JP2018013135A - Controller of automatic transmission - Google Patents

Abstract

A control device for an automatic transmission capable of suppressing an overshoot or undershoot of an input shaft rotation speed with respect to a synchronous rotation speed after a shift is provided. An ECU is configured to calculate a target input shaft acceleration, and to control the target input shaft acceleration so that the target input shaft acceleration is realized and to proceed with a gear shift. When the shaft rotation speed is controlled, a realizable input shaft acceleration is calculated as a guard value, and the target input shaft acceleration is corrected using the guard value. [Selection] Figure 6

Description

  The present invention relates to a control device for an automatic transmission.

  2. Description of the Related Art Conventionally, there is known an automatic transmission control device that controls an automatic transmission that establishes a plurality of shift stages by selectively engaging a plurality of friction engagement elements (see, for example, Patent Document 1).

  The control device for an automatic transmission described in Patent Document 1 is configured to perform shift control using a shift model that determines a control operation amount for realizing a shift target value. In the shift control based on this shift model, the input shaft acceleration and the output shaft torque, which are shift target values, are calculated, and the torque sharing ratio is calculated. Then, based on the shift target value and the torque sharing ratio, the input shaft torque, the engagement side clutch torque, and the release side clutch torque as the control operation amount are calculated. The control device for the automatic transmission controls the driving force source so as to obtain the calculated input shaft torque, and controls the hydraulic control device so as to obtain the calculated engagement side clutch torque and the calculated release side clutch torque. .

Japanese Patent Laid-Open No. 2006-035301

  However, the above-described conventional automatic transmission control apparatus may calculate a control operation amount that cannot be realized by the target input shaft acceleration, which is a shift target value. In such a case, the guard process is performed on the control operation amount. However, the target input axis acceleration and the actual input axis acceleration continue to deviate, and the integral term of the feedback control continues to accumulate. For this reason, feedback control does not work properly, and the input shaft rotation speed may overshoot or undershoot the synchronized rotation speed after the shift.

  The present invention has been made to solve the above-described problems, and an object of the present invention is to prevent the input shaft rotational speed from overshooting or undershooting the synchronous rotational speed after shifting. It is to provide a control device for a possible automatic transmission.

  The control apparatus for an automatic transmission according to the present invention is applied to an automatic transmission that establishes a plurality of shift stages by selectively engaging a plurality of friction engagement elements. The control device for the automatic transmission is configured to calculate the target input shaft acceleration, and to control the target input shaft acceleration so that the target input shaft acceleration is realized, and to proceed with the shift. When controlling the input shaft rotation speed, a realizable input shaft acceleration is calculated as a guard value, and the target input shaft acceleration is corrected using the guard value.

  With this configuration, it is possible to suppress the target input axis acceleration from continuing to deviate from the actual input axis acceleration, and thus it is possible to suppress the accumulation of the integral term of the feedback control. Thereby, since feedback control can work appropriately, it can suppress that an input shaft rotation speed overshoots or undershoots to a synchronous rotation speed after shifting.

  According to the control apparatus for an automatic transmission of the present invention, it is possible to suppress the overshoot or undershoot of the input shaft rotation speed with respect to the synchronous rotation speed after the shift.

It is the figure which showed schematic structure of the vehicle provided with ECU by one Embodiment of this invention. FIG. 2 is a skeleton diagram illustrating configurations of a torque converter and an automatic transmission of FIG. 1. 3 is an engagement table showing engagement states of a first clutch to a fourth clutch, a first brake, and a second brake for each gear position in the automatic transmission of FIG. 2. It is the block diagram which showed ECU of FIG. It is the timing chart which showed an example of the power-on downshift by a comparative example. 5 is a timing chart illustrating an example of a power-on downshift according to the present embodiment. It is a flowchart for demonstrating the operation example of the shift control by this embodiment. It is a flowchart for demonstrating an example of the correction process of the target input axis acceleration in step ST3 of FIG.

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

  First, with reference to FIGS. 1-4, the vehicle 100 provided with ECU5 by one Embodiment of this invention is demonstrated.

  As shown in FIG. 1, the vehicle 100 includes an engine 1, a torque converter 2, an automatic transmission 3, a hydraulic control device 4, and an ECU 5. The vehicle 100 is, for example, an FF (front engine / front drive) system, and the output of the engine 1 is transmitted to the differential device 6 via the torque converter 2 and the automatic transmission 3, and left and right drive wheels (front wheels) 7. To be distributed.

-Engine-
The engine (internal combustion engine) 1 is a driving force source for traveling, for example, a multi-cylinder gasoline engine. The engine 1 is configured such that its operating state can be controlled by the throttle valve opening (intake air amount), fuel injection amount, ignition timing, and the like.

-Torque converter-
As shown in FIG. 2, the torque converter 2 includes a pump impeller 21 connected to a crankshaft 1a that is an output shaft of the engine 1, a turbine runner 22 connected to the automatic transmission 3, and a stator having a torque amplification function. 23, and a lockup clutch 24 for directly connecting the engine 1 and the automatic transmission 3 to each other. In FIG. 2, the lower half is omitted and only the upper half is schematically shown with respect to the rotation center axes of the torque converter 2 and the automatic transmission 3.

-Automatic transmission-
The automatic transmission 3 is provided in a power transmission path between the engine 1 and the drive wheels 7, and is configured to shift the rotation of the input shaft 3a and output it to the output shaft 3b. In the automatic transmission 3, the input shaft 3 a is connected to the turbine runner 22 of the torque converter 2, and the output shaft 3 b is connected to the drive wheels 7 via the differential device 6 and the like.

  The automatic transmission 3 includes a first transmission unit (front planetary) 31 mainly composed of a first planetary gear unit 31a, a second planetary gear unit 32a, and a second planetary gear unit 32b. A transmission unit (rear planetary) 32, a first clutch C1 to a fourth clutch C4, a first brake B1, a second brake B2, and the like are configured.

  The first planetary gear device 31a constituting the first transmission unit 31 is a double pinion type planetary gear mechanism, and supports a sun gear S1, a plurality of pairs of pinion gears P1 meshing with each other, and the pinion gears P1 so as to be able to rotate and revolve. Planetary carrier CA1 and ring gear R1 meshing with sun gear S1 via pinion gear P1.

  The planetary carrier CA1 is coupled to the input shaft 3a and rotates integrally with the input shaft 3a. The sun gear S1 is fixed to the transmission case 30 and cannot rotate. The ring gear R1 functions as an intermediate output member, is decelerated with respect to the input shaft 3a, and transmits the decelerated rotation to the second transmission unit 32.

  The second planetary gear unit 32a constituting the second transmission unit 32 is a single pinion type planetary gear mechanism, which is a sun gear S2, a pinion gear P2, and a planetary carrier RCA that supports the pinion gear P2 so as to be capable of rotating and revolving. And a ring gear RR that meshes with the sun gear S2 via the pinion gear P2.

  The third planetary gear device 32b constituting the second transmission unit 32 is a double pinion type planetary gear mechanism, and includes a sun gear S3, a plurality of pairs of pinion gears P2 and P3 meshing with each other, and the pinion gears P2 and P3. A planetary carrier RCA that supports rotation and revolution is provided, and a ring gear RR that meshes with the sun gear S3 via pinion gears P2 and P3. The planetary carrier RCA and the ring gear RR are shared by the second planetary gear device 32a and the third planetary gear device 32b.

  The sun gear S2 is selectively connected to the transmission case 30 by the first brake B1. The sun gear S2 is selectively connected to the ring gear R1 via the third clutch C3. Further, the sun gear S2 is selectively coupled to the planetary carrier CA1 via the fourth clutch C4. Sun gear S3 is selectively coupled to ring gear R1 via first clutch C1. The planetary carrier RCA is selectively coupled to the transmission case 30 by the second brake B2. Further, the planetary carrier RCA is selectively coupled to the input shaft 3a via the second clutch C2. The ring gear RR is connected to the output shaft 3b and rotates integrally with the output shaft 3b.

  The first clutch C1 to the fourth clutch C4, the first brake B1, and the second brake B2 are all friction engagement elements that are frictionally engaged by a hydraulic actuator, and are controlled by the hydraulic control device 4 and the ECU 5.

  FIG. 3 is an engagement table showing an engaged state or a released state of the first clutch C1 to the fourth clutch C4, the first brake B1, and the second brake B2 for each shift speed (gear speed). In the engagement table of FIG. 3, a circle indicates an “engaged state”, and a blank indicates a “released state”.

  As shown in FIG. 3, in the automatic transmission 3 of this example, the first clutch C1 and the second brake B2 are engaged, so that the gear ratio (the rotational speed of the input shaft 3a / the rotational speed of the output shaft 3b). The first shift speed (1st) with the largest value is established. The second gear (2nd) is established by engaging the first clutch C1 and the first brake B1.

  The third shift stage (3rd) is established by engaging the first clutch C1 and the third clutch C3, and the fourth shift stage (4th) by engaging the first clutch C1 and the fourth clutch C4. Is established. The fifth gear (5th) is established by engaging the first clutch C1 and the second clutch C2, and the sixth gear (6th) by engaging the second clutch C2 and the fourth clutch C4. Is established. The seventh shift stage (7th) is established by engaging the second clutch C2 and the third clutch C3, and the eighth shift stage (8th) by engaging the second clutch C2 and the first brake B1. Is established. The reverse speed (Rev) is established when the third clutch C3 and the second brake B2 are engaged.

-Hydraulic control device-
The hydraulic control device 4 is provided to control the state (engaged state or released state) of the friction engagement element of the automatic transmission 3. The hydraulic control device 4 also has a function of controlling the lockup clutch 24 of the torque converter 2.

-ECU-
The ECU 5 is configured to perform operation control of the engine 1 and shift control of the automatic transmission 3. Specifically, as shown in FIG. 4, the ECU 5 includes a CPU 51, a ROM 52, a RAM 53, a backup RAM 54, an input interface 55, and an output interface 56. The ECU 5 is an example of the “automatic transmission control device” in the present invention.

  The CPU 51 executes arithmetic processing based on various control programs and maps stored in the ROM 52. The ROM 52 stores various control programs, maps that are referred to when the various control programs are executed, and the like. The RAM 53 is a memory that temporarily stores a calculation result by the CPU 51, a detection result of each sensor, and the like. The backup RAM 54 is a non-volatile memory that stores data to be stored when the ignition is turned off.

  A crank position sensor 81, an input shaft rotational speed sensor 82, an output shaft rotational speed sensor 83, an accelerator opening sensor 84, a throttle opening sensor 85, and the like are connected to the input interface 55.

  The crank position sensor 81 is provided for calculating the rotational speed (angular speed) of the engine 1. The input shaft rotational speed sensor 82 is provided for calculating the rotational speed (turbine rotational speed) of the input shaft 3 a of the automatic transmission 3. The output shaft rotation speed sensor 83 is provided for calculating the rotation speed of the output shaft 3 b of the automatic transmission 3. The accelerator opening sensor 84 is provided to detect an accelerator opening that is an accelerator pedal depression amount (operation amount). The throttle opening sensor 85 is provided for detecting the throttle opening of the throttle valve.

  To the output interface 56, an injector 91, an igniter 92, a throttle motor 93, the hydraulic control device 4, and the like are connected. The injector 91 is a fuel injection valve and can adjust the fuel injection amount. The igniter 92 is provided for adjusting the ignition timing by the ignition plug. The throttle motor 93 is provided to adjust the throttle opening of the throttle valve.

  The ECU 5 is configured to be able to control the operating state of the engine 1 by controlling the throttle opening, the fuel injection amount, the ignition timing, and the like based on the detection result of each sensor. Further, the ECU 5 is configured to be able to execute shift control of the automatic transmission 3 and control of the lock-up clutch 24 of the torque converter 2 by controlling the hydraulic control device 4.

  In the shift control by the ECU 5, for example, the target shift stage is set based on a shift map using the vehicle speed and the accelerator opening as parameters, and the hydraulic control device 4 is controlled so that the actual shift stage becomes the target shift stage. That is, the ECU 5 performs a shift determination based on the shift map, and executes the shift control so that the target shift stage is obtained when it is determined that the shift should be executed.

  In this speed change control, switching to a shift stage established by releasing one friction engagement element and engaging one friction engagement element is permitted, and release of two friction engagement elements and two friction engagement elements are permitted. Switching to a gear position that requires engagement of the engagement element is prohibited. Further, it is possible to switch to a shift stage that is two or more steps away from the current shift stage.

-Shift control of automatic transmission-
Here, as a general shift control, for example, whether or not a shift shock or a shift time is appropriate is evaluated with an actual vehicle, and each frictional factor at the time of shift is determined based on a control map predetermined by adaptation. There is a technique for determining the torque capacity (or hydraulic pressure command value) of the combination element (the clutch and the brake) and executing a shift. In the method using this control map, it is necessary to create a large number of control maps in accordance with a combination of a shift pattern such as a power-on downshift or a power-off upshift and a shift stage before and after the shift. For this reason, the greater the number of shift stages of the automatic transmission, the more labor is required for the adaptation work.

  Therefore, in the present embodiment, as a shift control, a method of executing a shift using a shift model that determines a control operation amount for realizing a shift target value is employed instead of the method using the control map. The shift target value is a target value of an element (for example, shift time, driving force, etc.) that determines a change mode to be realized at the time of shifting. The control operation amount is a required value of an element (engine torque, clutch torque, etc.) operated with respect to a control target.

  Hereinafter, the shift control using the shift model will be described. The equation of motion during the shift is expressed by the following equations (1) and (2).

  These expressions (1) and (2) are derived from the equations of motion for the mutually connected rotating elements constituting the automatic transmission 3 and the relational expressions in the planetary gear device constituting the automatic transmission 3. It is a thing. The equation of motion for each rotating element is a torque expressed by the product of the inertia in each rotating element and the rate of change in rotational speed with time, among the three members of the planetary gear unit and the members on both sides of the friction engagement element. The equation of motion is defined by the torque acting on the members involved in each rotating element. Further, the relational expression in the planetary gear unit is a relational expression that defines the relationship between the torque and the rate of change in the rotational speed time in the three members of the planetary gear unit using the gear ratio of the planetary gear unit.

  In the equations (1) and (2), dωt / dt is a time derivative of the turbine rotational speed ωt (that is, the transmission input shaft rotational speed ωi), that is, a time variation rate, and a speed variation of the rotating member on the input shaft 3a side. This represents acceleration (hereinafter referred to as input axis acceleration) dωin / dt of the input shaft 3a as a quantity. dωo / dt is the time change rate of the transmission output shaft rotational speed ωo and represents the output shaft acceleration. Tt represents turbine torque that is torque on the input shaft 3a as torque on the rotating member on the input shaft 3a side, that is, input shaft torque Tin. The turbine torque Tt agrees with the engine torque Te (= Tt / t) when the torque ratio t of the torque converter 2 is taken into consideration. To represents the output shaft torque that is the torque on the output shaft 3b as the torque on the rotating member on the output shaft 3b side. Tcapl is a torque capacity (hereinafter referred to as an engagement-side clutch torque) of a friction engagement element that performs an engagement operation at the time of shifting. Tcdrn is the torque capacity of the friction engagement element that performs the releasing operation at the time of shifting (hereinafter referred to as the release side clutch torque). a1, a2, b1, b2, c1, c2, d1, and d2 are constants when the equations (1) and (2) are derived, and the inertia and the planetary gears in each of the rotating elements It is a coefficient determined by design from the gear ratio of the device. The specific numerical value of this constant varies depending on, for example, the type of shift (for example, a shift pattern or a combination of shift stages before and after the shift). Therefore, although the above-mentioned equation of motion is one predetermined equation, the equation of motion corresponding to each type of shift, which is a constant different for each type of shift, is used for the shift of the automatic transmission 3.

  The equations (1) and (2) are gear train motion equations of the automatic transmission 3 in which the relationship between the shift target value and the control operation amount is formulated. The shift target value can express each target value of the shift time and the driving force and can be handled on the gear train motion equation. In the present embodiment, the input shaft acceleration dωt / dt is used as an example of a physical quantity that can express the shift time. Further, the output shaft torque To of the transmission 3 is used as an example of a physical quantity that can express the driving force. That is, in this embodiment, the shift target value is set with two values of the input shaft acceleration dωt / dt and the output shaft torque To.

  On the other hand, in the present embodiment, the control operation amount for establishing the shift target value is represented by three values of the turbine torque Tt (the engine torque Te is also agreed), the engagement side clutch torque Tcapl, and the release side clutch torque Tcdrn. It is set. Then, since there are three control operation amounts for the equation of motion being composed of the above two formulas (1) and (2), the control operation amount for establishing the two shift target values is uniquely solved. It is not possible. The output shaft acceleration dωo / dt in each equation is calculated from the transmission output shaft rotational speed ωo, which is a detection value of the output shaft rotational speed sensor 83.

  Therefore, a study was made to add a constraint condition to the equations of motion of the equations (1) and (2) to uniquely solve the control operation amount. In this embodiment, it is suitable for expressing and controlling the transfer of torque during a shift, and as a restraint condition that can correspond to any shift pattern, it is engaged with a disengagement clutch. The torque sharing rate of the transmission torque that is handled by the side clutch is used. That is, the torque sharing rate of the transmission torque that can incorporate the torque transfer during the shift into the equation of motion and uniquely solve the control operation amount is set as the constraint condition. The torque sharing ratio is the total transmission torque (total transmission torque) that must be handled by the disengagement side clutch and the engagement side clutch when shifting the automatic transmission 3, for example, torque on the input shaft 3a (total on the input shaft) (Transmission torque) is the ratio of the transmission torque shared by both friction engagement elements to the total transmission torque on the input shaft. In this embodiment, the torque sharing rate of the engagement side clutch is set to “xapl”, the torque sharing rate of the release side clutch is set to “xdrn”, and each torque sharing rate reflects the transfer of torque during shifting. Using the torque sharing ratio x (for example, 0 ≦ x ≦ 1) that changes in time series, the following equations (3) and (4) are defined.

xapl = x (3)
xdrn = 1-x (4)
The relational expression between the engagement-side clutch torque Tcapl and the disengagement-side clutch torque Tcdrn is based on “Tcapl” and “Tcdrn” replaced with the torque on the input shaft 3a, and the expressions (3) and (4). , “X” (= xapl) and “1-x” (= xdrn). Then, from the equation (1), the equation (2), and the relational expression between “Tcapl” and “Tcdrn”, the turbine operation torque Tt, the engagement side clutch torque Tcapl, and the disengagement side, which are control operation amounts, are obtained. A relational expression for calculating the clutch torque Tcdrn is derived. The turbine torque Tt (the engine torque Te is also agreed) is expressed by a relational expression using “x” (= xapl), “1-x” (= xdrn), input shaft acceleration dωt / dt, output shaft torque To, and the like. It is expressed as Similarly, the engagement side clutch torque Tcapl is represented by a relational expression using “x” (= xapl), the input shaft acceleration dωt / dt, the output shaft torque To, and the like. Similarly, the release side clutch torque Tcdrn is represented by a relational expression using “1-x” (= xdrn), input shaft acceleration dωt / dt, output shaft torque To, and the like.

  That is, the speed change model of the present embodiment has a relation between the equation of motion (the above formulas (1) and (2)) of the automatic transmission 3 including the shift target value and the control operation amount, and the torque sharing rate ( The control operation amount is calculated based on the shift target value using the equations (3) and (4)). As described above, in this embodiment, the shift of the automatic transmission 3 is executed using the shift model by adding the constraint condition set by the torque sharing ratio x to the equations (1) and (2). . Therefore, even if there are three control operation amounts for two shift target values, the three control operation amounts can be appropriately determined using the shift model. This shift model is one predetermined model. However, as described above, since the gear train equation of motion, which is a constant different for each type of shift (for example, a combination of shift patterns and shift stages before and after the shift), is used, For the speed change of the transmission 3, a speed change model corresponding to each type of speed change is used.

-Example of operation during shifting-
Next, an example of a power-on downshift will be described with reference to FIGS. In the following, the power-on downshift according to the conventional comparative example will be described with reference to FIG. 5, and then the power-on downshift according to the present embodiment will be described with reference to FIG.

  In the comparative example of FIG. 5, the downshift determination is made based on the shift map by depressing the accelerator pedal from the state where the eighth shift stage is established, and the fifth shift stage is set as the target shift stage. At time t1, shift control during a power-on downshift for switching from the eighth shift stage to the fifth shift stage is started.

  This shift control is based on a shift model, and the input shaft acceleration and output shaft torque, which are shift target values, are calculated, and the torque sharing ratio is calculated. The target input axis acceleration is calculated based on, for example, an input axis acceleration change map that defines a mode for changing the input axis acceleration. The input shaft acceleration change map is determined in advance so that the turbine rotation speed can be changed during the inertia phase while simultaneously suppressing the shift shock and shortening the shift time. The target output shaft torque is calculated based on, for example, an output shaft torque change map that defines a mode for changing the output shaft torque. Further, the torque sharing rate is calculated based on, for example, a map that defines a mode for changing the torque sharing rate. These maps are determined for each type of shift (for example, a shift pattern or a combination of shift stages before and after a shift).

  Then, based on the shift target value and the torque sharing rate, the turbine torque, the engagement side clutch torque, and the release side clutch torque as the control operation amount are calculated. The ECU controls the engine so as to obtain the calculated turbine torque, and controls the hydraulic control device so as to obtain the calculated engagement side clutch torque and the calculated release side clutch torque.

  First, the hydraulic pressure command for the first brake B1, which is the disengagement side frictional engagement element, starts to decrease from the time point t1 when the shift is started. Further, after a first fill pressure is output as a hydraulic pressure command for the first clutch C1, which is the friction engagement element on the engagement side, a predetermined standby pressure (a so-called pack packing pressure immediately before the friction engagement element starts to be engaged). Is output.

  Next, at the time t2, the target input shaft acceleration rises, and the target rotational speed of the input shaft increases from the synchronous rotational speed of the eighth shift stage. Then, the first brake B1 is released, and the actual rotational speed of the input shaft increases from the synchronous rotational speed of the eighth shift stage. That is, the inertia phase is started. In the inertia phase during the power-on downshift, the rotation of the input shaft is controlled mainly by the disengagement friction engagement element. Further, the synchronous rotational speed of each gear stage is calculated based on the gear ratio of each gear stage and the rotational speed of the output shaft.

  Here, there is a case where a control operation amount (release clutch torque) that cannot be realized depending on the target input shaft acceleration is calculated. In such a case, the release side clutch torque is guarded at a realizable value, and it becomes difficult to bring the actual input axis acceleration close to the target input axis acceleration. For this reason, as shown in FIG. 5, the actual input axis acceleration may be maintained at a lower level than the target input axis acceleration. In FIG. 5, the input axis acceleration that is not realized is indicated by hatching. If the target input axis acceleration and the actual input axis acceleration continue to deviate, the integral term of the feedback control continues to accumulate. Therefore, after that, when the actual input shaft acceleration increases, the integral term is accumulated, so that it takes time to reflect, the first brake B1 cannot be appropriately controlled, and the input shaft rotation speed becomes the first speed. Overshoot with respect to the synchronous rotation speed of the fifth gear.

  At time t3, the hydraulic pressure command for the first brake B1 is lowered and completely released, and the first clutch C1 is forcibly engaged by a backup sweep (an increase in the hydraulic pressure command for the first clutch C1). There was a shock.

  Therefore, in the present embodiment, the ECU 5 calculates a realizable input shaft acceleration as a guard value in the case of a power-on downshift and a power-off upshift, and corrects the target input shaft acceleration using the guard value. It is configured. Specifically, the ECU 5 calculates, as a guard value, the limit of the input shaft acceleration that can be realized by the disengagement side frictional engagement element that controls the rotation of the input shaft 3a at the time of shifting, and uses the guard value as a target input shaft. It is configured to correct the acceleration. As a result, it is possible to prevent the target input axis acceleration from continuing to deviate from the actual input axis acceleration. Here, the guard value is calculated using Equation (5), which is a gear train equation in the inertia phase.

Iin × dωin / dt = α · Tin + β · Tcdrn (5)
In equation (5), Iin is the input shaft inertia, and dωin / dt is the input shaft acceleration (angular acceleration of the input shaft rotation). Further, Tin is the input shaft torque, Tcdrn is the disengagement side clutch torque, and α and β are predetermined coefficients.

  In this equation (5), when the release side clutch torque Tcdrn is zero, the input shaft acceleration dωin / dt is expressed by the following equation (6).

dωin / dt = (α / Iin) × Tin (6)
Then, α / Iin is calculated in advance as a coefficient, and the input axis acceleration dωin / dt used as a guard value is calculated by multiplying the coefficient α / Iin by the input shaft torque Tin at the time of shifting. The equation (6) for calculating the guard value can be switched for each engagement holding element during shifting. In addition, when the input shaft torque Tin is a positive value and the power-on downshift, the guard value is a positive value, and the upper limit guard process is performed using the guard value, and the input shaft torque Tin is a negative value. In the case of the power-off upshift, the guard value is a negative value, and the lower limit guard process is performed using the guard value.

  In the present embodiment of FIG. 6, the downshift determination is made based on the shift map when the accelerator pedal is depressed from the state where the eighth shift stage is established, and the fifth shift stage is set as the target shift stage. . At time t11, shift control at the time of power-on downshift for switching from the eighth shift stage to the fifth shift stage is started.

  This shift control is based on a shift model, and the input shaft acceleration and output shaft torque, which are shift target values, are calculated, and the torque sharing ratio is calculated. Note that the target input shaft acceleration, the target output shaft torque, and the torque sharing rate are calculated based on a map, for example, as in the comparative example.

  Here, in this embodiment, a guard value is calculated based on the above formula (6), and the target input axis acceleration is corrected using the guard value. As a result, the target input shaft acceleration is corrected within a range that can be realized by the first brake B1 that is the disengagement side frictional engagement element that controls the rotation of the input shaft 3a. The guard value calculation formula for the second clutch C2, which is an engagement holding element at the time of shifting, is used. In the example of FIG. 6, since it is a power-on downshift, the upper limit guard process is performed.

  Therefore, when the target input axis acceleration before correction calculated based on the input axis acceleration change map (hereinafter also simply referred to as “map”) is larger than the guard value, the guard value becomes the corrected target value. . If the target input axis acceleration before correction calculated based on the map is smaller than the guard value, the target value before correction becomes the target value after correction as it is. That is, the target input shaft acceleration is corrected so as not to be higher than the guard value that is the limit that can be realized by the first brake B1.

  Then, based on the shift target value including the corrected target input shaft acceleration and the torque sharing rate, the turbine torque, the engagement side clutch torque, and the release side clutch torque as the control operation amount are calculated. The ECU 5 controls the engine 1 so as to obtain the calculated turbine torque, and controls the hydraulic control device 4 so as to obtain the calculated engagement side clutch torque and the calculated release side clutch torque.

  First, the hydraulic pressure command for the first brake B1, which is the disengagement side frictional engagement element, starts to decrease from the time t11 when the shift is started. In addition, a predetermined standby pressure is output after the first fill pressure is output as a hydraulic pressure command for the first clutch C1, which is the friction engagement element on the engagement side.

  Next, at time t12, the target input shaft acceleration rises, and the target rotational speed of the input shaft 3a increases from the synchronous rotational speed of the eighth shift stage. Then, the first brake B1 is released, and the actual rotational speed of the input shaft 3a increases from the synchronous rotational speed of the eighth shift stage. That is, the inertia phase is started. In the inertia phase during the power-on downshift, the rotation of the input shaft 3a is controlled mainly by the disengagement side frictional engagement element. Further, the synchronous rotational speed of each gear stage is calculated based on the gear ratio of each gear stage and the rotational speed of the output shaft 3b.

  Here, the target input axis acceleration is corrected using the guard value as described above. For this reason, as shown in FIG. 6, if the target input shaft acceleration before correction calculated based on the map cannot be realized by the first brake B1, it can be realized by the first brake B1 by the upper limit guard process. It is corrected to the target input axis acceleration. For this reason, rotation of the input shaft 3a can be controlled by the first brake B1 so that the actual input shaft acceleration becomes the corrected target input shaft acceleration. Therefore, it is possible to suppress the actual input axis acceleration from continuing to deviate from the corrected target input axis acceleration.

  After that, when the actual input shaft acceleration increases, the feedback control works appropriately, and the first brake B1 prevents the input shaft rotational speed from overshooting the fifth rotational speed synchronous rotational speed. can do. That is, since the integral term does not accumulate excessively, the feedback control works appropriately, and the hydraulic pressure command for the first brake B1 can be raised in order to suppress the increase in the input shaft rotation speed.

  At time t13, the hydraulic pressure command for the first brake B1 is decreased toward zero, and the hydraulic pressure command for the first clutch C1 is increased. Thus, the first brake B1 is completely released, the first clutch C1 is engaged, and the fifth shift speed is established. The target input axis acceleration is zero. Further, unlike the comparative example in which the input shaft rotation speed overshoots and the first clutch C1 is forcibly engaged, it is possible to suppress a shock when the first clutch C1 is engaged.

[Operation example of shift control]
Next, an operation example of the shift control according to the present embodiment will be described with reference to FIGS. The following steps are executed by the ECU 5.

  First, in step ST1 of FIG. 7, it is determined whether or not a shift is being performed. That is, it is determined whether or not switching to the target gear stage has been executed. If it is determined that shifting is in progress, the process proceeds to step ST2. On the other hand, if it is determined that the speed is not being changed, the routine returns.

  Next, in step ST2, a shift target value and a torque sharing rate are calculated. Specifically, the target input shaft acceleration is calculated based on the input shaft acceleration change map, and the target output shaft torque is calculated based on the output shaft torque change map. Further, the torque sharing rate is calculated based on a map that defines a mode for changing the torque sharing rate.

  Next, in step ST3, target input axis acceleration correction processing is performed. In this correction process, it is determined in step ST11 of FIG. 8 whether the power-on downshift or the power-off upshift. And in the case of a power-on downshift or a power-off upshift, it moves to step ST12. On the other hand, in the case of a power-on upshift or a power-off downshift, the target input axis acceleration is not corrected and the process proceeds to the end.

  Next, in step ST12, the input shaft torque (turbine torque) is estimated. The input shaft torque is calculated based on, for example, the estimated engine torque calculated based on the throttle opening and the engine rotation speed, and the torque ratio of the torque converter 2.

  Next, in step ST13, a guard value is calculated. Specifically, the guard value is calculated based on the input shaft torque calculated in step ST12 and the above equation (6).

  In step ST14, it is determined whether or not the input shaft torque calculated in step ST12 is a positive value. If it is determined that the input shaft torque is a positive value, it is a power-on downshift, and the process proceeds to step ST15. On the other hand, if it is determined that the input shaft torque is a negative value, it is a power-off upshift, and the process proceeds to step ST16.

  In step ST15, an upper limit guard process is performed using the guard value for the target input axis acceleration calculated in step ST2. That is, when the target input axis acceleration calculated in step ST2 is larger than the guard value, the guard value becomes the corrected target value, and the target input axis acceleration calculated in step ST2 is smaller than the guard value. In this case, the target input axis acceleration calculated in step ST2 becomes the corrected target value as it is. Thereby, the correction process of the target input axis acceleration is completed, and the process proceeds to the end.

  In step ST16, a lower limit guard process is performed using the guard value for the target input axis acceleration calculated in step ST2. That is, when the target input axis acceleration calculated in step ST2 is smaller than the guard value, the guard value becomes the corrected target value, and the target input axis acceleration calculated in step ST2 is larger than the guard value. In this case, the target input axis acceleration calculated in step ST2 becomes the corrected target value as it is. Thereby, the correction process of the target input axis acceleration is completed, and the process proceeds to the end.

  Next, in step ST4 of FIG. 7, the control operation amount is calculated. Specifically, in the case of a power-on downshift or a power-off upshift, the turbine is based on the target input shaft acceleration corrected in step ST3 and the target output shaft torque and torque sharing ratio calculated in step ST2. Torque, engagement side clutch torque, and release side clutch torque are calculated. In the case of a power-on upshift or a poweroff downshift, the turbine torque, the engagement side clutch torque, and the release are based on the target input shaft acceleration, the target output shaft torque, and the torque sharing ratio calculated in step ST2. The side clutch torque is calculated.

  In step ST5, the engine 1 and the hydraulic control device 4 are controlled. Specifically, the hydraulic control device 4 controls the engine 1 so as to obtain the turbine torque calculated in step ST4 and obtains the engagement-side clutch torque and the disengagement-side clutch torque calculated in step ST4. To control. Then move on to return.

-Effect-
In the present embodiment, as described above, when the input shaft rotation speed is controlled by the disengagement friction engagement element at the time of shifting (in the case of power-on downshift or power-off upshift), the realizable input shaft acceleration is It is calculated as a guard value, and the target input axis acceleration is corrected using the guard value. With this configuration, it is possible to suppress the target input axis acceleration from continuing to deviate from the actual input axis acceleration, and thus it is possible to suppress the accumulation of the integral term of the feedback control. Thereby, since feedback control can work appropriately, it can suppress that an input shaft rotation speed overshoots or undershoots to a synchronous rotation speed after shifting.

-Other embodiments-
In addition, embodiment disclosed this time is an illustration in all the points, Comprising: It does not become a basis of limited interpretation. Therefore, the technical scope of the present invention is not interpreted only by the above-described embodiments, but is defined based on the description of the scope of claims. Further, the technical scope of the present invention includes all modifications within the meaning and scope equivalent to the scope of the claims.

  For example, in the present embodiment, an example in which the vehicle 100 is an FF has been described. However, the present invention is not limited thereto, and the vehicle may be an FR (front engine / rear drive) or may be a four-wheel drive. .

  In the present embodiment, the example in which the present invention is applied in the case of the power-on downshift is shown in FIG. 6, but the present invention is not limited to this, and the present invention can be applied in the case of the power-off upshift.

  Further, in the present embodiment, the example in which the engine 1 is a multi-cylinder gasoline engine is shown, but the present invention is not limited thereto, and the engine may be a diesel engine or the like.

  In this embodiment, ECU5 may be constituted by a plurality of ECUs.

  The present invention is applicable to an automatic transmission control device that controls an automatic transmission that establishes a plurality of shift stages by selectively engaging a plurality of friction engagement elements.

3 Automatic transmission 5 ECU (Control device for automatic transmission)
C1 first clutch (friction engagement element)
C2 Second clutch (friction engagement element)
C3 3rd clutch (friction engagement element)
C4 4th clutch (friction engagement element)
B1 First brake (friction engagement element)
B2 Second brake (friction engagement element)

Claims (1)

A control device for an automatic transmission that is applied to an automatic transmission that establishes a plurality of shift stages by selectively engaging a plurality of friction engagement elements,
The target input shaft acceleration is calculated, and the shift is advanced by controlling the target input shaft acceleration to be realized.
When the input shaft rotation speed is controlled by the frictional engagement element on the release side at the time of shifting, the realizable input shaft acceleration is calculated as a guard value, and the target input shaft acceleration is corrected using the guard value A control device for an automatic transmission, characterized in that

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