JP2005170194A - Driving force control device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve operability of a driving force control device for a vehicle when starting. <P>SOLUTION: This driving force control device for a vehicle is provided with a target acceleration/deceleration setting means S4 for setting target acceleration/ deceleration (α*<SB>w</SB>) of the vehicle on the basis of the driving condition of the vehicle, a feed-forward compensating unit 31 for outputting feed-forward operation quantity T*<SB>p-ff</SB>, I*<SB>p-ff</SB>to the vehicle on the basis of the target acceleration/ deceleration, a feedback compensating unit 33 for outputting feedback operation quantity T*<SB>p-fb</SB>, I*<SB>p-fb</SB>to the vehicle on the basis of a difference Δα<SB>w</SB>between estimated acceleration/deceleration α<SB>w</SB>of the vehicle estimated on the basis of the wheel speed V<SB>w</SB>of the vehicle and the target acceleration/deceleration α*<SB>w</SB>, and an operation quantity setting means S13 for setting operation quantity T*<SB>p</SB>, I*<SB>p</SB>to the vehicle on the basis of the feed-forward operation quantity and the feedback operation quantity. A feedback operation quantity inhibiting means S4 and S5 are provided to inhibit setting of operation quantity to the vehicle based on the feedback operation quantity to set the operation quantity of the vehicle on the basis of the only feed-forward operation quantity in the case wherein the wheel speed of the vehicle does not exceeds the predetermined speed V1. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

  The present invention relates to a driving force control apparatus for a vehicle.

As a conventional driving force control device, there is one that controls a driving force of a vehicle so that an actual wheel acceleration matches a target acceleration calculated from a driver's accelerator pedal operation amount or the like (see Patent Document 1). In this conventional technology, a two-degree-of-freedom control method composed of a feed-forward control unit for adjusting responsiveness to a target value and a feedback control unit for adjusting stability and disturbance resistance is applied to acceleration control. doing. Further, the wheel acceleration as a feedback signal is inputted by subjecting a wheel speed that can be directly measured by an inexpensive sensor to a band pass filter (hereinafter referred to as BPF) process. In other words, only the low frequency range that is not easily affected by sensor noise is subjected to differential processing approximately.
Japanese Patent Application No. 2002-097135

  However, a normal wheel speed sensor measures the frequency of a pulse corresponding to the wheel speed, and accurate measurement is impossible unless the wheel speed actually increases to some extent. Accordingly, the wheel acceleration (estimated value) obtained by subjecting the wheel speed thus measured to the BPF processing described above becomes an inaccurate value in the one pole low speed region. That is, acceleration feedback control cannot be performed when starting acceleration is performed from a vehicle stopped state. For example, as shown in FIG. 12, when accelerating from a stopped state, the wheel speed fluctuates suddenly from a zero value. If this value is processed as it is, a large overshoot occurs in the wheel acceleration (estimated value). End up. If such a signal is used as an acceleration feedback signal, the vehicle becomes vibrated at the time of starting acceleration, and the drivability is deteriorated contrary to the aim.

  The present invention has been developed in view of these problems, and an object of the present invention is to provide a driving force control apparatus that prevents the control amount overshoot by the feedback compensator from being suppressed and improves vehicle drivability. Is.

  In order to solve the above problem, a driving force control apparatus according to claim 1 of the present invention includes target acceleration / deceleration setting means for setting a target acceleration / deceleration of a vehicle based on a driving state of the vehicle, and the target acceleration / deceleration to the vehicle. A feedforward compensator that outputs a feedforward manipulated variable of the vehicle, a feedback compensator that outputs a feedback manipulated variable to the vehicle based on a difference between the estimated acceleration / deceleration of the vehicle estimated from the wheel speed of the vehicle and the target acceleration / deceleration, and the feed An operation amount setting means for setting an operation amount to the vehicle based on the forward operation amount and the feedback operation amount; and when the vehicle wheel speed does not exceed a predetermined speed, The setting of the operation amount to the vehicle based on the feedback operation amount is prohibited, and the operation amount to the vehicle is set only by the feedforward operation amount. Provided to feedback manipulated variable prohibiting means - that Fi.

  According to the driving force control apparatus of the first aspect of the present invention, when starting acceleration is performed from a vehicle stop state or a state close thereto, first, acceleration control is performed only by a feedforward compensator having excellent responsiveness to a target value. Next, as soon as the measurement accuracy of the wheel acceleration is obtained, the feedforward compensator + feedback compensator (2-degree-of-freedom control) performs acceleration control, so in addition to quick response, disturbance resistance (influence of various disturbances) The feedback effect is suppressed. That is, it is possible to improve the drivability by improving the followability of the actual acceleration to the target acceleration even at the time of starting acceleration.

  Hereinafter, an embodiment in which a feedback control device of the present invention is applied to a vehicle acceleration / deceleration control device will be described with reference to the accompanying drawings.

  FIG. 1 is a schematic configuration diagram showing an embodiment of an acceleration / deceleration control apparatus according to this embodiment. In the figure, reference numeral 1 denotes an engine, reference numeral 2 denotes a continuously variable transmission, reference numeral 3 denotes a torque converter with a lockup mechanism interposed between the engine 1 and the continuously variable transmission 2, and reference numeral 4 denotes a drive wheel. The engine 1 is configured to be able to control the engine torque by adjusting the opening of the throttle valve 12 by the throttle actuator 11 and controlling the intake air amount. The continuously variable transmission 2 is a so-called belt-type continuously variable transmission, and the gear ratio is controlled by controlling the belt contact radii of the primary pulley (input side pulley) 13 and the secondary pulley (output side pulley) 14. It is comprised so that it can control. The secondary pulley 14 of the continuously variable transmission 2 is connected to the drive wheels 4 via the final reduction gear 15. The torque converter 3 includes a lockup clutch 16.

  The engine 1 is controlled by an engine controller 7. Therefore, a crank angle sensor 21 for detecting the rotational speed of the engine 1 is provided, and the operating state of the engine 1 is controlled based on the detected value. The continuously variable transmission 2 and the lockup clutch 16 of the torque converter 3 are controlled by the transmission controller 5. Therefore, the primary speed sensor 22 for detecting the rotational speed of the primary pulley 13, that is, the transmission input shaft rotational speed, and the secondary output for detecting the rotational speed of the secondary pulley 14, which is the rotational speed of the transmission output shaft and also the traveling speed of the vehicle. A speed sensor 23 is provided, and the gear ratio of the continuously variable transmission 2 and the engagement state of the lockup clutch 16 are controlled based on the detected value. Incidentally, in the present embodiment, the lockup clutch 16 is released only in the extremely low speed range, and is engaged in almost all speed ranges except that it can be stopped and started.

  Further, this vehicle has an acceleration / deceleration for controlling the acceleration / deceleration of the host vehicle. The acceleration / deceleration controller 6 is connected to the engine torque controller 7 and the transmission controller 5 through a high-speed communication line, and the information, the accelerator opening detected by the accelerator sensor 24 and the wheel speed detected by the wheel speed sensor 23. The acceleration / deceleration of the host vehicle is controlled based on the above. Specifically, based on the difference between the target acceleration / deceleration and gear ratio and the actual acceleration / deceleration and gear ratio, target values of the transmission input torque and gear ratio are set, and these are set as the engine controller 4 and the gear shift respectively. It outputs toward the machine controller 5 to control the acceleration / deceleration of the host vehicle. The conventional engine torque controller controls the engine torque according to the accelerator opening and the engine speed, and the conventional transmission controller controls the gear ratio based on the accelerator opening, the engine speed and the traveling speed. The acceleration / deceleration required by the driver and the fuel consumption were compatible to some extent, but in order to further improve the acceleration / deceleration and the fuel consumption, an acceleration / deceleration controller that takes the entire vehicle into account is provided. They are controlled in accordance with the transmission input torque calculated, that is, the engine torque and the gear ratio. Each controller includes an arithmetic processing unit such as a microcomputer.

In this vehicle, the acceleration / deceleration control system is configured as shown in FIG. The plant model 34 in the figure is the host vehicle. The output of the vehicle is the acceleration alpha _w and the transmission input shaft rotation speed omega _p. For example, the target acceleration / deceleration α ^* _w is determined from the accelerator opening Apo and the traveling speed, that is, the wheel speed V _w, and the target is determined from the engine rotational speed, that is, the transmission input shaft rotational speed and the engine torque, that is, the transmission input torque. When the transmission input shaft rotation speed omega ^* _p is determined, the feedforward compensator 31, the transfer in accordance with the function G _FF (s), feed forward control amount of the target transmission input torque command value from the target acceleration alpha ^* _w Calculate and set T ^* _p-ff and the feedforward control amount I ^* _p-ff of the target gear ratio command value. On the other hand, the normative model unit 32 calculates and sets the normative acceleration / deceleration α _w-ref and the normative transmission input shaft rotation speed ωp _-ref according to a predetermined normative model G _M (s), and the adder / subtractors 35 and 36 are respectively calculated. in the acceleration alpha _w and the transmission input shaft rotation speed omega is subtracted _p acceleration difference △ alpha _w and the transmission input shaft rotational speed difference (- △ ω _p) is calculated. The feedback compensator 33 feeds back the target transmission input torque command value according to a predetermined transfer function G _FB (s) with respect to the acceleration / deceleration difference Δα _w and the transmission input shaft rotational speed difference (−Δω _p ). The control amount T ^* _p-fb and the feedback control amount I _p-fb of the target gear ratio command value are calculated and set. The target transmission input torque command value T ^* _p-ff and the target transmission input torque command value feedback control value T ^* _p-fb are added by the adder 37 to obtain the target transmission input. The torque command value T ^* _p is calculated, and the adder 38 adds the feedforward control component I ^* _p-ff of the target gear ratio command value and the feedback control component I ^* p-fb of the target gear ratio command value. A target gear ratio command value I ^* p is calculated.

FIG. 3 shows a reference transmission input for calculating a reference transmission input shaft rotational speed ω _p-ref from the target transmission input torque command value T ^* _p in the vehicle model as the plant model 34 and the reference model unit 32. The shaft rotation speed calculation unit 32a and the adder / subtractor 36 are shown. First, in the vehicle that is the plant model 34, the upper / lower limiter 301 regulates the target transmission input torque command value T ^* _p according to the transmission input shaft rotational speed ωp (substantially the feedback compensator). 33), the value becomes the transmission input torque T _p via the primary slow red engine torque control system 302. On the other hand, in another on the lower limiter 303, in response to the wheel speed V _w by regulating the target speed ratio command value I ^* _p (is substantially performed in the said feedback compensator 33), the value is a primary delay The transmission ratio I _p and the transmission ratio change rate I ′ _p are obtained via the transmission ratio control system 304 of the system. The wheel speed V _w, in divider 305, the effective tire radius dividing to the wheel angular velocity omega _w is obtained, multiplied by the the wheel angular velocity omega _w and the gear ratio rate of change r _p at the multiplier 306, further The multiplier 307 multiplies the drive system inertia J1 and the final reduction ratio _If to obtain an inertia torque _Tine .

Accordingly, a value obtained by subtracting the inertia torque T _ine from the transmission input torque T _p by the adder / subtractor 308 becomes the drive torque T _w . This drive torque T _w, at the multiplier 309, multiplied by the speed ratio I _p, further multiplier 310, a driving force F _w by dividing the and the tire effective radius R multiplied by the final drive ratio I _f. Further, since the traveling resistance system 311 obtains the traveling resistance force F corresponding to the wheel speed V _w , the value obtained by subtracting the traveling resistance force F _r by the adder / subtractor 312 from the driving force F _w is the wheel driving force F. _d becomes, in divider 313 this next wheel acceleration alpha _w by dividing the vehicle mass M, by integrating further integrator 314 becomes the wheel speed V _w. Further, the speed change ratio I _p is multiplied by the wheel angular speed ω _w by the multiplier 316 and further multiplied by the final reduction ratio _If by the multiplier 317 to obtain the transmission input shaft rotational speed ω _p . In the present embodiment, the wheel speed V _w is passed through the band pass filter 315 to calculate the wheel acceleration / deceleration α _wf .

Meanwhile, the in normative transmission input shaft rotational speed calculation unit 32a, from the target transmission input torque command value T ^* _p via the engine torque control system 302 as described above is the transmission input torque T _p, the transmission From the machine input torque T _p , the target transmission input shaft rotation speed ω ^* _p is calculated and set by the target transmission input shaft rotation speed setting unit 318 according to the engine operation restriction map ^. _{The reference} transmission input shaft rotational speed ω _p-ref is obtained by normalizing _p by the reference model unit 319 of the first order lag system. The engine operation restriction map will be described in detail later.

FIG. 4 shows the feedforward compensator 31. In the feedforward compensator 31, first, the target acceleration / deceleration α ^* _w is multiplied by the vehicle mass M by the multiplier 401 to obtain the target wheel driving force F ^* _d . On the other hand, the flat road running resistance force calculation unit 402 calculates a flat road running resistance force F _r according to the wheel speed V _w according to the flat road running resistance map, and the flat road running resistance force F _r and the target wheel driving force are calculated. The target driving force F ^* _w is obtained by adding P _d with the adder 403. The target driving force F ^* _w is multiplied by the tire effective radius R by the multiplier 404 and divided by the final reduction ratio _If , and further divided by the speed ratio _Ip by the subtractor 405, thereby feed-forward control. Target transmission input torque command value T ^* _po is obtained. A formula for calculating the target transmission input torque command value T ^* _po for feedforward control is shown in the following one formula.

Then, the acceleration / deceleration model matching compensator 406 obtains a feedforward control amount T ^* _{p-ff of the} transmission input torque command value from the target transmission input torque command value T ^* _po for feedforward control. The transfer function G _FF-1 (s) of the acceleration / deceleration model matching compensator 406 is shown in the following two equations. This acceleration / deceleration model matching compensator 406 standardizes using a normative model relating to acceleration / deceleration (denominator in the equation), and at the same time, a first-order delay transmission input torque control system, that is, an engine torque control system Phase matching is performed by multiplying the reciprocal of the response delay (numerator in the equation). In the equation, s is a Laplace operator, τ _eng is a response delay time constant of the engine torque control system, ω _r , ζ _r are reference model responses of the acceleration / deceleration α _{p to} the target acceleration / deceleration α ^* _p (secondary delay model) ) Cutoff frequency and damping constant.

However, in order to perform arithmetic processing by the microcomputer described above, for example, it is discretized by Tustin approximation or the like, and a differential equation that can be executed by software is obtained as shown by the following equation 2a to obtain the transmission input torque command value. The feedforward control amount T ^* _p-ff is calculated. Note that MTN0, MTN1, MTN2, MTD1, and MTD2 in the equation are constants determined from the time constant τ _eng , the cutoff frequency ω _r , the damping constant ζ _r , and the sampling period ΔT of the arithmetic processing. Further, (k) indicates the current value, (k-1) indicates the previous value, and (k-2) indicates the same value immediately before.

On the other hand, the when multiplying the target driving force F ^* _w and the wheel speed V _w at the multiplier 407 since the target engine power (output) P ^* is obtained, the target transmission input rotation speed setting unit 408 in the engine operating constraints described above The target transmission input rotational speed ω _p that achieves the target engine power P ^* and obtains the optimum fuel consumption is calculated and set using the map. The following three equations are used to calculate the target engine power P ^* .

Accordingly, the divider 409 divides the target transmission input rotational speed ω ^* _p by the wheel speed V _w , and the multiplier 410 multiplies the tire effective radius R and divides by the final reduction ratio _If , A target speed ratio I ^* _po for feedforward control is obtained. The following four equations are used to calculate the feedforward control target gear ratio I ^* _po .

Then, the transmission ratio model matching compensator 411 obtains the feedforward control component I ^* _{p-ff of the} transmission ratio command value from the target forward transmission control ratio I ^* _po . The transfer function G _FF-2 (s) of the gear ratio model matching compensator 411 is shown in the following equation (5). The transmission ratio model matching compensator 411 normalizes using a reference model relating to the transmission ratio (denominator in the equation), and at the same time, the reciprocal of the response delay of the transmission ratio control system of the first-order lag system ahead of the output (in the equation) Phase alignment is performed by multiplying by (numerator). In the equation, τ _cvt is the response delay time constant of the gear ratio control system, and τ _ref-wp is the reference model response (first-order lag model) of the transmission input shaft rotational speed ω _{p to} the target transmission input shaft rotational speed ω ^* _p . ) Time constant.

However, in order to perform arithmetic processing by the above-described microcomputer, for example, it is discretized by Tustin approximation or the like, and a difference equation that can be executed by software is obtained as shown by the following equation 5a to feed forward the transmission ratio command value. The control component I ^* _p-ff is calculated. Note that MIN0, MIN1, and MID1 in the equation are constants determined from the sampling period ΔT of the time constants τ _cvt and τ _ref-wp arithmetic processing.

Next, the engine operation restriction map will be described with reference to FIG. For example, when the engine speed ω _e (= transmission input rotational speed ω _p ) is taken on the horizontal axis and the engine torque _Te (= transmission input torque T _p ) is taken on the vertical axis as shown in FIG. An equal output line (broken line in the figure) connecting power (output) and an equal fuel consumption line (one-dot chain line in the figure) centering on the optimum fuel consumption point can be drawn. A curve in which the optimum fuel consumption points on the iso-output line are continuous becomes the optimum fuel consumption driving line. Generally, in recent engines, fuel is not injected with the accelerator off, so that the optimal fuel consumption point and the iso fuel consumption line exist only in the region where the engine torque _Te is positive. Therefore, the optimum fuel consumption driving line exists only in the region where the engine torque _Te is positive. Conversely, in a region where the engine torque _Te is negative, an engine brake characteristic line showing the relationship between the engine brake torque and the engine speed appears. As described above, since fuel is not injected when the engine torque _Te is negative, it is necessary to control the engine speed in order to control the engine brake torque. In this embodiment, since a continuously variable transmission is used as the transmission, in order to obtain a desired engine brake torque at an arbitrary traveling speed, the speed ratio of the continuously variable transmission may be controlled. An engine operation restriction map is obtained by mapping the relationship between these curves in consideration of fuel consumption.

Next, a design method of the feedback compensator 33 in this embodiment will be briefly described. The above-described non-linear control target model of FIG. 3 has two outputs of the transmission input torque command value T ^* _p and the gear ratio command value I ^* _p , the acceleration / deceleration α _w , and the transmission input shaft rotational speed difference Δω _p . It is assumed that the model is a nonlinear control target model. In the present embodiment, for the purpose of ensuring the stability of the feedback control system, the detection unit and the partial control unit are combined with a vehicle model as a controlled object model. This nonlinear control target model is linearly approximated at a specific operating point to derive a linear approximate control target model for control system design. In order to design a feedback compensator using “μ synthesis”, which is one of “robust control theory”, it is necessary to further model variable elements and extend them to a generalized plant model. Will be omitted.

Since the input to the feedback compensator 33 is the acceleration / deceleration difference Δα _w and the transmission input shaft rotational speed difference (−Δω _p ) as described above, the output transmission input torque command value T ^* _p , When the ratio command value I ^* _{p is represented} by the transfer function G _FB (s) of the feedback compensator 33, the following six formulas are obtained, and each element of the transfer function G _FB (s) is represented by the seven formulas.

Substituting actual vehicle specifications or required response characteristics and obtaining each element G ₁₁ (s) to G ₂₂ (s) by the “μ synthesis”, the following equations 8 to 11 are obtained.

Considering each of these elements G ₁₁ (s) to G ₂₂ (s) in detail, there is a portion where the pole is on or near the imaginary axis. The pole on or in the vicinity of the imaginary axis is a pole with a slow response and accumulates the difference between the target value and the actual value, in other words, it is a part having an integral characteristic. Therefore, the transfer function of the feedback compensator of Equation 7 is separated into a portion G _A (s) having integral characteristics and a portion G _B (s) other than that, and is expressed as the following Equation 12.

Specific elements G _11-A (s) to G _22-A (s) and G _11-B (s) to G _22-B (s) are expressed by the following 13 to 20 equations.

Then, elements G _11-B (s) to G _22-B (s) that do not have the integral characteristics are added to the acceleration / deceleration α _w and the transmission input shaft rotational speed difference (−Δω _p ). X ₁₁ to x _22, and the elements G _11-A (s) to G _22-A (s) having the integral characteristics in these elements are y ₁₁ to y _22. The feedback control component T ^* _p-fb and the gear ratio command value feedback control component I ^* _p-fb are expressed by the following equations (21) and (22), respectively.

  Next, arithmetic processing performed by the feedforward compensator 31 and the feedback compensator 33 will be described with reference to the flowchart of FIG. This calculation process is performed at a predetermined sampling period ΔT of about 10 msec., For example. Although all steps for communication are not described in this calculation process, necessary information is exchanged with other controllers or storage devices at any time, and information obtained in the calculation process is always changed with other controllers or storage devices. Exchanged with storage device.

  In this calculation process, first, in step S1, the accelerator opening Apo detected by the accelerator sensor 24 is read.

Next, the process proceeds to step S2, and the wheel speed V _w detected by the wheel speed sensor 25 is read.

Next, the process proceeds to step S3 where the primary rotational speed ω _p , the secondary rotational speed ω _s , and the speed ratio I _p , which is the ratio of both, are read from the transmission controller 5 and the engine rotational speed ω is read from the engine torque controller 7. Read _e .

Next, the process proceeds to step S4, and in accordance with the control map shown in FIG. 7, the target acceleration / deceleration α ^* _w is determined based on the accelerator opening Apo read in step S1 and the wheel speed V _w read in step S2. Set the calculation.

Next conditional branch by comparing the wheel speed V _w and the predetermined value V1 proceeds to step S4.5. If V _w ≧ V1, the process proceeds to step S5 to perform normal bandpass filter processing. Otherwise, the process proceeds to step S5.5 to initialize the bandpass filter. Predetermined value V1 is a threshold for the accuracy of the wheel speed V _w read in step 2 it is determined whether sufficient or not in utilizing the feedback control. For example, the wheel speed is such that the number of pulse edges read from the wheel speed sensor within a calculation cycle is equal to or greater than a predetermined value.

Next, the process proceeds to step S5, and a wheel acceleration / deceleration speed α _w only in a predetermined frequency range is calculated by removing noise from the wheel speed V _w by using a bandpass filter of a transfer function G _bp (s) expressed by the following equation (23). To do. In the equation, ω _n is the natural angular frequency, ζ _n is the attenuation rate, and ω _n and ζ _n are determined by the detected noise level of the wheel speed.

However, in order to perform the arithmetic processing by the microcomputer described above, the wheel acceleration / deceleration α _w is calculated by obtaining a differential equation that can be executed by software as shown by the following equation 23a after being discretized by, for example, Tustin approximation. . BPN0, BPN1, and BPN2 in the equation are constants determined by the natural angular frequency ω _n , the attenuation rate ζ _n , and the sampling period ΔT.

  In step S5, Timer1 is further decremented to zero. This timer measures the elapsed time of the bandpass filter processing in S5, and is referred to in step S9.5.

In step S5.5, the band pass filter is initialized. Specifically, a value is substituted for each variable used in equation 23a as in the following equation. Also, the set time T ₁ is substituted for Timer1. The set time T ₁ is set in consideration of the wheel acceleration estimation delay by the band-pass filter, that is, the natural angular frequency ω _n in Equation 23.

At the next step S6, the calculated transmission input torque T _p by the engine torque response delay model for example from the transmission input torque gist command value T ^* _p of the previous operation. This engine torque response delay model is a first-order delay system represented by the following 25 transfer functions.

However, in order to perform arithmetic processing by the microcomputer described above, for example, it is discretized by Dustin approximation or the like, and a differential equation that can be executed by software is calculated as shown by the following equation 25a to calculate the transmission input torque T _p . To do. In the equation, TEN0, TEN1, and TEN2 are constants determined by the time constant τ _eng and the sampling period ΔT.

Next, the routine proceeds to step S7, calculates the transmission input shaft rotational speed difference △ omega _p of the norm transmission input shaft rotational speed omega _p-ref and the transmission input shaft rotation speed omega _p by the engine operating constraints described above To do. Specifically, first, the transmission input shaft rotational speed on the optimum fuel consumption driving line or engine brake characteristic line is set as the target transmission input shaft rotational speed ω ^* _p from the transmission input torque T _p calculated in step S6. _Is normalized using a reference model response characteristic represented by the following 26 transfer function G _ref-wp (s) to calculate a reference transmission input rotational speed ω _p-ref .

However, in order to perform arithmetic processing by the microcomputer described above, for example, it is discretized by Tustin approximation or the like, and a differential equation that can be executed by software is obtained as shown by the following equation 26a to obtain a reference transmission input rotational speed ω _{p. -ref} is calculated. Note that PRN0, PRN1, and PRD2 in the equation are constants determined by the time constant τ _ref-wp and the sampling period ΔT.

Then, as shown in the following equation 27, the transmission input shaft rotational speed ω _p is subtracted from the obtained reference transmission input rotational speed ω _p-ref to calculate the transmission input shaft rotational speed difference Δω _p .

In step S8, the feedforward compensator 31 feeds the target transmission input torque command value T ^* _p-ff and the target transmission ratio command value I ^* _p-ff. Is calculated.

At the next step S9, it calculates an acceleration difference △ alpha _w with feedback control norms acceleration alpha _w-ref and acceleration alpha _w. Specifically, the delay compensation (second-order lag model) corresponding to the reference model response of the acceleration / deceleration composed of the transfer function G _ref-a expressed by the following equation 28, and the acceleration / deceleration similarly composed of the transfer function G _bp (s) The reference acceleration / deceleration α _w-ref is calculated by applying delay compensation (second-order delay model) corresponding to the bandpass filter for calculation to the target acceleration / deceleration α ^* _w .

However, in order to perform arithmetic processing by the microcomputer described above, for example, it is discretized by Tustin approximation or the like, and a differential equation that can be executed by software is obtained as shown by the following equation 28a to obtain a reference acceleration / deceleration α _w-ref Is calculated. In the equation, REN0, REN1, REN2, REN3, REN4, RED1, RED2, RED3, RED4 are the cutoff frequency ω _r , damping constant ζ _r , natural angular frequency ω _n , attenuation rate ζ _n , sampling period It is a constant determined from ΔT. In addition, (k-4) indicates the value two times before.

  Next, the process proceeds to step S9.5, where a conditional branch is made based on Timer1. If Timer1 is zero, the process proceeds to step S10 and the feedback compensator is operated. Otherwise, the process proceeds to step S12.5, and the feedback compensator is initialized.

Next, the process proceeds to step S10, and the feedback compensator G _11-B excluding the integral characteristic with respect to the acceleration / deceleration difference Δα _w and the transmission input shaft rotation speed difference (−Δω _p ) as described above. _{(s) ~G 22-B (} s) according to the transmission input torque command value of the feedback control amount T ^* _p-fb element x _11, x _12, and the gear ratio of the command value of the feedback control amount I ^* _p-fb Elements x ₂₁ and x ₂₂ are calculated.

Next, the process proceeds to step S11, and the elements x ₁ , x ₁₂ of the feedback control component T ^* _p-fb of the transmission input torque command value calculated in step S10 are calculated according to the arithmetic processing of FIG. A transmission using feedback compensators G _11-A (s) to G _22-A (s) having the above integral characteristics with respect to the elements x ₂₁ and x ₂₂ of the feedback control component I ^* _p-fb of the gear ratio command value. The elements y ₁₁ and y ₁₂ of the feedback control component T ^* _p-fb of the input torque command value and the elements y ₂₁ and y ₂₂ of the feedback control component I ^* _p-fb of the transmission ratio command value are calculated.

At the next step S12, the 21 formula, according to 22 wherein calculating a feedback control amount T ^* _p-fb transmission input torque command value from the sum of the elements y _11, y ₁₂ calculated at step S11 At the same time, the feedback control component I ^* _p-fb of the gear ratio command value is calculated from the added value of the elements y ₂₁ and y ₂₂ .

  In step S12, Timer2 is further decremented to zero. This timer measures the elapsed time since the start of feedback compensation, and is referred to in step S13.

Next, the process proceeds to step S12.5, where the feedback compensator is initialized. Specifically, all the variables of the difference equation obtained by discretizing equations 13 to 20 are all cleared to zero. In addition, substituting the set time T ₂ to Timer2. The set time T ₂ sets a transient time for gradually adding the output of the feedback compensator to the output of the feedforward compensator.

Next, the process proceeds to step S13, and according to the following formulas 29 and 30, the transmission input is determined from the addition value of the feedforward control portion T ^* _p-ff and the feedback control portion T ^* _p-fb of the transmission input torque command value. The torque command value T ^* _p is calculated, and the gear ratio command value I ^* _p is calculated from the addition value of the feedforward control component I ^* _p-ff and the feedback control component I ^* _p-fb .

Next, the process proceeds to step S14, and a restriction process is performed on the transmission input torque command value T ^* _p and the transmission ratio command value I ^* _p from the control maps shown in FIGS. That is, the target value is limited by the upper and lower limit values of the control amount that can actually be generated. In step S15, the transmission input torque command value T ^* _p and the transmission ratio command value I ^* _p are output to the engine torque controller 7 and the transmission controller 5, respectively. Return to.

Next, the minor program executed in step S11 of the calculation process of FIG. 6 will be described with reference to the flowchart of FIG. Or this operation process, first in step S21, it is the previous value T ^* _p of the transmission input torque command value (k-1) is and the acceleration speed difference at the upper limit value △ alpha _w is "0" or more, or shift If the machine input torque command immediately preceding value of the value tT ^* (k-1) is and the acceleration speed difference at the lower values △ alpha _w is "0" determines whether less or is, any of the conditions are satisfied If not, the process proceeds to step S22. If not, the process proceeds to step S23.

In the step S23, the relative feedback control amount T ^* _p-fb elements x ₁₁ transmission input torque command value, the transmission input torque command by the feedback with integral characteristic compensator G _11-A (s) After calculating the element y ₁₁ of the value feedback control component T ^* _p-fb , the process _proceeds to step S24. Specifically, the current value y ₁₁ (k) of the element is updated based on the difference equation obtained by discretization as described above.

In step S22, the previous value y ₁₁ (k-1) of the feedback control component T ^* _p-fb of the transmission input torque command value is output as the current value y ₁₁ (k), and then the process proceeds to step S24. Transition. Essentially, the current value y ₁₁ (k) of the elements of the difference equation described in step S23 is not updated, and the previous value y ₁₁ (k−1) is stored. At the step S24, whether the previous value T ^* _p of the transmission input torque command value (k-1) is and the transmission input shaft rotational speed difference at the upper limit △ omega _p is "0" or more, or transmission It is determined whether or not the previous value T ^* _p (k−1) of the input torque command value is a lower limit value and the transmission input shaft rotational speed difference Δω _p is “0” or less. If satisfied, the process proceeds to step S25, and if not, the process proceeds to step S26.

At the step S26, the relative transmission input torque command value of the feedback control amount T ^* _p-fb elements x ₁₂ of the transmission input torque command by the feedback with integral characteristic compensator G _11-A (s) After calculating the element y ₁₂ of the value feedback control component T ^* _p-fb , the process _proceeds to step S27. Specifically, the current value y ₁₂ (k) of the element is updated based on the difference equation obtained by discretization as described above.

In step S25, the previous value y ₁₂ (k-1) of the feedback control component T ^* _p-fb of the transmission input torque command value is output as the current value y ₁₂ (k), and then the process proceeds to step S27. Transition. Essentially, the current value y ₁₂ (k) of the elements of the difference equation described in step S26 is not updated, and the previous value y ₁₂ (k−1) is stored.

At the step S27, whether the previous value I ^* _p (k-1) is and the acceleration speed difference at the upper limit value △ alpha _w of the gear ratio command value is a predetermined value △ alpha _w1 above, or the previous gear ratio command value It is determined whether or not the value I ^* _p (k−1) is a lower limit value and the acceleration / deceleration difference Δα _w is equal to or _smaller than a predetermined value Δα _w2 . If any of the conditions is satisfied, step S28 is performed. If not, the process proceeds to step S29. The predetermined value Δα _w1 is a positive value, meaning an acceleration command, and the predetermined value Δα _w2 is a negative value, meaning a deceleration command.
In the step S29, to the feedback control amount I ^* _p-fb elements x ₂₁ of the speed change ratio command value, the feedback compensator G _21-A (s) by the feedback control of the gear ratio command value with the integral characteristic After calculating the element y ₂₁ of the minute I ^* _p-fb, the process _proceeds to step S30. Specifically, the current value y ₂₁ (k) of the element is updated based on the difference equation obtained by discretization as described above.

In step S28, the previous value y ₂₁ (k−1) of the feedback control component I ^* _p-fb of the gear ratio command value is output as the current value y ₂₁ (k), and then the process proceeds to step S30. . In practice, the current value y ₂₁ (k) of the elements of the difference equation described in step S29 is not updated, and the previous value y ₂₁ (k−1) is stored.

In step S30, the previous value I ^* _p (k−1) of the gear ratio command value is the upper limit value and the transmission input shaft rotational speed difference Δω _p is “0” or more, or the gear ratio command value. When the previous value I ^* _p (k−1) of the current value is the lower limit value and the transmission input shaft rotational speed difference Δω _p is “0” or less, and any of the conditions is satisfied Shifts to step S31, otherwise shifts to step S32.

At the step S32, the relative feedback control amount I ^* _p-fb elements x ₂₂ of the gear ratio command value, the feedback compensator G _22-A (s) by the feedback control of the gear ratio command value with the integral characteristic After calculating the component y ₂₂ of the minute I ^* _p-fb, the process proceeds to step S12 of the calculation process of FIG. Specifically, the current value y ₂₂ (k) of the element is updated based on the difference equation obtained by discretization as described above.

In step S31, the previous value y ₂₂ (k−1) of the feedback control component I ^* _p-fb of the gear ratio command value is output as the current value y ₂₂ (k), and then the calculation process of FIG. The process proceeds to step S12. Essentially, the current value y ₂₂ (k) of the elements of the difference equation described in step S32 is not updated, and the previous value y ₂₂ (k−1) is stored.

According to these processing, feed forward control amount of target acceleration α feedforward control component of the target transmission input torque command value corresponding to _w T ^* _p-ff, and target speed ratio command value I ^* _p-ff There is calculated and set, also acceleration difference △ alpha _w and the transmission input shaft rotational speed difference (- △ ω _p) to the transmission input torque command value of the feedback control amount T ^* _p-fb and the speed ratio command value in accordance with The feedback control component I ^* _p-fb is calculated and set, and the transmission input torque command value T ^* _p and the gear ratio command value I ^* _p are calculated and set from the added value of both. However, when the transmission input torque command value T ^* _p and the transmission ratio command value I ^* _p, which are operation amounts, are saturated, the feedback control amount T ^* _p-fb of the transmission input torque command value and the transmission ratio command value Only the feedback compensator having an integral characteristic when calculating and setting the feedback control component I ^* _p-fb is stopped. As described above, the feedback compensator having an integral characteristic has a characteristic of accumulating the difference between the target value and the actual value. Therefore, when the manipulated variable is saturated, the feedback compensator having the integral characteristic. Is stopped, the difference between the target value and the actual value is not accumulated, and the overshoot of the controlled variable is suppressed and prevented when the manipulated variable is no longer saturated.

Further, the stop condition of the feedback compensator having the integral characteristic is to the transmission input torque command value T ^* _p is the upper limit value, acceleration difference △ alpha _w is positive, i.e. the further acceleration and when There is required, or to the transmission input torque command value T ^* _p is the lower limit, acceleration difference △ alpha _w is negative value, when the words further deceleration is required, the acceleration operation of only the elements y ₁₁ for calculating a transmission input torque command value T ^* _p from the difference △ alpha _w is stopped. Further, although the transmission input torque command value T ^* _p is the upper limit value, and when a positive transmission input shaft rotational speed difference △ omega _p is, i.e. further increasing speed is required, or transmission When the input torque command value T ^* _p is the lower limit value but the transmission input shaft rotational speed difference Δω _p is a negative value, that is, when further deceleration is required, the transmission input shaft rotational speed difference Δ operation of only the elements y ₁₂ for calculating a transmission input torque command value T ^* _p from omega _p is stopped. Further, when the gear ratio command value I ^* _p is the upper limit value, the acceleration / deceleration difference Δα _w is greater than or equal to the predetermined value Δα _w1, that is, when further acceleration is required, or the gear ratio command value for I ^* _p is the lower limit, acceleration difference △ alpha _w is less than or equal to a predetermined value △ alpha _w2, i.e. when the further deceleration is required, the speed change ratio command value from the acceleration difference △ alpha _w The calculation of only element y ₂₁ for calculating I ^* _p is stopped. Further, although the gear ratio command value I ^* _p is the upper limit value, and when a positive transmission input shaft rotational speed difference △ omega _p is, i.e. further increasing speed is required, or speed ratio command value When I ^* _p is the lower limit value but the transmission input shaft rotational speed difference Δω _p is a negative value, that is, when further deceleration is required, a shift is made from the transmission input shaft rotational speed difference Δω _p. The calculation of only the element y ₂₂ for calculating the ratio command value I ^* _p is stopped. Therefore, the calculation of the element in which the operation amount is not saturated is continued, and more robust feedback control can be performed correspondingly.

  Further, as described above, when the manipulated variable is saturated, the calculation of only the feedback compensator having the integral characteristic is stopped, and the calculation of the other feedback compensators is continued. As described above, the feedback compensator based on the modern control theory is separated and the integral characteristic is removed. Conversely, the feedback compensator also has a differential characteristic, that is, a phase advance characteristic. Therefore, stopping the entire feedback compensator may cause the feedback compensator having this differential characteristic to react sensitively when restarting the feedback compensator operation. In the present embodiment, by continuing the operation of the feedback compensator excluding the integral characteristic in this way, it is possible to suppress and prevent the sensitive response of the control amount.

  Furthermore, in the present invention, when it is determined that the measurement accuracy of the wheel speed can be ensured (vehicle speed ≧ predetermined value V1, that is, when a predetermined number or more of sensor pulse edges are detected, step S4.5), the bandpass filter processing is performed. When wheel acceleration estimation calculation is executed (step S5) and it is determined that the output of the bandpass filter process has converged to a true value (vehicle speed ≧ predetermined value V1 continues for a set time T1 or more, step S9.5), In addition to the feedforward compensator, the feedback compensator is also operated to perform control in the original form of two-degree-of-freedom control. If not (the above condition is not satisfied), only the feedforward compensator calculation is performed. By adopting such control contents, acceleration feedback control can be started smoothly without causing overshoot or vibration even when starting acceleration is performed from a vehicle stop state or a state close thereto. The time T1 is set in consideration of the estimated delay due to the bandpass filter process.

  In this way, when starting and accelerating from a vehicle stop state or a state close thereto, acceleration control is performed only with a feedforward compensator that is excellent in responsiveness to the target value, and then the measurement accuracy of wheel acceleration is obtained. As soon as the acceleration control is performed by the feedforward compensator + feedback compensator (2-degree-of-freedom control), disturbance resistance (feedback effect for suppressing the influence of various disturbances) is ensured in addition to quick response. That is, it is possible to improve the drivability, which is the original aim, by improving the followability of the actual acceleration to the target acceleration even during start acceleration.

  This effect will be described with reference to FIG. 11. In the conventional control in the upper stage, since control is performed using a feedforward compensator and a feedback compensator from the start, overshoot occurs in the estimated value for the acceleration target value, When the vehicle starts to accelerate, the vehicle becomes vibrated and the drivability deteriorates.

  On the other hand, in the control of the present invention shown in the middle stage, first, the control is performed by the feedforward compensator until the vehicle speed reaches the predetermined value V1 to improve the responsiveness to the acceleration target value. Next, when the vehicle speed reaches a predetermined value and the state continues for the set time T1 or longer, control by the feedforward compensator and the feedback compensator can prevent deterioration of drivability as described above.

  Further, a value obtained by multiplying the output value of the feedback compensator that has started the calculation by a coefficient (gain) K that continuously changes from zero to 1 with the elapsed time from the start of the calculation is obtained as a feedforward compensator. The torque command value, which is the final manipulated variable, is calculated by adding to the output of. By such control, when shifting from the state of accelerating only by feedforward control to the state of accelerating by feedforward control + feedback control, the transition can be made more smoothly without causing an acceleration step. For example, when starting acceleration is performed on an uphill road, the output of the feedback compensator that has started the calculation outputs a value commensurate with the gradient resistance in order to suppress the influence of the uphill. However, if this is added directly to the output of the feedforward compensator, a step may occur in the torque command value or actual acceleration. Therefore, by gradually adding the output of the feedback compensator to the output of the feedforward compensator, the start acceleration can be realized smoothly without a step. The effect of this control content will be explained using the lower part of FIG. 11. When the set time T1 elapses after reaching the vehicle speed V1, the feedforward control is switched to the feedforward control + feedback control. As shown in the figure, the driver feels a difference in acceleration due to the deviation from the target value, and the drivability deteriorates. For this reason, the step of acceleration can be suppressed by switching to feedforward control + feedback control and correcting the output value of the feedback compensator with the gain K to obtain the torque command value.

  The present invention is a driving force control device that improves drivability when a vehicle starts, and is useful for a vehicle including a continuously variable transmission.

It is a schematic block diagram of the acceleration / deceleration feedback control apparatus which shows one Embodiment of the feedback control apparatus of this invention. FIG. 2 is a system configuration diagram of the acceleration / deceleration feedback control device of FIG. 1. It is a block diagram which shows the plant model of FIG. It is a block diagram of the feedforward compensator of FIG. It is an engine driving | operation restraint map. It is a flowchart of the arithmetic processing performed with the feedback compensator and feedforward compensator of FIG. It is a control map used for the arithmetic processing of FIG. It is a control map used for the arithmetic processing of FIG. It is a control map used for the arithmetic processing of FIG. It is a flowchart of the minor program performed by the arithmetic processing of FIG. It is a figure which shows the effect of the control in a feedforward compensator + feedback compensator. It is a figure explaining the subject of this invention.

Explanation of symbols

1 is engine 2 is continuously variable transmission 3 is torque converter 4 is wheel 5 is transmission controller 6 is acceleration / deceleration controller 7 is engine torque controller 11 is throttle actuator 12 is throttle valve 13 is primary pulley 14 is secondary pulley 16 is locked Up clutch 31 is feedforward compensator 33 is feedback compensator

Claims (3)

Target acceleration / deceleration setting means for setting the target acceleration / deceleration of the vehicle based on the driving state of the vehicle;
A feedforward compensator that outputs a feedforward manipulated variable from the target acceleration / deceleration to the vehicle;
A feedback compensator that outputs a feedback operation amount to the vehicle from a difference between the estimated acceleration / deceleration of the vehicle estimated from the wheel speed of the vehicle and the target acceleration / deceleration;
An operation amount setting means for setting an operation amount to the vehicle based on the feedforward operation amount and the feedback operation amount;
In a vehicle driving force control device comprising:
When the vehicle wheel speed does not exceed a predetermined speed, the setting of the operation amount to the vehicle based on the feedback operation amount is prohibited, and the feedback operation to set the operation amount to the vehicle only by the feedforward operation amount A vehicle driving force control apparatus comprising a quantity prohibiting means.   The feedback operation amount prohibiting means sets an operation amount to the vehicle only by the feedforward operation amount until a first predetermined time elapses after the wheel speed of the vehicle exceeds a predetermined speed. Item 2. The vehicle driving force control device according to Item 1. A feedback manipulated variable correcting means for correcting the feedback manipulated variable so as to converge to the feedback manipulated variable step by step and outputting the corrected feedback manipulated variable;
The second predetermined time after the prohibition of setting of the operation amount to the vehicle based on the feedback operation amount by the feedback operation amount prohibiting means is released based on the corrected feedback operation amount and the feedforward operation amount. The vehicle driving force control device according to claim 2, wherein an operation amount is set.

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