JP6852464B2 - Control device for automatic transmission - Google Patents

Description

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

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 engaging 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 gear to a target gear via an intermediate gear. Specifically, the control device of this automatic transmission has a sixth speed stage established by the engagement of the second clutch and the first brake, and a third speed stage established by the engagement of the first clutch and the third clutch. When shifting to, the vehicle goes through the fourth speed stage, which is established by the engagement of the first clutch and the second clutch. That is, since it is necessary to change the grip of four elements for shifting from the sixth speed to the third speed, shifting from the sixth speed to the fourth speed, which can be switched by two elements (first shifting). And, the shift from the 4th speed to the 3rd speed (second shift), which can be switched by two elements, is continuously performed. In such a 6-4-3 shift, by preparing for the second shift in advance during the execution of the first shift, the first shift and the second shift are individually performed in two stages, as compared with the case where the first shift and the second shift are individually performed in two stages. , It is possible to shorten the shifting time.

JP-A-2002-310281

Here, in the conventional automatic transmission described above, the internal rotation of the rear planetary constituting the transmission mechanism is reversed when passing through the intermediate transmission stage. The reversal of the internal rotation of the rear planetary means that when the rotation speed of the ring gear connected to the output shaft is constant, the direction of change in the rotation speed of the sun gear and the planetary carrier with the progress of shifting is reversed. At that time, the torque capacity of the first brake, which is the friction engagement element on the release side of the first shift, is small, and the torque capacity at the time of preparation of the first clutch, which is the friction engagement element on the engagement side of the first shift. If is relatively large, the torque applied to the output shaft becomes smaller than zero, and the backlash (backlash) of the output shaft becomes clogged on the opposite side. In that state, when the input shaft rotation speed passes through the synchronous rotation speed of the intermediate shift stage, the direction in which the torque capacity of the first clutch, which is the frictional engagement element on the engaging side, acts is reversed, and the direction is positive to the output shaft. Since torque is applied, a shock due to rattling may occur.

Therefore, it is conceivable to increase the torque capacity of the friction engaging element on the release side when the shift progress to the intermediate shift becomes equal to or higher than a predetermined value. As a result, the backlash of the output shaft can be brought into a state of being clogged in the positive direction when passing through the intermediate shift stage, so that it is possible to suppress a shock.

However, when the actual input shaft acceleration fluctuates with respect to the target input shaft acceleration, the torque capacity of the friction engaging element on the release side of the first shift when passing through the synchronous rotation speed of the intermediate shift stage is appropriately secured. It becomes difficult. Specifically, when the actual input shaft acceleration is larger than the target input shaft acceleration, the shift progresses faster, so that the input shaft rotation speed synchronizes with the intermediate shift stage after the shift progress exceeds a predetermined value. A shock occurs because the time required to reach the rotation speed is shortened and the torque capacity of the friction engaging element on the release side is insufficient. On the other hand, if the actual input shaft acceleration is smaller than the target input shaft acceleration, the speed change progresses slowly. Therefore, the input shaft rotation speed becomes the synchronous rotation of the intermediate speed change stage after the speed change progress exceeds a predetermined value. It takes a long time to reach the speed, and the torque capacity of the friction engaging element on the release side becomes excessive, so that the shift is stagnant. It should be noted that such a problem is unknown.

The present invention has been made to solve the above problems, and an object of the present invention is to perform a power-on-downshift that shifts from a current gear to a target gear via an intermediate gear. , Even when the actual input shaft acceleration fluctuates with respect to the target input shaft acceleration, it is possible to suppress the stagnation of the shift progress while suppressing the shock when passing through the intermediate shift stage. It is to provide a control device for the machine.

The control device for an automatic transmission according to the present invention is applied to an automatic transmission that establishes a plurality of gears by selectively engaging a plurality of friction engaging elements. The automatic transmission control device predicts the time required for the input shaft rotation speed to reach the synchronous rotation speed of the intermediate gear during a power-on-down shift that shifts from the current gear to the target gear via the intermediate gear. Is calculated, and when the predicted time becomes less than or equal to the first predetermined value, the torque capacity of the friction engaging element on the release side is corrected to increase , and the predicted time is smaller than the first predetermined value and is less than or equal to the second predetermined value. if it becomes, the increase correction of the torque capacity of the frictional engagement element on the release side is configured so that the ends.

With this configuration, even if the actual input shaft acceleration varies with respect to the target input shaft acceleration, the torque capacity of the friction engaging element on the release side can be increased at an appropriate timing. It is possible to suppress the stagnation of the progress of the shift while suppressing the shock when passing through the shift stage.

According to the control device of the automatic transmission of the present invention, when performing a power-on-down shift that shifts from the current shift stage to the target shift stage via the intermediate shift stage, the actual input shaft is relative to the target input shaft acceleration. Even when the acceleration fluctuates, it is possible to suppress the stagnation of the shift progress while suppressing the shock when passing through the intermediate shift stage.

It is a figure which showed the schematic structure of the vehicle which comprises the ECU by one Embodiment of this invention. It is a skeleton diagram which showed the structure of the torque converter and the automatic transmission of FIG. It is an engagement table which showed the engagement state of the 1st clutch to the 4th clutch, the 1st brake and the 2nd brake for each shift stage in the automatic transmission of FIG. It is a block diagram which showed the ECU of FIG. It is a timing chart which showed an example of the case where the actual input shaft acceleration is as intended with respect to the target input shaft acceleration at the time of power-on-downshift via the intermediate shift stage according to the comparative example. It is a timing chart which showed an example of the case where the actual input shaft acceleration is larger than the target input shaft acceleration at the time of power-on-downshift via the intermediate shift stage according to the comparative example. It is a timing chart which showed an example of the case where the actual input shaft acceleration is smaller than the target input shaft acceleration at the time of power-on-downshift via the intermediate shift stage according to the comparative example. It is a timing chart which showed an example of the case where the actual input shaft acceleration is as intended with respect to the target input shaft acceleration at the time of power-on-downshift via the intermediate shift stage according to the present embodiment. It is a timing chart which showed an example of the case where the actual input shaft acceleration is larger than the target input shaft acceleration at the time of power-on-downshift via the intermediate shift stage according to this embodiment. It is a timing chart which showed an example of the case where the actual input shaft acceleration is smaller than the target input shaft acceleration at the time of power-on-downshift via the intermediate shift stage according to the present embodiment. It is a flowchart for demonstrating the calculation method of the clutch torque at the time of shift control by this Embodiment.

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

First, the vehicle 100 including the ECU 5 according to the embodiment of the present invention will be described with reference to FIGS. 1 to 4.

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 the left and right drive wheels (front wheels) 7 are transmitted. It is designed to be distributed to.

-Engine-
The engine (internal combustion engine) 1 is a driving force source for traveling, and is, for example, a multi-cylinder gasoline engine. The engine 1 is configured so that the operating state can be controlled by the throttle opening (intake air amount) of the throttle valve, the fuel injection amount, the 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 an engine 1, a turbine runner 22 connected to an 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 are included. In FIG. 2, with respect to the rotation center axis of the torque converter 2 and the automatic transmission 3, only the upper half is schematically shown by omitting the lower half.

-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 3a is connected to the turbine runner 22 of the torque converter 2, and the output shaft 3b is connected to the drive wheels 7 via a differential device 6 or the like.

The automatic transmission 3 is composed mainly of a first transmission unit (front planetary) 31, a second planetary gear device 32a, and a third planetary gear device 32b, which are mainly composed of a first planetary gear device 31a. It is composed of a transmission unit (rear planetary) 32, a first clutch C1 to a fourth clutch C4, a first brake B1 and a second brake B2.

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

The planetary carrier CA1 is connected 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 deceleration rotation to the second transmission unit 32.

The second planetary gear device 32a constituting the second transmission unit 32 is a single pinion type planetary gear mechanism, and includes a sun gear S2, a pinion gear P2, and a planetary carrier RCA that supports the pinion gear P2 so as to rotate and revolve. , A ring gear RR that meshes with the sun gear S2 via a pinion gear P2 is provided.

Further, 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 that mesh with each other, and the pinion gears P2 and P3. It includes a planetary carrier RCA that supports rotation and revolution, 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. Further, the sun gear S2 is selectively connected to the ring gear R1 via the third clutch C3. Further, the sun gear S2 is selectively connected to the planetary carrier CA1 via the fourth clutch C4. The sun gear S3 is selectively connected to the ring gear R1 via the first clutch C1. The planetary carrier RCA is selectively connected to the transmission case 30 by the second brake B2. Further, the planetary carrier RCA is selectively connected 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 engaging elements that are frictionally engaged by the hydraulic actuator, and are controlled by the hydraulic control device 4 and the ECU 5.

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

As shown in FIG. 3, in the automatic transmission 3 of this example, the gear ratio (rotational speed of the input shaft 3a / rotational speed of the output shaft 3b) is caused by engaging the first clutch C1 and the second brake B2. The first gear (1st) having the largest value is established. The second shift stage (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) is established 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 shift (7th) is established by engaging the second clutch C2 and the third clutch C3, and the eighth shift (8th) is established by engaging the second clutch C2 and the first brake B1. Is established. The reverse stage (Rev) is established by engaging the third clutch C3 and the second brake B2.

-Flood control device-
The hydraulic control device 4 is provided to control the state (engaged state or released state) of the friction engaging 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 "control device for an automatic transmission" of 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 to be referred to when executing the various control programs, and the like. The RAM 53 is a memory that temporarily stores the calculation result by the CPU 51 and the detection result of each sensor. The backup RAM 54 is a non-volatile memory that stores data and the like to be saved when the ignition is turned off.

A crank position sensor 81, an input shaft rotation speed sensor 82, an output shaft rotation 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 rotation speed sensor 82 is provided to calculate the rotation speed (turbine rotation speed) of the input shaft 3a of the automatic transmission 3. The output shaft rotation speed sensor 83 is provided to calculate the rotation speed of the output shaft 3b of the automatic transmission 3. The accelerator opening sensor 84 is provided to detect the accelerator opening, which is the amount of depression (operation amount) of the accelerator pedal. The throttle opening sensor 85 is provided to detect the throttle opening of the throttle valve.

An injector 91, an igniter 92, a throttle motor 93, a hydraulic control device 4, and the like are connected to the output interface 56. The injector 91 is a fuel injection valve, and the fuel injection amount can be adjusted. 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 degree 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 degree, the fuel injection amount, the ignition timing, and the like based on the detection results of each sensor. Further, the ECU 5 is configured to be able to execute the shift control of the automatic transmission 3 and 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 is set based on the shift map with the vehicle speed and the accelerator opening as parameters, and the hydraulic control device 4 is controlled so that the actual shift becomes the target shift. That is, the ECU 5 makes a shift determination based on the shift map, and executes the shift control so that the target shift stage can be obtained when it is determined that the shift should be executed.

In this shift control, switching to a shift stage established by releasing one friction engaging element and engaging one friction engaging element is permitted, and releasing two friction engaging elements and two frictions. Switching to a gear that requires engagement of the engaging element is prohibited. In addition, it is possible to switch to a shift gear that is two or more gears away from the current shift gear.

-Automatic transmission shift control-
Here, as general shift control, for example, while evaluating whether or not the shift shock and shift time are appropriate in the actual vehicle, each friction clutch during shifting is based on a control map predetermined by conformity. There is a method of determining the torque capacity (or hydraulic command value) of the combined element (the clutch and the brake) and executing the shifting. In the method using this control map, it is necessary to create a large number of control maps according to the combination of shift patterns such as power-on-downshift and power-off-upshift and shift stages before and after shift. Therefore, the more gears the automatic transmission has, the more labor is required for the fitting work.

Therefore, in the present embodiment, as the shift control, instead of the method using the control map, a method of executing the shift using a shift model for determining the control operation amount for realizing the shift target value is adopted. 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, etc.) to be operated with respect to the control target.

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

The equations (1) and (2) are derived from the equations of motion for each of the interconnected rotating elements constituting the automatic transmission 3 and the relational equations in the planetary gear device constituting the automatic transmission 3. It is a thing. In the equation of motion for each rotating element, the torque represented by the product of the inertia and the rate of change in rotational speed and time in each rotating element is obtained from the three members of the planetary gear device and the members on both sides of the friction engaging element. It is an equation of motion defined by the torque acting on the members involved in each rotating element. Further, the relational expression in the planetary gear device is a relational expression in which the relationship between the torque and the relationship between the rotation speed time change rate in the three members of the planetary gear device are defined by using the gear ratio of the planetary gear device.

In equations (1) and (2), dωt / dt is the time derivative of the turbine rotation speed ωt (that is, the transmission input shaft rotation speed ωi), that is, the time change rate, and the speed change 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 the input shaft acceleration). dωo / dt is the time change rate of the transmission output shaft rotation speed ωo and represents the output shaft acceleration. Tt represents the turbine torque, that is, the transmission input torque Ti, which is the torque on the input shaft 3a as the torque on the rotating member on the input shaft 3a side. This turbine torque Tt agrees with the engine torque Te (= Tt / t) in consideration of the torque ratio t of the torque converter 2. To represents the transmission output torque, which is the torque on the output shaft 3b as the torque on the rotating member on the output shaft 3b side. Tcapl is the torque capacity (hereinafter, also referred to as the engaging side clutch torque) of the friction engaging element that engages in the engaging operation at the time of shifting. Tcdrn is the torque capacity (hereinafter, also referred to as release side clutch torque) of the friction engaging element that performs the release operation at the time of shifting. a1, a2, b1, b2, c1, c2, d1, d2 are constants when the equations (1) and (2) are derived, respectively, and the inertia and the planetary gear in each of the rotating elements are used. 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 shifting (for example, the shifting pattern and the combination of shifting stages before and after shifting). Therefore, although the equation of motion is one predetermined one, the equation of motion corresponding to each type of shifting, which is a constant different for each type of shifting, is used for the shifting of the automatic transmission 3.

The equations (1) and (2) are gear train equations of motion of the automatic transmission 3 that formulate the relationship between the shift target value and the control operation amount. 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 equation of motion. In this embodiment, the input shaft acceleration dωt / dt is used as an example of the physical quantity capable of expressing the shift time. Further, as an example of a physical quantity capable of expressing a driving force, a transmission output torque To is used. That is, in the present embodiment, the shift target value is set by two values, 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 set to three values of turbine torque Tt (engine torque Te also agrees), engaging side clutch torque Tcapl, and releasing side clutch torque Tcdrn. It is set. Then, since there are three control operation amounts for the equation of motion composed of the two equations (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 rotation speed ωo, which is a detection value of the output shaft rotation speed sensor 83.

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

xapl = x ... (3)
xdrn = 1-x ... (4)
The relational expression between the engaging side clutch torque Tcapl and the releasing side clutch torque Tcdrn is based on "Tcapl" and "Tcdrn" replaced with the torque on the input shaft 3a, and the above equations (3) and (4). , "X" (= xapl) and "1-x" (= xdrn) can be defined. Then, from the equation (1), the equation (2), and the relational expression between "Tcapl" and "Tcdrn", the turbine torque Tt, the engaging side clutch torque Tcapl, and the release side, which are the control operation amounts, 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, and transmission output torque To. It is represented by. Similarly, the engaging 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), the input shaft acceleration dωt / dt, the transmission output torque To, and the like.

That is, the shift model of the present embodiment has a relationship (the equations (1) and (2)) of the automatic transmission 3 including the shift target value and the control operation amount and the torque sharing ratio (the torque sharing ratio). The control operation amount is calculated based on the shift target value by using the equations (3) and (4)). As described above, in the present embodiment, the automatic transmission 3 is changed by using the speed change model by adding the restraint condition set by the torque sharing ratio x to the above equations (1) and (2). .. Therefore, even if there are three control operation amounts for the two shift target values, the three control operation amounts can be appropriately determined using the shift model. This shifting model is one predetermined model, but as described above, a gear train equation of motion with different constants is used for each type of shifting (for example, a combination of shifting patterns and shifting stages before and after shifting), so that it is automatic. For the shifting of the transmission 3, a shifting model corresponding to each type of shifting is used.

-Power on downshift via intermediate gears-
Next, an example of a power-on-downshift via the intermediate shift stage will be described with reference to FIGS. 5 to 10.

In the examples of FIGS. 5 to 10, a downshift is determined based on the shift map by depressing the accelerator pedal from the state where the eighth shift is established as the current shift, and the third shift is set as the target shift. The gear is set. At this time, in order to shift from the eighth gear to the third gear, it is necessary to release the second clutch C2 and the first brake B1 and engage the first clutch C1 and the third clutch C3. , The fifth gear may be set as the intermediate gear. In this case, the shift from the 8th shift (current shift) to the 5th shift (intermediate shift) (1st shift) and the shift from the 5th shift to the 3rd shift (target shift). (Second shift) is performed continuously. The synchronous rotation 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 rotation speed of the third gear is the gear ratio of the third gear. And the rotation speed of the output shaft 3b.

Here, in the power-on-downshift via the intermediate gears as shown in FIGS. 5 to 10, the internal rotation of the second transmission (rear planetary) 32 is reversed when passing through the intermediate gears. .. Therefore, the torque capacity of the friction engaging element (first brake B1) on the release side of the first shift is small, and the friction engagement element (first clutch C1) on the engagement side of the first shift is prepared. When the torque capacity is relatively large, the torque applied to the output shaft 3b becomes smaller than zero due to the drag of the first clutch C1, and the backlash of the output shaft 3b becomes clogged on the opposite side. In this state, when the input shaft rotation speed passes through the synchronous rotation speed of the intermediate shift stage, the direction in which the torque capacitance of the first clutch C1 acts is reversed (transmitted to the second transmission unit 32 side via the first clutch C1). The direction of the torque is reversed), and the backlash of the output shaft 3b may be clogged in the positive direction, causing a shock due to the backlash.

(In the case of comparative example)
Therefore, in the comparative example, the torque capacity of the friction engaging element on the release side is increased when the shift progress to the intermediate shift stage becomes a predetermined value Th1 or more. In such a comparative example, when the actual input shaft acceleration (actual input shaft acceleration) is as intended with respect to the target input shaft acceleration, it is possible to suppress the shock, but the target input shaft acceleration is used. On the other hand, when the actual input shaft acceleration varies, the problem described later occurs. First, with reference to FIG. 5, a case where the actual input shaft acceleration is as intended with respect to the target input shaft acceleration at the time of power-on-downshift via the intermediate shift stage according to the comparative example will be described. After that, with reference to FIGS. 6 and 7, a case where the actual input shaft acceleration varies from the target input shaft acceleration during the power-on-downshift via the intermediate shift stage according to the comparative example will be described.

[When the input shaft rotation speed is as intended in the comparative example]
As shown in FIG. 5, when the third gear is set as the target gear at the time point t1 from the state where the eighth gear is established, the fifth gear is set as the intermediate gear. As a result, the first shift for switching from the eighth shift to the fifth shift is started. Then, at time t2, the inertia phase of the first shift is started. That is, the required torque for the first brake B1, which is the frictional engaging element on the release side of the first shift, is reduced, so that the first brake B1 is released and the input shaft rotation speed is the synchronous rotation speed of the eighth shift stage. Rise from.

As the input shaft rotation speed approaches the synchronous rotation speed of the fifth gear from the synchronous rotation speed of the eighth gear, the shift progress of the first gear increases. 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 rotation speed-synchronous rotation speed of the eighth shift) / (synchronous rotation speed of the fifth shift-synchronous rotation speed of the eighth shift)} x 100 ... (5)
Then, at the time point t3, when the shift progress of the first shift reaches the predetermined value Th1, the required torque for the first brake B1 which is the friction engaging element on the release side of the first shift is temporarily increased. As a result, at the time t4 when the input shaft rotation speed reaches the synchronous rotation speed of the fifth shift stage, the backlash of the output shaft 3b becomes clogged in the positive direction. The predetermined value Th1 is a preset value (for example, 80%).

At time point t4, the torque phase of the first shift is started. That is, 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 clutch C1 which is the friction engagement element on the engagement side of the first shift is reduced. The required torque for is increased. At this time, since the first clutch C1 is engaged in a state where the backlash of the output shaft 3b is clogged on the positive direction side, it is possible to suppress the shock.

Further, at the time point t4, the inertia phase of the second shift which switches from the fifth shift to the third shift is started. That is, the required torque for the second clutch C2, which is the frictional engaging element on the release side of the second shift, is reduced, so that the second clutch C2 is released and the input shaft rotation speed is the synchronous rotation speed of the fifth shift stage. Rise from.

Then, at the time point t5, the torque phase of the first shift is terminated. Therefore, the replacement of the first brake B1 which is the friction engaging element on the release side and the first clutch C1 which is the friction engaging element on the engaging side in the first shift is completed.

After that, at the time t6 when the input shaft rotation speed reaches the synchronous rotation speed of the third shift stage, the torque phase of the second shift is started. That is, the required torque for the second clutch C2, which is the friction engagement element on the release side of the second shift, is reduced toward zero, and the third clutch C3, which is the friction engagement element on the engagement side of the second shift, is reduced. The required torque for is increased. Then, at the time point t7, the torque phase of the second shift is terminated. Therefore, the replacement of the second clutch C2, which is the friction-engaging element on the release side, and the third clutch C3, which is the friction-engagement element on the engagement side, in the second shift is completed.

As described above, in the comparative example, at the time of power-on-downshift via the intermediate shift stage, the input shaft rotation is performed by temporarily increasing the release side clutch torque of the first shift according to the shift progress of the first shift. If the speed is as intended, it is possible to suppress the shock.

[When the input shaft rotation speed is large in the comparative example]
Since the time points t11 to t12 in FIG. 6 are the same as those in the above time points t1 to t2, the description thereof will be omitted.

Then, as shown in FIG. 6, when the actual input shaft acceleration is larger than the target input shaft acceleration in the inertia phase of the first shift, the actual value (actual value) of the input shaft rotation speed is larger than the target value. growing. That is, since the degree of increase in the input shaft rotation speed is large, the progress of the first shift becomes fast. Therefore, when the input shaft rotation speed is high, the input is performed from the time t13 when the shift progress of the first shift becomes the predetermined value Th1 as compared with the case where the input shaft rotation speed is as intended (in the case of FIG. 5). The time until t14 when the shaft rotation speed reaches the synchronous rotation speed of the fifth gear is shortened.

Therefore, at time t13, even if the required torque for the first brake B1, which is the frictional engaging element on the release side of the first shift, is increased, the time required for increasing the required torque is short, so that the first brake B1 The torque capacity of the output shaft 3b cannot be secured, and the backlash of the output shaft 3b becomes clogged on the opposite side.

In this state, when the torque phase of the first shift is started at the time point t14 and the required torque for the first clutch C1 which is the friction engagement element on the engagement side of the first shift is increased, the first clutch C1 At the time of engagement, the backlash of the output shaft 3b is clogged in the positive direction, and a shock due to the backlash occurs. Since the time points t15 to t17 are the same as the above-mentioned time points t5 to t7, the description thereof will be omitted.

As described above, in the comparative example, the input shaft rotation speed is increased even if the release side clutch torque is temporarily increased according to the shift progress of the first shift at the time of power-on-downshift via the intermediate shift stage. If it is large, a shock due to rattling occurs.

[When the input shaft rotation speed is small in the comparative example]
Since the time points t21 to t22 in FIG. 7 are the same as the above-mentioned time points t1 to t2, the description thereof will be omitted.

Then, as shown in FIG. 7, when the actual input shaft acceleration is smaller than the target input shaft acceleration in the inertia phase of the first shift, the actual value (actual value) of the input shaft rotation speed is larger than the target value. It becomes smaller. That is, since the degree of increase in the input shaft rotation speed is small, the progress of the first shift is slowed down. Therefore, when the input shaft rotation speed is small, the input starts from t23 when the shift progress of the first shift reaches the predetermined value Th1 as compared with the case where the input shaft rotation speed is as intended (in the case of FIG. 5). The time until t24 when the shaft rotation speed reaches the synchronous rotation speed of the fifth shift stage becomes longer.

Therefore, at time t23, if the required torque for the first brake B1, which is the frictional engaging element on the release side of the first shift, is increased, the required torque is increased for a longer time, so that the torque capacity of the first brake B1 is increased. Is more than necessary, and the progress of shifting is stagnant. Since the time points t24 to t27 are the same as the above-mentioned time points t4 to t7, the description thereof will be omitted.

As described above, in the comparative example, the input shaft rotation speed is increased even if the release side clutch torque is temporarily increased according to the shift progress of the first shift at the time of power-on-downshift via the intermediate shift stage. If it is small, the progress of shifting is stagnant.

(In the case of this embodiment)
Therefore, in the present embodiment, at the time of power-on-downshift via the intermediate shift stage, when the synchronous predicted time until the input shaft rotation speed reaches the synchronous rotation speed of the intermediate shift stage becomes a predetermined value Th2 or less, the release is released. It is configured to increase the torque capacity of the side friction engagement element. As a result, even if the actual input shaft acceleration varies with respect to the target input shaft acceleration, the torque capacity of the friction engaging element on the release side is immediately before the input shaft rotation speed reaches the synchronous rotation speed of the intermediate shift stage. It is possible to make the time when the increase is appropriate. First, with reference to FIG. 8, a case where the actual input shaft acceleration is as intended with respect to the target input shaft acceleration at the time of power-on-downshift via the intermediate shift stage according to the present embodiment will be described. After that, with reference to FIGS. 9 and 10, a case where the actual input shaft acceleration varies from the target input shaft acceleration during the power-on-downshift via the intermediate shift stage according to the present embodiment will be described.

[When the input shaft rotation speed is as intended in this embodiment]
As shown in FIG. 8, when the third gear is set as the target gear at the time point t31 from the state where the eighth gear is established, the fifth gear is set as the intermediate gear. As a result, the first shift for switching from the eighth shift to the fifth shift is started. Then, at the time point t32, the inertia phase of the first shift is started. That is, the required torque for the first brake B1, which is the frictional engaging element on the release side of the first shift, is reduced, so that the first brake B1 is released and the input shaft rotation speed is the synchronous rotation speed of the eighth shift stage. Rise from. The rotation of the input shaft 3a in the first shift is mainly controlled by the first brake B1 which is a friction engaging element on the release side. Further, the torque capacity (release side clutch torque) of the first brake B1 is calculated by using the above-mentioned gear train equation of motion.

In the present embodiment, in the inertia phase of the first shift, the synchronous predicted time until the input shaft rotation speed becomes the synchronous rotation speed of the fifth shift stage is calculated. This synchronous prediction time is the remaining time until the input shaft rotation speed reaches the synchronous rotation speed of the fifth shift stage, and becomes smaller as the first shift progresses. The synchronous prediction time is calculated by, for example, dividing the difference between the synchronous rotation speed of the fifth shift stage and the current input shaft rotation speed by the amount of change in the difference per unit time.

Then, at the time point t33, when the synchronization prediction time reaches the predetermined value Th2, the required torque for the first brake B1 which is the friction engaging element on the release side of the first shift becomes high. Specifically, the release side clutch torque calculated by using the gear train equation of motion is increased-corrected, and the increased-corrected value is required for the first brake B1. Next, when the synchronous prediction time decreases as the first shift progresses and the synchronous predicted time reaches a predetermined value Th3 at the time point t34, a request for the first brake B1 which is a friction engaging element on the release side of the first shift is made. The torque is restored. Specifically, the value of the release side clutch torque calculated by using the gear train equation of motion is required for the first brake B1. That is, when the synchronization prediction time is equal to or less than the predetermined value Th2 and larger than the predetermined value Th3 (between the time point t33 and the time point t34), the release side clutch torque is increased and corrected. Therefore, at the time t35 when the input shaft rotation speed reaches the synchronous rotation speed of the fifth shift stage, the backlash of the output shaft 3b becomes clogged in the positive direction.

The predetermined value Th2 is a preset value (for example, 100 msec) and is an example of the " first predetermined value" of the present invention. Predetermined value Th3 is smaller preset value than the predetermined value Th2 (e.g., 40 msec) der is, Ru example der of the "second predetermined value" of the present invention. The predetermined values Th2 and Th3 and the increase correction amount are preset values so that the backlash of the output shaft 3b can be reduced to the positive direction at the time point t35.

At time point t35, the torque phase of the first shift is started. That is, 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 clutch C1 which is the friction engagement element on the engagement side of the first shift is reduced. The required torque for is increased. At this time, since the first clutch C1 is engaged in a state where the backlash of the output shaft 3b is clogged on the positive direction side, it is possible to suppress the shock. Since the time points t36 to t38 are the same as those at the time points t5 to t7 described above, the description thereof will be omitted.

As described above, in the present embodiment, the shock can be suppressed by temporarily increasing the release side clutch torque of the first shift according to the synchronous predicted time at the time of power-on-downshift via the intermediate shift stage. It is possible.

[When the input shaft rotation speed is high in this embodiment]
Since the time points t41 to t42 in FIG. 9 are the same as those at the time points t31 to t32 described above, the description thereof will be omitted.

Then, as shown in FIG. 9, when the actual input shaft acceleration is larger than the target input shaft acceleration in the inertia phase of the first shift, the actual value (actual value) of the input shaft rotation speed is larger than the target value. growing. That is, since the degree of increase in the input shaft rotation speed is large, the progress of the first shift becomes fast. Therefore, the synchronous prediction time is smaller than that in the case of FIG. 8 at the start of the inertia phase of the first shift, but the inclination (decrease degree) is almost the same as in the case of FIG. That is, the time from the time t42 when the inertia phase of the first shift is started to the time t43 when the synchronization prediction time reaches the predetermined value Th2 is shorter than the time from the time t32 to the time t33 in FIG. The time from time to time t45 when the input shaft rotation speed reaches the synchronous rotation speed of the fifth shift stage is almost the same as the time from time point t33 to time point t35 in FIG.

Then, at the time point t43, when the synchronization prediction time reaches the predetermined value Th2 earlier than in the case of FIG. 8, the required torque for the first brake B1 which is the friction engaging element on the release side of the first shift becomes higher. Next, when the synchronous prediction time decreases as the first shift progresses and the synchronous predicted time reaches a predetermined value Th3 at the time point t44, a request for the first brake B1 which is a friction engaging element on the release side of the first shift is made. The torque is restored. Therefore, at the time point t45, the backlash of the output shaft 3b is clogged in the positive direction. That is, since the increase correction of the release side clutch torque is started earlier than in the case of FIG. 8, it is possible to secure the time for performing the increase correction.

At time t45, the torque phase of the first shift is started. That is, 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 clutch C1 which is the friction engagement element on the engagement side of the first shift is reduced. The required torque for is increased. At this time, since the first clutch C1 is engaged in a state where the backlash of the output shaft 3b is clogged on the positive direction side, it is possible to suppress the shock. Since the time points t46 to t48 are the same as the above-mentioned time points t5 to t7, the description thereof will be omitted.

As described above, in the present embodiment, the input shaft rotation speed is increased by temporarily increasing the release side clutch torque of the first shift according to the synchronous predicted time at the time of power-on-downshift via the intermediate shift stage. Even in some cases, it is possible to suppress the shock.

[When the input shaft rotation speed is small in this embodiment]
Since the time points t51 to t52 in FIG. 10 are the same as those at the time points t31 to t32 described above, the description thereof will be omitted.

Then, as shown in FIG. 10, when the actual input shaft acceleration is smaller than the target input shaft acceleration in the inertia phase of the first shift, the actual value (actual value) of the input shaft rotation speed is larger than the target value. It becomes smaller. That is, since the degree of increase in the input shaft rotation speed is small, the progress of the first shift is slowed down. Therefore, the synchronous prediction time is larger than that in the case of FIG. 8 at the start of the inertia phase of the first shift, but the inclination (decrease degree) is almost the same as in the case of FIG. That is, the time from the time point t52 when the inertia phase of the first shift is started to the time point t53 when the synchronization prediction time reaches the predetermined value Th2 is longer than the time from the time point t32 to the time point t33 in FIG. The time from time to time t55 when the input shaft rotation speed reaches the synchronous rotation speed of the fifth shift stage is almost the same as the time from time point t33 to time point t35 in FIG.

Then, at the time point t53, when the synchronization prediction time reaches the predetermined value Th2 later than in the case of FIG. 8, the required torque for the first brake B1 which is the friction engaging element on the release side of the first shift becomes higher. Next, when the synchronous prediction time decreases as the first shift progresses and the synchronous predicted time reaches a predetermined value Th3 at the time point t54, a request for the first brake B1 which is a friction engaging element on the release side of the first shift is made. The torque is restored. Therefore, at the time point t55, the backlash of the output shaft 3b is clogged in the positive direction. That is, the increase correction of the release side clutch torque is started later than in the case of FIG. 8, and the stagnation of the progress of the shift can be suppressed by not lengthening the time for performing the increase correction more than necessary.

At time point t55, the torque phase of the first shift is started. That is, 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 clutch C1 which is the friction engagement element on the engagement side of the first shift is reduced. The required torque for is increased. At this time, since the first clutch C1 is engaged in a state where the backlash of the output shaft 3b is clogged on the positive direction side, it is possible to suppress the shock. Since the time points t56 to t58 are the same as the above-mentioned time points t5 to t7, the description thereof will be omitted.

As described above, in the present embodiment, the input shaft rotation speed is small by temporarily increasing the release side clutch torque of the first shift according to the synchronous predicted time at the time of power-on-downshift via the intermediate shift stage. Even in this case, it is possible to suppress the stagnation of the progress of the shift.

(Calculation of clutch torque)
Next, a method of calculating the clutch torque during shift control will be described with reference to FIG. The following flow is repeated at predetermined time intervals. Further, each step is executed by the ECU 5.

First, in step ST1 of FIG. 11, it is determined whether or not the power is on / downshift. For example, when the accelerator pedal is depressed to make a downshift determination, and when shift control based on the downshift determination is being executed, it is determined to be a power-on downshift. Then, when it is determined that the power-on-downshift is performed, the process proceeds to step ST2. On the other hand, if it is determined that the shift is not power-on-downshift (for example, in the case of other shift patterns such as power-on-upshift), the process proceeds to step ST7.

Next, in step ST2, it is determined whether or not to pass through the intermediate shift stage. Whether or not to pass through the intermediate gear is determined by, for example, whether or not the friction engaging element on the release side is switched when passing through the destination gear of the currently executed shift. Then, if it is determined that the vehicle passes through the intermediate shift stage, the process proceeds to step ST3. On the other hand, if it is determined not to pass through the intermediate shift stage, the process proceeds to step ST7.

Next, in step ST3, it is determined whether or not the element is a friction engaging element on the release side of the first shift. Then, when it is determined that the element is the friction engaging element on the release side of the first shift, the process proceeds to step ST4. On the other hand, if it is determined that the element is not the friction engagement element on the release side of the first shift (for example, the friction engagement element on the engagement side of the first shift), the process proceeds to step ST7.

Next, in step ST4, the synchronous predicted time until the input shaft rotation speed reaches the synchronous rotation speed of the intermediate shift stage is calculated. This synchronous prediction time is calculated, for example, by dividing the difference between the synchronous rotation speed of the intermediate shift stage and the current input shaft rotation speed by the amount of change in the difference per unit time. That is, the synchronous predicted time is the remaining time from the present time until the input shaft rotation speed reaches the synchronous rotation speed of the intermediate transmission stage. The input shaft rotation speed is calculated based on, for example, the detection result of the input shaft rotation speed sensor 82. Further, the synchronous rotation speed of the intermediate gear is calculated based on, for example, the gear ratio of the intermediate gear and the detection result of the output shaft rotation speed sensor 83.

Next, in step ST5, it is determined whether or not the synchronization prediction time is equal to or less than the predetermined value Th2. Then, when it is determined that the synchronization prediction time is equal to or less than the predetermined value Th2, the process proceeds to step ST6. On the other hand, if it is determined that the synchronization prediction time is not less than or equal to the predetermined value Th2 (when the synchronization prediction time is larger than the predetermined value Th2), the process proceeds to step ST7.

Next, in step ST6, it is determined whether or not the synchronization prediction time is equal to or less than the predetermined value Th3. Then, when it is determined that the synchronization prediction time is equal to or less than the predetermined value Th3, the process proceeds to step ST7. On the other hand, if it is determined that the synchronization prediction time is not less than or equal to the predetermined value Th3 (when the synchronization prediction time is larger than the predetermined value Th3), the process proceeds to step ST8.

Then, in step ST7, the torque capacitance of each friction engaging element is calculated using the gear train equation of motion. Then, the clutch torque calculated by using the gear train equation of motion is required for each friction engaging element. For example, in the friction engaging element on the release side of the first shift, when the synchronization prediction time is larger than the predetermined value Th2 (time points t31 to t33 in FIG. 8, time points t41 to t43 in FIG. 9 and time points t51 to t53 in FIG. 10). ), And when the synchronization prediction time is the predetermined value Th3 or less (time points t34 to t36 in FIG. 8, time points t44 to t46 in FIG. 9 and time points t54 to t56 in FIG. 10), the release side clutch is used using the gear train equation of motion. The torque is calculated. Then, the calculated release side clutch torque is required for the friction engagement element on the release side of the first shift.

Further, in step ST8, the torque capacity of the friction engaging element on the release side of the first shift is calculated using the gear train equation of motion and is increased and corrected. That is, the release side clutch torque calculated by using the gear train equation of motion is increased and corrected, and the increased and corrected value is required for the friction engaging element on the release side of the first shift. For example, in the friction engagement element on the release side of the first shift, when the synchronization prediction time is a predetermined value Th2 or less and larger than the predetermined value Th3 (time points t33 to t34 in FIG. 8 and time points t43 to t44 in FIG. 9). And at the time points t53 to t54) in FIG. 10, the release side clutch torque calculated by using the gear train equation of motion is increased and corrected. Then, the increased and corrected release side clutch torque is required for the friction engagement element on the release side of the first shift.

-Effect-
In the present embodiment, as described above, at the time of power-on-downshift in which the current shift is shifted to the target shift via the intermediate shift, the input shaft rotation speed becomes the synchronous rotation speed of the intermediate shift. The synchronization prediction time is calculated, and when the synchronization prediction time becomes a predetermined value Th2 or less, the torque capacity of the friction engaging element on the release side is increased. With this configuration, even if the actual input shaft acceleration varies with respect to the target input shaft acceleration, the torque capacity of the friction engaging element on the release side can be increased at an appropriate timing. It is possible to suppress the stagnation of the progress of the shift while suppressing the shock when passing through the shift stage.

Further, in the present embodiment, when the synchronization prediction time becomes the predetermined value Th3 or less, the increase correction of the torque capacity of the friction engaging element on the release side is completed, so that the release side is released at the start of the torque phase of the first shift. Since it is possible to suppress the clutch torque from becoming excessive, it is possible to suppress the delay in increasing the driving force due to the delay in the progress of the torque phase.

-Other embodiments-
It should be noted that the embodiment disclosed this time is an example in all respects and does not serve as a basis for a limited interpretation. Therefore, the technical scope of the present invention is not construed solely by the above-described embodiments, but is defined based on the description of the claims. In addition, the technical scope of the present invention includes all modifications within the meaning and scope equivalent to the claims.

For example, in the present embodiment, the example in which the vehicle 100 is FF is shown, but the present invention is not limited to this, and the vehicle may be FR (front engine / rear drive) or four-wheel drive. ..

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

Further, in the present embodiment, an example is shown in which the correction for increasing the torque capacity of the friction engaging element on the release side is terminated when the synchronization prediction time becomes the predetermined value Th3 or less, but the present invention is not limited to this, and the release side is not limited to this. The increase correction may be ended when a predetermined time has elapsed from the start of the increase correction of the torque capacity of the friction engaging element of the above. Further, the increase correction may be ended when the torque phase of the first shift is started.

Further, in the present embodiment, an example is shown in which the intermediate shift stage is set because it is necessary to re-grasp the four elements, but the present invention is not limited to this, and even if the two-element re-grasping is sufficient, friction Intermediate gears may be set to reduce the frictional load on the engaging elements.

Further, in the present embodiment, 8-5-3 shift is shown as an example of shifting from the current shift to the target shift via the intermediate shift, but the present invention is not limited to this, and other patterns are available. You may.

Further, in the present embodiment, an example in which the target shift stage is set based on the shift map is shown, but the target shift is not limited to this, and the shift stage that can be realized in consideration of the failure state, the amount of heat generated, the overrev, etc. is the target shift. It may be set as a stage.

Further, in the present embodiment, the ECU 5 may be composed of a plurality of ECUs.

The present invention can be used in an automatic transmission control device that controls an automatic transmission that establishes a plurality of gears by selectively engaging a plurality of friction engaging elements.

3 Automatic transmission 5 ECU (control device for automatic transmission)
C1 1st clutch (friction engagement element)
C2 2nd clutch (friction engagement element)
C3 3rd clutch (friction engagement element)
C4 4th clutch (friction engagement element)
B1 1st brake (friction engagement element)
B2 2nd brake (friction engagement element)

Claims (1)

An automatic transmission control device applied to an automatic transmission that establishes a plurality of gears by selectively engaging a plurality of friction engaging elements.
At the time of power-on-downshift to shift from the current shift to the target shift via the intermediate shift, the predicted time until the input shaft rotation speed becomes the synchronous rotation speed of the intermediate shift is calculated, and the predicted time is calculated. Is less than or equal to the first predetermined value, the torque capacity of the friction engaging element on the release side is corrected to increase , and the predicted time is less than or equal to the second predetermined value, which is smaller than the first predetermined value. a control device for an automatic transmission, wherein the increase correction of the torque capacity of the frictional engagement element of the release side is configured so that the ends.

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