coefficient when the vehicle is in the dead zone. γf ,ζ γr and are set as different parameters. f ,S r ) is in the dead zone. Then, for the drive system outside the dead zone, the basic torque command value (T mf * ,T mr * ) is set, and in the drive system in the dead zone, the dead zone damping coefficient ζ γf ,ζ γr Based on this, the basic torque command value (T mf * ,T mr * ) correction amount (ΔT mf ,ΔT mr ) is set.

[0157] This allows each drive system S f ,S r Considering the variation in torque response characteristics due to the length of the dead zone, which varies depending on the mechanical characteristics and driving force distribution of the mf * ,T mr * Vibration compensation is performed on the final torque command value T mff * ,T mrf * Therefore, multiple drivetrains Sf ,S r In the electric vehicle 100 equipped with the f ,S r This realizes a control configuration that reduces discomfort felt by occupants due to variations in torque response characteristics resulting from mechanical characteristics and drive force distribution.

[0158] In the vibration suppression process S203 in each of the above-described embodiments and modifications, in particular, the dead-zone damping coefficient ζ is set to make the correction amount in the dead-zone section smaller than the correction amount in sections other than the dead-zone section. γf ,ζ γr is set.

[0159] In this way, the dead-zone damping coefficient ζ that reduces the correction amount in the dead-zone section γf ,ζ γr By setting each drive system S f ,S r Even if the dead zone is usually long due to the mechanical characteristics and driving force distribution of the drivetrain, the torque response in the dead zone is quickened. f ,S r This can reduce or suppress variations in torque response characteristics regardless of the mechanical characteristics or driving force distribution.

[0160] In the vibration suppression process S203 of each of the above-described embodiments and modifications, f ,S r The dead zone damping coefficient ζ at which the timing at which γf ,ζ γr Set.

[0161] In this way, each drive system S f ,S r Each drive system S is set so that the timing at which it passes through the dead zone is substantially constant. f ,S r The dead band damping coefficient ζ is determined according to the mechanical characteristics and driving force distribution of the γf ,ζ γr By setting the above, the variation in the torque response characteristics of the electric vehicle 100 is particularly effectively reduced or suppressed, and the discomfort felt by the occupants due to this is reduced.

[0162] In the vibration suppression process S203 in each of the above-described embodiments and modifications, the dead zone damping coefficient ζ that shortens the timing of exiting the dead zone is γf ,ζ γr Set.

[0163] In this way, each drive system S f ,S r The dead zone damping coefficient ζ is set so that the "timing of leaving the dead zone" (t2 to t3), which can change depending on the mechanical characteristics or driving force distribution, becomes short (for example, the shortest time t2). γf ,ζ γr By setting the above, the variation in the torque response characteristics of the electric vehicle 100 is reduced or suppressed without impairing the torque response performance.

[0164] In particular, in the first embodiment and each of the modifications, each drive system S f ,S r When the combination of mechanical characteristics can change, in the vibration suppression process S203, the dead zone damping coefficient ζ γf ,ζ γr Each drive system S f ,S r It is set (adjusted) according to the combination of mechanical properties.

[0165] In this way, each drive system S f ,S r The gear ratios of each drive system f ,S r The dead zone damping coefficient ζ depending on the combination of mechanical properties γf ,ζ γr By setting (adjusting) each drive system S f ,S r This makes it possible to particularly effectively reduce or suppress variations in the torque response characteristics of the electric vehicle 100, regardless of changes in the mechanical characteristics of the electric vehicle 100.

[0166] In particular, in the second embodiment and each of the modifications, each drive system S f ,S r Driving force distribution (K f ,1-K f) may change, in the vibration suppression process S203, the dead zone damping coefficient ζ γf ,ζ γr The driving force distribution (K f ,1-K f ) and set (adjust) accordingly.

[0167] In this way, each drive system S f ,S r The dead band damping coefficient ζ is calculated according to the driving force distribution. γf ,ζ γr By setting (adjusting) each drive system S f ,S r This makes it possible to particularly effectively reduce or suppress variations in the torque response characteristics of the electric vehicle 100, regardless of the drive force distribution.

[0168] In the vibration suppression process S203 in each of the above-described embodiments and modifications, specifically, f ,S r Drive shaft (DS f ,DS r ) torsional angular velocity estimate (ω df ^,ω dr ^) to a given feedback gain (k f1 ,k f2 ,k r1 ,k r2 ) to obtain the vibration compensation torque (ΔT mf ,ΔT mr ) is calculated. Also, the basic torque command value (T mf * ,T mr * ) to the vibration compensation torque (ΔT mf ,ΔT mr ) to obtain the corrected torque command value (T mf1 * ,T mr1 * , or T mff * ,T mrf * ) is calculated. Then, the feedback gain (k f2 ,k r2 ) and the dead-band damping coefficient ζ γf ,ζ γr The dead band gain k according tof2 ,k r2 By setting mf * ,T mr * ) to adjust the correction amount.

[0169] This allows each drive system S f ,S r For each of these, a specific and excellent control logic for adjusting the torque response characteristics in the dead band section is realized.

[0170] In addition to the above, in particular in the vibration suppression process S203 of the first modified example, each drive system S f ,S r The driving force transmission characteristics and each drive system S f ,S r The calculation is performed using a vehicle model S1301 that includes the mutual driving force transmission characteristics of the above.

[0171] In this way, each drive system S f ,S r The driving force transmission characteristics and each drive system S f ,S r By using the vehicle model S1301 including the mutual driving force transmission characteristics of the drive trains Sf and Sr, particularly good vibration damping processing that reduces the interaction between the drive trains Sf and Sr can be achieved.

[0172] Furthermore, in the vibration suppression process S203 of the first modified example, the vehicle model S1301 is used to calculate the vibration of each drivetrain S f ,S r The estimated twist angle (ω df ^,ω dr ^) and calculate the estimated torsion angle (ω df ^,ω dr ^) to calculate the correction amount (ΔT mf ,ΔT mr ) is calculated.

[0173] In this way, in the process of performing the vibration suppression process S203 using the vehicle model S1301, the correction amount, i.e., the vibration compensation torque (ΔT mf ,ΔT mr) parameters (ω df ^,ω dr ^,θ df ^,θ dr By calculating ^, it is possible to reduce or suppress variations in the torque response characteristics of the electric vehicle 100 while reducing the calculation load.

[0174] In addition to the electric vehicle control method, each of the above-described embodiments and modifications provides a motor controller 2 that functions as an electric vehicle control device for executing the method.

[0175] Specifically, the motor controller 2 controls a plurality of drive systems (S f ,S r In the electric vehicle 100 equipped with each drive system (S f ,S r ) driving force (T mf ,T mr ) for the electric vehicle 100. The motor controller 2 controls the total required driving force (T m * ) and each drive system (S f ,S r ) to the driving force distribution (K f ,1-K f ) based on the basic torque command value (T mf * ,T mr * ) and the basic torque command value (T mf * ,T mr * ) is corrected to suppress vibration in the driving force transmission system, and the corrected torque command value (T mf1 * ,T mr1 * , or T mff * ,T mrf * ) and a vibration suppression unit (S203) that calculates the correction torque command value (T mf1 * ,Tmr1 * , or T mff * ,T mrf * ) based on the driving force (T mf ,T mr The driving force control unit (S204, S205) executes a driving force control process to control each driving system (S f ,S r The damping coefficient of the reference response related to the driving force transmission in the dead zone is the reference damping coefficient ("1"), which is the damping coefficient when the vehicle is in a zone other than the dead zone, and the dead zone damping coefficient ζ, which is the damping coefficient when the vehicle is in the dead zone. γf ,ζ γr and are set as different parameters. f ,S r ) is in the dead zone. Then, for the drive system outside the dead zone, the basic torque command value (T mf * ,T mr * ) is set, and in the drive system in the dead zone, the dead zone damping coefficient ζ γf ,ζ γr Based on this, the basic torque command value (T mf * ,T mr * ) correction amount (ΔT mf ,ΔT mr ) is set.

[0176] The above describes embodiments of the present invention, but the configurations described in the above embodiments and each modified example only show some of the application examples of the present invention and are not intended to limit the technical scope of the present invention. [Explanation of symbols]

[0177] 1: Battery, 2: Motor controller, 3: Inverter, 3f: Front inverter, 3r: Rear inverter, 4: Drive motor, 4f: Front motor, 4r: Rear motor, 5: Transmission, 5f: Front transmission, 5r: Rear transmission, 20: Current sensor, 21: Rotation sensor, 100, 200, 300, 400: Electric vehicle, S1301: Vehicle model, S1302: Front torque correction unit, S1303: Rear torque correction unit, S202: Basic torque distribution processing, S203: Vibration suppression processing, S204, S205: Driving force control processing, S601: Vehicle model, S602: Front torque correction unit, S701: Vehicle model, S702: Rear torque correction unit, S f : Front drivetrain, S r : Rear drivetrain

Claims

1. 1. An electric vehicle control method for controlling the driving force of each drive system in an electric vehicle equipped with a plurality of drive systems each having a drive motor, comprising: a basic torque distribution process for determining a basic torque command value for each of the drive motors based on a total required drive force for the electric vehicle and a drive force distribution for each of the drive trains; a vibration suppression process for correcting each of the basic torque command values ​​to suppress vibration in the driving force transmission system, thereby obtaining a corrected torque command value; a driving force control process for controlling the driving forces generated by the respective driving motors based on the correction torque command values, In the vibration damping treatment, a damping coefficient of a reference response relating to the transmission of driving force in each of the drive trains is set as a reference damping coefficient, which is the damping coefficient when the drive train is in a section other than a dead-band section, and a dead-band damping coefficient, which is the damping coefficient when the drive train is in the dead-band section, as different parameters; individually estimating whether each of the drive systems is in the dead zone; In the drive system outside the dead band section, a correction amount for the basic torque command value is set based on the reference damping coefficient; In the drive system in the dead-band section, a correction amount for the basic torque command value is set based on the dead-band damping coefficient; When a combination of mechanical characteristics of the drive trains changes, the dead-band damping coefficient is set in accordance with the combination of the mechanical characteristics. Electric vehicle control method.

2. 1. An electric vehicle control method for controlling the driving force of each drive system in an electric vehicle equipped with a plurality of drive systems each having a drive motor, comprising: a basic torque distribution process for determining a basic torque command value for each of the drive motors based on a total required drive force for the electric vehicle and a drive force distribution for each of the drive trains; a vibration suppression process for correcting each of the basic torque command values ​​to suppress vibration in the driving force transmission system, thereby obtaining a corrected torque command value; a driving force control process for controlling the driving forces generated by the respective driving motors based on the correction torque command values, In the vibration damping treatment, a damping coefficient of a reference response relating to the transmission of driving force in each of the drive trains is set as a reference damping coefficient, which is the damping coefficient when the drive train is in a section other than a dead-band section, and a dead-band damping coefficient, which is the damping coefficient when the drive train is in the dead-band section, as different parameters; individually estimating whether each of the drive systems is in the dead zone; In the drive system outside the dead band section, a correction amount for the basic torque command value is set based on the reference damping coefficient; In the drive system in the dead-band section, a correction amount for the basic torque command value is set based on the dead-band damping coefficient; When the driving force distribution of each of the drive trains changes, the dead-band damping coefficient is set in accordance with the driving force distribution. Electric vehicle control method.

3. 3. The electric vehicle control method according to claim 1, In the vibration damping process, the dead-zone damping coefficient is set so that the correction amount in the dead-zone section is smaller than the correction amount in sections other than the dead-zone section. Electric vehicle control method.

4. The electric vehicle control method according to any one of claims 1 to 3, In the vibration damping process, the dead-zone damping coefficient is set so that the timing at which each of the drive systems exits the dead-zone interval is constant. Electric vehicle control method.

5. 5. The electric vehicle control method according to claim 4, In the vibration suppression process, the dead-zone damping coefficient is set so that the timing is shortened. Electric vehicle control method.

6. The electric vehicle control method according to any one of claims 1 to 5, In the vibration damping treatment, calculating a vibration compensation torque by multiplying the estimated torsional angular velocity of the drive shaft in each of the drive systems by a predetermined feedback gain; determining the correction torque command value by subtracting the vibration compensation torque from the basic torque command value; The correction amount is adjusted by setting the feedback gain in the dead-band section to a dead-band gain according to the dead-band damping coefficient. Electric vehicle control method.

7. 7. The electric vehicle control method according to claim 6, The vibration damping process is performed using a vehicle model that includes a driving force transmission characteristic in each of the drive systems and a mutual driving force transmission characteristic between the drive systems. Electric vehicle control method.

8. 8. The electric vehicle control method according to claim 7, In the vibration damping treatment, calculating an estimated torsion angle of each of the drivetrains using the vehicle model; calculating the correction amount using the torsion angular velocity estimated value; Electric vehicle control method.

9. An electric vehicle control device that controls the driving force of each drive system in an electric vehicle equipped with multiple drive systems each having a drive motor, a basic torque distribution unit that determines a basic torque command value for each of the drive motors based on a total required drive force for the electric vehicle and a drive force distribution for each of the drive trains; a vibration suppression unit that performs corrections on the respective basic torque command values ​​to suppress vibrations in the driving force transmission system, thereby obtaining corrected torque command values; a driving force control unit that executes a driving force control process to control the driving forces generated by the respective driving motors based on the correction torque command values, The vibration damping unit is a damping coefficient of a reference response relating to the transmission of driving force in each of the drive trains is set as a reference damping coefficient, which is the damping coefficient when the drive train is in a section other than a dead-band section, and a dead-band damping coefficient, which is the damping coefficient when the drive train is in the dead-band section, as different parameters; individually estimating whether each of the drive systems is in the dead zone; In the drive system outside the dead band section, a correction amount for the basic torque command value is set based on the reference damping coefficient; In the drive system in the dead-band section, a correction amount for the basic torque command value is set based on the dead-band damping coefficient; When a combination of mechanical characteristics of the drive trains changes, the dead-band damping coefficient is set in accordance with the combination of the mechanical characteristics. Electric vehicle control device.

10. An electric vehicle control device that controls the driving force of each drive system in an electric vehicle equipped with multiple drive systems each having a drive motor, a basic torque distribution unit that determines a basic torque command value for each of the drive motors based on a total required drive force for the electric vehicle and a drive force distribution for each of the drive trains; a vibration suppression unit that performs corrections on the respective basic torque command values ​​to suppress vibrations in the driving force transmission system, thereby obtaining corrected torque command values; a driving force control unit that executes a driving force control process to control the driving forces generated by the respective driving motors based on the correction torque command values, The vibration damping unit is a damping coefficient of a reference response relating to the transmission of driving force in each of the drive trains is set as a reference damping coefficient, which is the damping coefficient when the drive train is in a section other than a dead-band section, and a dead-band damping coefficient, which is the damping coefficient when the drive train is in the dead-band section, as different parameters; individually estimating whether each of the drive systems is in the dead zone; In the drive system outside the dead band section, a correction amount for the basic torque command value is set based on the reference damping coefficient; In the drive system in the dead-band section, a correction amount for the basic torque command value is set based on the dead-band damping coefficient; When the driving force distribution of each of the drive trains changes, the dead-band damping coefficient is set in accordance with the driving force distribution. Electric vehicle control device.

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