JP6551329B2 - Control device of automatic transmission - Google Patents

Description

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

  2. Description of the Related Art A control device for an automatic transmission is known which controls an automatic transmission that establishes a plurality of shift speeds 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 a power-on downshift that shifts from a current shift speed to a target shift speed via an intermediate shift speed. Specifically, the control device for the automatic transmission operates from the sixth speed established by engagement of the second clutch and the first brake to the third speed established by engagement of the first clutch and the third clutch. At the time of transition to the second gear stage, the fourth speed stage established by the engagement of the first clutch and the second clutch is passed. That is, since shifting from the sixth speed to the third speed requires shifting of four elements, shifting from the sixth speed to the fourth speed, which can be switched by two elements, and two elements. Shifting from the switchable fourth speed to the third speed is continuously performed. In such a 6-4-3 shift, the 6-4 shift and the 4-3 shift are individually performed in two steps by preparing for the 4-3 shift in advance during the execution of the 6-4 shift. Compared to the case, the shift time can be shortened.

JP 2002-310281 A

  Here, in the above-described conventional automatic transmission, when passing through the intermediate gear, the internal rotation of the rear planetary gear constituting the transmission mechanism is reversed. The internal rotation of the rear planetary is reversed when the rotation speed of the sun gear and the planetary carrier changes with the progress of the shift when the rotation speed of the ring gear connected to the output shaft is constant. At this time, if the torque capacity of the first brake, which is the disengagement friction engagement element at the time of shifting from the sixth speed to the fourth speed, is small, the torque applied to the output shaft becomes smaller than zero, and the output shaft The backlash is stuck on the opposite side. In this state, when the first clutch, which is a frictional engagement element on the engagement side, is engaged, torque in the positive direction is applied to the output shaft, which may cause a shock due to rattling.

  The present invention has been made to solve the above-described problems, and an object of the present invention is to perform a power-on downshift in which the current shift speed shifts to the target shift speed via the intermediate shift speed. It is an object of the present invention to provide a control device of an automatic transmission capable of suppressing a shock when passing through an intermediate gear.

  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 such that the shift progress during the power-on downshift is performed by controlling the frictional engagement element on the release side based on the target inertia torque. Further, the control device for the automatic transmission calculates a guard value based on the degree of shift progress to the intermediate gear during a power-on downshift where the current gear shifts to the target gear via the intermediate gear. The target inertia torque is corrected using the guard value.

  By this configuration, when approaching the intermediate gear, the torque capacity of the frictional engagement element on the release side can be increased by correcting the target inner torque low by the guard value. For this reason, it is possible to make the backlash of the output shaft clogged in the forward direction when passing through the intermediate speed, and it is possible to engage the friction engagement element in this state.

  According to the control device for an automatic transmission of the present invention, when performing a power-on downshift that shifts from the current shift speed to the target shift speed via the intermediate shift speed, the shock when passing through the intermediate shift speed is reduced. It can be suppressed.

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 which passes through the intermediate gear stage by the comparative example. 5 is a timing chart showing an example of a power on downshift via an intermediate gear according to the present embodiment. It is a flowchart for demonstrating the calculation method of the releasing side clutch torque at the time of a power-on downshift.

  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. This vehicle 100 is, for example, an FF (front engine and front drive) method, and the output of the engine 1 is transmitted to the differential device 6 through the torque converter 2 and the automatic transmission 3 to drive 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 which 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 this automatic transmission 3, an input shaft 3a is connected to a turbine runner 22 of the torque converter 2, and an output shaft 3b is connected to a drive wheel 7 via a differential device 6 or 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 is configured to rotate 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 meshing 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 32 a and the third planetary gear device 32 b.

  The sun gear S2 is selectively coupled to the transmission case 30 by the first brake B1. In addition, the sun gear S2 is selectively coupled to the ring gear R1 via the third clutch C3. Furthermore, the sun gear S2 is selectively coupled to the planetary carrier CA1 via the fourth clutch C4. The sun gear S3 is selectively coupled to the ring gear R1 via the first clutch C1. The planetary carrier RCA is selectively coupled to the transmission case 30 by the second brake B2. 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 is configured to rotate 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 engagement of the first clutch C1 and the first brake B1 establishes a second gear (2nd).

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

-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 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 for storing 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 to calculate the rotational speed (angular velocity) 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 rotational speed sensor 83 is provided to calculate the rotational 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 degree sensor 85 is provided to detect the throttle opening degree 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 to adjust the ignition timing by the spark 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 the shift control of the automatic transmission 3 and the control of the lockup 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 speed is set based on the shift map using the vehicle speed and the accelerator opening as parameters, and the hydraulic control device 4 is controlled such that the actual shift speed becomes the target shift speed. 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 general shift control, for example, each friction relating to shift is determined based on a control map predetermined by adaptation while evaluating whether or not shift shock, shift time, etc. are appropriate on an actual vehicle. 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 combinations of shift patterns such as power on downshift and power off upshift, and gear positions 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 shift. The control operation amount is a required value of an element (engine torque, clutch torque or the like) operated on the 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 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. It represents the acceleration of the input shaft 3a as a quantity (hereinafter referred to as input axis acceleration). dω o / dt is a time change rate of the transmission output shaft rotational speed ω o and represents 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, transmission input torque Ti. 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 a transmission output torque which is a torque on the output shaft 3b as a 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 differs depending on, for example, the type of shift (for example, a combination of shift pattern and shift speeds before and after shift). Therefore, although one predetermined motion equation is used as the equation of motion, an 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 represent the shift time. Further, the transmission output torque To 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 transmission output 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, for example, the torque on the input shaft 3a (total on the input shaft 3a) that needs to be held by the release side clutch and the engagement side clutch when shifting the automatic transmission 3 (total transmission torque). (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 release side clutch torque Tcdrn is based on "Tcapl" and "Tcdrn" replaced with the torque on the input shaft 3a, and the above-mentioned equations (3) and (4). , “X” (= xapl) and “1-x” (= xdrn). Then, from the equation (1), the equation (2), and the relational equation between “Tcapl” and “Tcdrn”, the turbine torque Tt, the engagement side clutch torque Tcapl, which is the control operation amount, and the release side A relational expression for calculating the clutch torque Tcdrn is derived. Turbine torque Tt (also agrees with engine torque Te) is a relational expression using “x” (= xapl), “1-x” (= xdrn), input shaft acceleration dωt / dt, transmission output torque To, etc. Is represented by Similarly, the engagement-side clutch torque Tcapl is expressed by a relational expression using “x” (= xapl), the input shaft acceleration dωt / dt, the transmission output torque To, and the like. Similarly, the release side clutch torque Tcdrn is expressed by a relational expression using “1−x” (= xdrn), input shaft acceleration dωt / dt, transmission output 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 the present embodiment, the shift condition of the automatic transmission 3 is executed using the shift model by adding the constraint condition set at 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 transmission of the transmission 3, a transmission model corresponding to each type of transmission is used.

-Power-on downshift via intermediate gear-
Next, an example of a power-on downshift that passes through the intermediate shift speed will be described with reference to FIGS. 5 and 6. In the following, the power on downshift via the intermediate gear according to the conventional comparative example will be described with reference to FIG. 5, and then the power via the intermediate gear according to the present embodiment will be described with reference to FIG. The on-down shift will be described.

  In the example of FIGS. 5 and 6, 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 as the current shift stage, and the third shift stage as the target shift stage. The gear is set. At this time, the shift from the eighth shift stage to the third shift stage requires the release of the second clutch C2 and the first brake B1 and the engagement of the first clutch C1 and the third clutch C3. The fifth shift speed is set as the intermediate shift speed. Therefore, a shift (first shift) from the eighth shift stage (current shift stage) to the fifth shift stage (intermediate shift stage) and a shift from the fifth shift stage to the third shift stage (target shift stage). (Second shift) is continuously performed. The synchronous rotational speed of the fifth gear is calculated based on the gear ratio of the fifth gear and the rotational speed of the output shaft 3b, and the synchronous rotational speed of the third gear is the gear ratio of the third gear. And the rotational speed of the output shaft 3b.

  In the comparative example, as shown in FIG. 5, when the target inertia torque for proceeding with the shift rises at time t <b> 1, the required torque for the first brake B <b> 1 that is the friction engagement element on the release side of the first shift. Is lowered. As a result, the first brake B1 is released, and the input shaft rotational speed is increased from the synchronous rotational speed of the eighth shift stage. The reduction of the required torque shown by hatching in FIG. 5 corresponds to the inner torque. The target inertia torque is calculated based on the target input axis acceleration. Further, the required torque for the first brake B1 is calculated based on the turbine torque and the target inertia torque.

  Then, at time t2 when the input shaft rotational speed reaches the synchronous rotational speed of the fifth shift stage, a request for the second clutch C2, which is a friction engagement element on the release side of the second shift, so as to obtain a target inner inertia. The torque is reduced. Further, the required torque for the first clutch C1, which is an engagement holding element for the second shift, rises, and the first clutch C1 is engaged. The required torque for the first brake B1, which is a friction engagement element on the release side of the first shift, is reduced toward zero, and the first brake B1 is completely released.

  Thereafter, when the input shaft rotational speed is near the synchronous rotational speed of the third gear, the target inner inertia torque is reduced, and the required torque for the second clutch C2 is increased accordingly. At time t3 when the input shaft rotational speed reaches the synchronous rotational speed of the third gear, the required torque for the second clutch C2 decreases toward zero and the required torque for the third clutch C3 rises. For this reason, the second clutch C2 is completely released and the third clutch C3 is engaged. As a result, the first clutch C1 and the third clutch C3 are engaged to establish the third shift speed.

  Here, in this comparative example, when passing through the intermediate gear, that is, when the input shaft rotational speed passes the synchronous rotational speed of the fifth gear, a fluctuation (shock) of the vehicle longitudinal G occurs. This is a movement in which the internal rotation of the second transmission unit (rear planetary) 32 is reversed when passing through the intermediate shift stage. At this time, since the torque capacity of the first brake B1 is small, the backlash of the output shaft 3b ( This is because the first clutch C1 is engaged in this state and a torque in the positive direction is applied to the output shaft 3b.

  Therefore, in the present embodiment, the ECU 5 calculates a guard value (upper limit guard value) based on the degree of shift progress to the intermediate shift stage and corrects the target inertia torque using the guard value in order to suppress the shock. Is configured as. Specifically, when the ECU 5 approaches the intermediate gear, the ECU 5 increases the torque capacity of the friction engagement element on the release side of the first shift by correcting the target inner torque low by the guard value. As a result, the backlash of the output shaft 3b is clogged in the forward direction when passing through the intermediate speed, and the engagement holding element of the second speed change is engaged in this state.

  In the present embodiment, as shown in FIG. 6, when the target inertia torque for proceeding with the shift rises at time t11, the required torque for the first brake B1, which is the friction engagement element on the release side of the first shift, is increased. Be reduced. As a result, the first brake B1 is released, and the input shaft rotational speed is increased from the synchronous rotational speed of the eighth shift stage. The reduction of the required torque shown by hatching in FIG. 6 corresponds to the inner torque. Further, the required torque for the first brake B1 is calculated based on the turbine torque and the target inertia torque.

  Here, in the present embodiment, the target inertia torque is calculated based on the target input axis acceleration, and the target inertia torque is corrected by performing a guard process. The guard value used for the guard processing is calculated based on the shift progress degree of the first shift. Note that the shift progress of the first shift is calculated by, for example, the following equation (5).

Shift progress of the first shift = (current input shaft rotational speed−synchronous rotational speed before start of shifting) / (synchronous rotational speed of intermediate shift stage−synchronous rotational speed before start of shift)
Then, if the degree of shift progress of the first shift is less than the predetermined value, a guard value higher than the target inertia torque calculated based on the target input shaft acceleration is set, so that the target value before correction is The target value after correction is used as it is. That is, when the degree of shift progression of the first shift is less than a predetermined value, the target inner inertia is a value calculated based on the target input shaft acceleration, and is controlled in the same manner as in the comparative example. This predetermined value is a threshold value for determining whether or not the vehicle has approached the intermediate gear, and is a preset value (for example, 0.7).

  After that, when the degree of shift progression of the first shift becomes equal to or greater than a predetermined value and the input shaft rotational speed approaches the synchronous rotational speed of the intermediate gear, a low guard value is set. If guard processing is performed using this guard value, the target inner inertia will be a value limited by the guard value. That is, the target value after correction becomes lower than the target value before correction calculated based on the target input shaft acceleration. That is, in the present embodiment, the target inner torque is corrected to be lower as the intermediate gear is approached.

  For this reason, the required torque for the first brake B1, which is the friction engagement element on the release side of the first shift, becomes high. For this reason, at time t12 when the input shaft rotational speed reaches the synchronous rotational speed of the fifth shift stage, with the backlash of the output shaft 3b being clogged in the forward direction, The required torque for the clutch C1 rises and the first clutch C1 is engaged. For this reason, in this embodiment, the fluctuation (shock) of the vehicle front-rear G when the input shaft rotational speed passes the synchronous rotational speed of the fifth gear is suppressed as compared with the comparative example.

  When the second shift is started, the guard process for the target inertia torque is not performed. That is, the target inertia torque is calculated based on the target input axis acceleration. Further, the required torque for the second clutch C2, which is the friction engagement element on the release side of the second shift, is reduced so that the target inertia torque is obtained. Note that the required torque for the first brake B1, which is the friction engagement element on the release side of the first shift, is reduced toward zero, and the first brake B1 is completely released.

  Thereafter, when the input shaft rotational speed is near the synchronous rotational speed of the third gear, the target inner inertia torque is reduced, and the required torque for the second clutch C2 is increased accordingly. At time t13 when the input shaft rotational speed reaches the synchronous rotational speed of the third gear, the required torque for the second clutch C2 is reduced toward zero and the required torque for the third clutch C3 rises. For this reason, the second clutch C2 is completely released and the third clutch C3 is engaged. As a result, the first clutch C1 and the third clutch C3 are engaged to establish the third shift speed.

[Calculation of release side clutch torque]
Next, with reference to FIG. 7, a calculation method of the release side clutch torque at the time of power-on downshift will be described. The following steps are executed by the ECU 5.

  First, in step ST1 of FIG. 7, it is determined whether a power on downshift is performed. Specifically, it is determined that a power-on downshift is performed when the accelerator pedal is depressed and a downshift determination is made. If it is determined that the power on downshift is to be performed, the process proceeds to step ST2. On the other hand, when it is determined that the power-on downshift is not performed, step ST1 is repeatedly performed. That is, it waits until a power-on downshift is performed.

  Next, in step ST2, a target input shaft acceleration (input shaft acceleration which is a shift target value) 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.

  Next, in step ST3, the target input axis acceleration is converted into a target inertia torque. That is, the target input shaft acceleration is multiplied by a predetermined moment of inertia to calculate the target inner inertia.

  Next, in step ST4, turbine torque (input shaft torque) is estimated. The turbine torque is calculated based on, for example, the estimated engine torque calculated based on the throttle opening and the engine speed, and the torque ratio of the torque converter 2.

  Next, in step ST5, it is determined whether or not the intermediate speed stage is passed. Whether or not the intermediate gear position is to be passed is determined by, for example, whether or not the release side clutch is switched when passing the destination gear position of the currently executed shift. If it is determined that the intermediate speed stage is to be passed, the process proceeds to step ST6. On the other hand, when it is determined that the intermediate gear is not to be passed, the process proceeds to step ST9.

  Next, in step ST6, the shift progress of the first shift (the currently executed shift) is calculated. The shift progress degree of the first shift is calculated, for example, based on the above-mentioned equation (5).

  Then, in step ST7, a guard value (upper limit guard value) is calculated based on the shift progress degree of the first shift. For example, when the shift progress is less than a predetermined value, a high guard value is calculated so as not to unnecessarily limit the target inertia calculated based on the target input shaft acceleration, and the shift progress is greater than or equal to the predetermined value. In this case, a low guard value is calculated in order to increase the required torque for the friction engagement element on the release side of the first shift.

  Next, in step ST8, a guard process is performed on the target inertia torque. Specifically, when the target inertia torque calculated in step ST3 is lower than the guard value calculated in step ST7, the value calculated in step ST3 becomes the target inertia torque as it is and is calculated in step ST3. If the target inner inertia is higher than the guard value calculated at step ST7, the guard calculated at step ST7 becomes the target inner inertia. For example, when the shift progress of the first shift is less than a predetermined value and the guard value is high, the value calculated in step ST3 becomes the target inertia torque as it is, and the shift progress of the first shift is greater than or equal to the predetermined value. If the guard value is low, the guard value becomes the target inertia torque.

  Next, in step ST9, the release side clutch torque is calculated based on the turbine torque and the target inner torque. Specifically, the release-side clutch torque is calculated by subtracting an inner torque for advancing the shift from the clutch torque capacity that is in balance with the turbine torque. When the target inner torque is corrected by performing guard processing, the release side clutch torque is calculated using the corrected target inner torque.

  Next, in step ST10, it is determined whether or not the power-on downshift is completed. That is, it is determined whether the switching to the target shift speed set in step ST1 is completed. If it is determined that the power-on downshift has not been completed, the process returns to step ST2. On the other hand, if it is determined that the power-on downshift has been completed, the process returns.

-Effect-
In the present embodiment, as described above, the guard value (upper limit guard value) is calculated based on the degree of shift progress to the intermediate gear position, and the target inertia torque is corrected using the guard value to obtain the intermediate gear position. When approaching, the torque capacity of the frictional engagement element on the release side of the first shift can be increased by correcting the target inner torque low by the guard value. As a result, when passing through the intermediate gear, the backlash of the output shaft 3b is clogged in the forward direction, and the engagement / holding element of the second gear can be engaged in that state. As a result, when performing a power on downshift that shifts from the current gear position to the target gear position via the intermediate gear position, it is possible to suppress the shock when passing the intermediate gear position.

-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, although the example in which the vehicle 100 is an FF is shown in the present embodiment, the present invention is not limited thereto, and the vehicle may be an FR (front engine rear drive) or four-wheel drive. .

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

  In the present embodiment, an example in which the guard process is performed on the target inertia torque is shown, but the present invention is not limited to this, and the guard process may be performed on the target input axis acceleration.

  Further, in this embodiment, an example is shown in which the guard value is increased when the shift progress degree of the first shift is less than the predetermined value, and the guard value is decreased when the shift progress degree of the first shift is equal to or more than the predetermined value. However, the present invention is not limited to this, and the guard value may be calculated based on the differential rotation between the synchronous rotational speed of the intermediate gear and the current input shaft rotational speed.

  Further, in the present embodiment, an example in which the intermediate shift stage is set because it is necessary to change the four elements is not limited to this. An intermediate speed may be set to reduce the friction load of the engagement element.

  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 Third 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 shift progress at the time of power-on downshift is configured to be performed by controlling the frictional engagement element on the release side based on the target inertia torque,
During a power-on downshift to shift from the current gear position to the target gear position via the intermediate gear position, a guard value is calculated based on the degree of shift progress to the intermediate gear position, and the target is calculated using the guard value. A control device for an automatic transmission, which is configured to correct an inertia torque.

Images