JP2013024401A - Drive force control device of vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a drive force control device of a vehicle that can obtain a drive force requested by the vehicle and reducing a shock by shifting.SOLUTION: This drive force control device of a vehicle includes a power transmission device for transmitting power by engaging a one-way clutch for transmitting only a power in one rotating direction with a clutch the engagement and the disengagement of which can be selectively changed. When the one-way clutch is changed from the condition in which the clutch does not transmit power to the condition in which the clutch transmits power, the clutch is constituted to be slipped. The drive force control device further includes a first torque calculation means for calculating the control amounts of a torque input into the power transmission device and the torque capacity of the clutch based on the angular acceleration of the output shaft of the power transmission device and a target slip amount through which the clutch is slipped.

Description

  The present invention relates to a driving force of a vehicle including a power transmission device having a clutch capable of selectively changing engagement and release of a power transmission member and a one-way clutch that transmits only power in one rotational direction. More particularly, the present invention relates to a device that controls the driving force of a power transmission device that is configured such that both a clutch and a one-way clutch are engaged to transmit power.

  The one-way clutch does not require hydraulic pressure or electric power to transmit power, and does not require control for engaging the clutch, and thus is used for a power transmission device in a vehicle. In Patent Document 1, a one-way clutch is provided between the engine and the automatic transmission, and there is an acceleration request from a state where the vehicle is coasting. The rotation speed of the input shaft on the engine side of the one-way clutch is automatically A device is described that is configured such that, when synchronized with the rotational speed of the output shaft on the transmission side, the one-way clutch is engaged and the power of the engine is transmitted to the automatic transmission. In the apparatus configured as described above, when the one-way clutch is engaged after the vehicle starts to accelerate, the power of the engine is rapidly input to the automatic transmission. For this reason, when the driving force of the vehicle rapidly increases, that is, when the acceleration increases rapidly, a shock occurs in the vehicle. Therefore, the apparatus described in Patent Document 1 reduces the output of the engine by ignition timing delay control when the vehicle is requested to accelerate, and reduces the fastening force of the hydraulic clutch provided in the automatic transmission. By doing so, it is configured to suppress a sudden shift shock when the one-way clutch is engaged.

  Further, Patent Document 2 discloses an automatic transmission configured such that a one-way clutch and a hydraulic clutch are engaged to select a gear position. The automatic transmission described in Patent Document 2 outputs the engine torque according to the acceleration request when there is an acceleration request and the shift stage is selected, so that the rotational speed on the input side of the one-way clutch is The time for synchronizing the output side rotation speed is configured to be shortened. In addition, in order to suppress a shock caused by the engagement of the one-way clutch, the engine output torque is reduced before the one-way clutch is engaged, so that even if the engine output torque is input, it does not slip and is low in pressure. The hydraulic pressure of the hydraulic clutch is set so as to be engaged with. Further, when the one-way clutch is engaged, the engine torque is increased. In this case, the hydraulic clutch slips by increasing the time for increasing the engine torque rather than the time for increasing the torque capacity of the hydraulic clutch. In this way, it is configured to suppress a shock caused by a shift.

  Further, Patent Document 3 describes a power transmission device including a sub-transmission and a main transmission having a forward / reverse switching function between an engine and a toroidal-type continuously variable transmission. A one-way clutch is provided on the output shaft of the sub-transmission, and a hydraulic clutch that is engaged when the vehicle travels forward is provided on the main transmission. When the accelerator opening is equal to or greater than a predetermined value, the control duty of the hydraulic clutch is reduced in order to suppress a shift shock when the one-way clutch is engaged. That is, it is configured to suppress a rapid torque fluctuation when the one-way clutch is engaged by slipping the hydraulic clutch. Further, the target slip time for slipping the hydraulic clutch and the slip rotation speed are set from the gear ratio deviation amount and the accelerator opening difference.

  In Patent Document 4, the clutch output torque is reduced by lowering the hydraulic pressure of the engaged hydraulic clutch at the time of clutch-to-clutch shifting. The described apparatus is described.

JP-A-10-324178 JP 2010-242926 A JP 2006-348984 A JP 2007-2899 A

  As described above, by changing the gear ratio, the driving force of the vehicle before and after the gear shift changes abruptly. Therefore, as described in each patent document, it is engaged when setting the gear ratio after the gear shift. By lowering the clutch engagement pressure and slipping the clutch, the torque transmitted from the engine to the output shaft of the vehicle can be reduced, and a sudden change in the driving force of the vehicle can be suppressed. . That is, the torque transmitted to the output shaft can be reduced by the relationship between the engine output torque and the clutch engagement pressure. Therefore, it is possible to suppress a sudden change in the driving force of the vehicle by reducing the output torque of the engine.

  On the other hand, if the torque of the engine is reduced or the engagement pressure of the clutch is reduced, the driving force required for the vehicle may not be obtained. Therefore, conventionally, the engine torque and the clutch engagement pressure are set based on a map prepared in advance by experiments or the like so that the driving force does not change abruptly and the required driving force can be obtained. It was. Therefore, it is necessary to prepare a map based on various parameters such as required driving force, engine torque, or clutch engagement pressure, and there is a possibility that man-hours for preparing the map may increase.

  The present invention has been made against the background described above, and provides a driving force control device for a vehicle that can obtain a driving force required for the vehicle and can reduce a shock caused by a shift. It is the purpose.

  In order to achieve the above object, the invention of claim 1 is to engage each of a one-way clutch that transmits only power in one direction in the rotational direction and a clutch that can selectively switch engagement and disengagement. In a vehicle driving force control device configured to slip the clutch when the one-way clutch is switched from a state in which power is not transmitted to a state in which power is transmitted. The control amount of the torque input to the power transmission device and the torque capacity of the clutch is calculated based on the target angular acceleration of the output shaft of the power transmission device and the target slip amount for slipping the clutch. A torque calculation means is provided.

  According to a second aspect of the present invention, in the first aspect of the present invention, the first torque calculation means is configured such that the clutch is completely engaged when the angular acceleration of the output shaft of the power transmission device reaches the target angular acceleration. Thus, the vehicle driving force control device includes means for calculating a control amount between the torque input to the power transmission device and the torque capacity of the clutch.

  According to a third aspect of the present invention, in the first or second aspect of the present invention, the torque input to the power transmission device and the clutch so that the rate of change of the target slip amount before and after the clutch is completely engaged is small. A second torque calculating means for calculating a control amount with respect to the torque capacity of the first torque calculating means; and a torque input by the first torque calculating means before the clutch is completely engaged; The torque input to the power transmission device is switched from the control amount with the torque capacity of the clutch to the control amount between the torque input to the power transmission device calculated by the second torque calculating means and the torque capacity of the clutch. And a torque capacity of the clutch are output.

  According to a fourth aspect of the present invention, in the third aspect of the present invention, the second torque calculating means is configured such that when the clutch is completely engaged, the control amount of the torque input to the power transmission device is the target angle. A driving force control apparatus for a vehicle, comprising means for calculating a control amount of the torque capacity of the clutch according to the calculated control amount, so as to obtain a torque corresponding to the acceleration.

  According to a fifth aspect of the present invention, in any one of the first to fourth aspects, the apparatus further comprises slip amount detecting means for detecting the slip amount of the clutch, and the torque input to the power transmission device, the torque capacity of the clutch, And a deviation between an actual slip amount detected by the slip amount detecting means with a delayed response between the torque input to the power transmission device and the torque capacity of the clutch. Thus, the vehicle driving force control device is configured to feedback control the control amount of either the torque input to the power transmission device or the torque capacity of the clutch.

  According to the first aspect of the present invention, the clutch is configured to slip when the one-way clutch is switched from the state where the power is not transmitted to the state where the power is transmitted. The generated pulse-shaped shift shock can be suppressed or prevented. The control amount between the torque input to the power transmission device and the torque capacity of the clutch is calculated based on the angular acceleration of the output shaft of the power transmission device and the target slip amount for slipping the clutch. Therefore, the man-hour for preparing the map of the torque input into the power transmission device and the torque capacity of the clutch can be reduced.

  According to the invention of claim 2, the torque input to the power transmission device and the torque capacity of the clutch are set such that the clutch is completely engaged when the angular acceleration of the output shaft of the power transmission device reaches the target angular acceleration. Therefore, it is possible to suppress or prevent the clutch from slipping when the angular acceleration of the output shaft of the power transmission device reaches the target angular acceleration. That is, it is possible to suppress or prevent an excessive increase in the rotational speed of the power input to the power transmission device. Therefore, it is possible to suppress or prevent power loss caused by excessively increasing the rotational speed.

  According to the invention of claim 3, before the clutch is completely engaged, the torque input to the power transmission device and the clutch are reduced so that the rate of change of the target slip amount before and after the clutch is completely engaged becomes small. Since the control amount with the torque capacity is calculated, and the torque input to the power transmission device and the torque capacity of the clutch are output by switching from the control amount calculated before that, the pulse-like speed change caused by the engagement of the clutch Shock can be suppressed or prevented.

  According to the invention of claim 4, when the clutch is completely engaged, the control amount of the torque input to the power transmission device is calculated so as to be a torque corresponding to the target angular acceleration, and the calculated Since the control amount of the torque capacity of the clutch according to the control amount is calculated, there is no need to perform control to increase or decrease the torque input to the power transmission device after the clutch is engaged.

  According to the invention of claim 5, further comprising a slip amount detecting means for detecting the slip amount of the clutch, the slip amount calculated based on a control amount between the torque input to the power transmission device and the torque capacity of the clutch, Based on the deviation between the actual slip amount detected by the slip amount detection means due to a delay in the response between the torque input to the power transmission device and the torque capacity of the clutch, the torque input to the power transmission device and the torque capacity of the clutch Any one of the control amounts is feedback-controlled. Therefore, control can be performed in consideration of the response input delay of the torque input to the power transmission device and the torque capacity of the clutch, and as a result, the controllability of the slip amount of the clutch can be improved.

It is a time chart which shows the change of the control amount of the turbine torque by the vehicle driving force control apparatus which concerns on this invention, and the torque capacity of a 1st clutch, and the change of motive powers, such as each rotating member. It is a schematic diagram which shows the structure of the drive system and control system of the vehicle which can be made into object by this invention. FIG. 3 is a skeleton diagram of the torque converter and the automatic transmission in FIG. 2. FIG. 4 is a collinear diagram of the automatic transmission in FIG. 3. It is a chart which shows the operation table. It is a flowchart for demonstrating the example of control for suppressing the shift shock by a one-way clutch engaging. It is a flowchart for demonstrating the example of control for suppressing the shift shock by a 1st clutch engaging. It is a flowchart for demonstrating feedback control. It is a time chart which shows delay of engine torque and torque capacity of the 1st clutch.

  Next, the present invention will be described based on specific examples. FIG. 2 is a schematic diagram for explaining the configuration of the drive system and the control system of the vehicle Ve that can be the subject of the present invention. First, a power source (hereinafter simply referred to as the engine 1) configured by an internal combustion engine such as a gasoline engine, a diesel engine, or an LPG engine, or an electric motor is provided. The engine 1 is connected to an automatic transmission 3 via a torque converter 2 so that the power output from the engine 1 can be increased or decreased. Drive wheels 7 are connected to the output side of the automatic transmission 3 via a propeller shaft 4, a differential gear 5, and a drive shaft 6.

  FIG. 3 is a skeleton diagram for explaining the configurations of the torque converter 2 and the automatic transmission 3. The torque converter 2 shown in the figure is disposed so as to face the pump impeller 2a, which rotates integrally with the output shaft 1a of the engine 1 and generates a flow of oil that is liquid-tightly flowed therein. The turbine runner 2b that is driven by the oil flow generated by the pump impeller 2a and outputs power, and is arranged between the pump impeller 2a and the turbine runner 2b, and rectifies the flow of oil discharged from the turbine runner 2b. And a stator 2c that produces a torque amplifying action by being reduced to the pump impeller 2a. Furthermore, a lockup clutch 2d is provided for transmitting power by directly connecting the turbine runner 2b and the output shaft 1a of the engine 1 when a speed change action is not required. The lockup clutch 2d is configured to be able to switch between a state in which power is transmitted and a state in which the power is shut off in accordance with hydraulic pressure and electric power.

  The automatic transmission 3 is connected to the output shaft 2e of the torque converter 2. The automatic transmission 3 shown in the figure includes two sets of planetary gear mechanisms, that is, a single pinion type planetary gear mechanism 8 and a Ravigneaux type planetary gear mechanism 9. Each of the planetary gear mechanisms 8 and 9 includes a clutch capable of connecting or disconnecting the rotating elements, a brake capable of selectively inhibiting (fixing) rotation of any of the rotating elements, and one-way. A fastening device such as a one-way clutch that allows only power transmission is provided.

  The configuration of the automatic transmission 3 shown in FIG. 3 will be described more specifically. The automatic transmission 3 shown in the figure is arranged in the order of a torque converter 2, a single pinion type planetary gear mechanism 8, and a Ravigneaux type planetary gear mechanism 9. The single pinion type planetary gear mechanism 8 meshes with a sun gear S10 that rotates integrally with the output shaft 2e of the turbine runner 2b, a ring gear R10 that is arranged coaxially with the sun gear S10, and the sun gear S10 and the ring gear R10. And a plurality of pinion gears held by the carrier C10 so as to rotate and revolve freely. The Ravigneaux type planetary gear mechanism 9 includes a sun gear S20 disposed coaxially with the rotation axis of the output shaft 2e of the turbine runner 2b, a ring gear R20 disposed coaxially with the sun gear S20, and the sun gear S20. A plurality of first pinion gears meshed with the ring gear R20 and longer than the shaft lengths of the sun gear S20 and the ring gear R20, and the carrier C10 constituting the single pinion type planetary gear mechanism 8 are rotated integrally with the turbine runner. A sun gear S30 disposed coaxially with the output shaft 2e of 2b, a plurality of second pinion gears meshed with the first pinion gear and the sun gear S30, and a carrier C20 that holds the first and second pinion gears in a freely rotating and revolving manner. And is composed of. The carrier C20 of the Ravigneaux type planetary gear mechanism 9 functions as an output shaft and is connected to the propeller shaft 4.

  Further, the automatic transmission 3 shown in the figure includes a first clutch C1 that can selectively transmit or cut power from the output shaft 2e of the turbine runner 2b to the sun gear S20, and a ring gear R20 from the output shaft 2e of the turbine runner 2b. And a second clutch C2 capable of selectively transmitting or interrupting power to a carrier C10 constituting a single pinion type planetary gear mechanism 8 by engagement, that is, a sun gear constituting a Ravigneaux type planetary gear mechanism 9 Rotation of the ring gear R10 that constitutes the single pinion planetary gear mechanism 8 by engaging the first brake B1 that prohibits the rotation of S30 and the second brake B2 that prohibits the rotation of the ring gear R20 by engaging. The third brake B3 prohibiting the rotation and the ring gear R20 rotated in one direction To release the ring gear R20 to focus idly rotating the ring gear R20, and a one-way clutch F1 which is configured to engage to prohibit the rotation of the ring gear R20 when rotated in the other direction. Each of the clutches C1, C2 and the brakes B1, B2, B3 except for the one-way clutch F1 described above may be a hydraulic engagement mechanism or an electric engagement mechanism. That is, each clutch C1, C2 and brake B1, B2, B3 may be any engagement mechanism that can control the engaged state and the disconnected state. Furthermore, each clutch C1, C2 should just be what can control engagement force. In the following description, a case where a so-called hydraulic fastening device in which each clutch C1, C2 and brakes B1, B2, B3 are engaged or released by hydraulic pressure is used will be described as an example.

  And the sensor 10 which detects the rotation speed of the output shaft 2e of the turbine runner 2b, the sensor 11 which detects the rotation speed of the output shaft of the automatic transmission 3, that is, the carrier C20, the accelerator operation by the driver, that is, the accelerator opening degree. An electronic control unit (ECU) that includes an accelerator opening sensor 12 to detect, and outputs a signal for controlling the engine torque and the hydraulic pressure of each fastening device based on signals input from the sensors 10, 11, 12. 13 is provided.

  Next, the operation of the automatic transmission 3 described above will be briefly described based on the alignment chart shown in FIG. 4 and the operation table shown in FIG. The automatic transmission 3 shown in the figure is a transmission capable of setting four shift speeds. First, when setting the first gear, the first clutch C1 and the one-way clutch F1 are engaged as shown in the operation table of FIG. That is, when the first clutch C1 is engaged, the sun gear S20 functions as an input element, and the one-way clutch F1 is engaged when the rotational speed of the output shaft of the automatic transmission 3, that is, the vehicle speed is equal to or less than a constant speed. In other words, when the vehicle Ve is in a low speed or stopped state and the acceleration is requested, if the first clutch C1 is engaged, the one-way clutch F1 is necessarily engaged, and the one-way clutch F1 is counteracted. It functions as a force element and power is output from the carrier C20. When the engine brake is applied, that is, when the first speed is selected by the shift lever, the second brake B2 is engaged and the first speed is forcibly selected.

  The second speed is set by engaging the first clutch C1 and the first brake B1. That is, since the rotation speed of the ring gear R20 increases as the rotation speed of the carrier C20 increases while the first speed is set, the one-way clutch F1 is released and the first brake B1 is engaged. As a result, the second speed is set. Further, when the vehicle speed increases, the first clutch C1 is engaged, the first brake B1 is released, the second clutch C2 is engaged, the sun gear S20 and the ring gear R20 are rotated together, and Ravigneo The third speed is set by bringing the type planetary gear mechanism 9 into a directly connected state.

  Furthermore, when the vehicle speed increases, the first brake B1 is engaged and the fourth speed is set while releasing the first clutch while the second clutch C2 is engaged. That is, since power is transmitted from the output shaft 2e of the torque converter 2 to the ring gear R20, the ring gear R20 functions as an input element, and the sun gear S30 functions as a reaction force element whose rotation is prohibited by the first brake B1. C20 functions as an output element to set the fourth speed. When the first to fourth speeds are selected, the ring gear R10 in the single pinion type planetary gear mechanism 8 is in an idle state. Then, when reverse travel is selected by the shift lever, the second and third brakes B2 and B3 are engaged. Then, power is transmitted to the sun gear S30 in the Ravigneaux planetary gear mechanism 9 via the carrier C10 in the single pinion planetary gear mechanism 8, and the carrier C20 is reversed by the engagement of the second brake B2. Since the vehicle rotates, the vehicle Ve can travel backward. That is, the single pinion type planetary gear mechanism 8 is configured to function as a forward / reverse switching device.

  In the power transmission device having the one-way clutch F1 as described above, the shift shock occurs due to the occurrence of pulse-like torque fluctuations when the one-way clutch F1 is engaged from the released state. It is for suppressing that. FIG. 6 shows a flowchart for explaining the control example. In the flowchart shown in FIG. 6, it is assumed that the vehicle Ve is traveling at low speed and coasting. Specifically, it is assumed that the vehicle is traveling in a state where the second speed or the third speed is set and the accelerator opening is OFF. First, it is determined whether or not the accelerator opening is changed from OFF to ON (step S11). This determination can be made based on whether or not the accelerator is depressed by the driver by the accelerator opening sensor 12. If the determination is negative, the control is temporarily terminated as it is. On the other hand, if the accelerator is depressed and an affirmative determination is made in step S11, the hydraulic pressure is reduced so that the torque capacity of the first clutch C1 becomes 0 (zero) Nm (step S12). That is, the first clutch C1 is released. This is because the output shaft 2e of the turbine runner 2b, that is, the input shaft of the automatic transmission 3 (hereinafter, the input shaft of the automatic transmission 3 is simply referred to as the input shaft 3a) with only the first clutch C1 engaged. The one-way clutch F1 is engaged due to the relationship between the input shaft speed and the carrier C20 speed (hereinafter, the carrier C20 speed is referred to as the output shaft speed). This is to prevent a shift shock from occurring.

  Next, it is determined whether or not the input shaft rotational speed matches the rotational speed obtained by integrating the output shaft rotational speed and the speed ratio after the shift, that is, the speed ratio when the first speed is set (step S13). . That is, it is determined whether or not the rotation speed at which the one-way clutch F1 is engaged by engaging the first clutch C1, that is, the rotation speed at which the one-way clutch F1 is synchronized. This can be determined based on signals detected by the sensors 10 and 11 that detect the input shaft rotation speed and the output shaft rotation speed. In the state where the first clutch C1 is disengaged, even if the input shaft speed is increased, the power is not transmitted to the output shaft, so the output shaft speed is kept constant. If a negative determination is made in step S13, the determination in step S13 is repeated until the input shaft rotational speed matches the rotational speed obtained by integrating the output shaft rotational speed and the gear ratio after shifting. On the contrary, if the determination in step S13 is affirmative, that is, if the input shaft rotational speed matches the rotational speed obtained by integrating the output shaft rotational speed and the speed ratio after the shift, the acceleration required for the vehicle Ve. (DWOTGT) and a time required to reach the acceleration DWOTGT (hereinafter simply referred to as a first predetermined time TC1SLP1) and a turbine blowing amount (WTSLPTGT) are determined (step S14). The acceleration DWOTGT in step S14 can be determined from the value detected by the accelerator opening sensor 12. Further, the first predetermined time TC1SLP1 can be arbitrarily determined for each vehicle. In other words, the time TC1SLP1 can be shortened in the case of a sports type vehicle, and the time TC1SLP1 can be arbitrarily increased in the case of a family type vehicle. The time TC1SLP1 may be changed based on the accelerator depression speed or the like. Further, the turbine blow amount WTSLPTGT is proportional to the slip amount of the first clutch C1, and specifically, in the case where the first speed is set with the first clutch C1 completely engaged. The number of rotations is faster than the number of rotations of the input shaft. That is, when the first clutch C1 slips, the reaction force against the rotation of the input shaft 3a is weak, and as a result, the rotation speed of the input shaft 3a increases. The increased amount corresponds to the turbine blow amount WTSLPTGT.

なお、上式におけるＴＴはタービントルク、ＴＣ１は第１クラッチＣ１のトルク容量、tは経過時間である。また、この発明は、タービントルクＴＴと第１クラッチＣ１のトルク容量ＴＣ１とを演算式によって定めるものであり、それらタービントルクＴＴおよび第１クラッチＣ１のトルク容量ＴＣ１を簡易的な関数で設定するために、上式のように一次関数としている。 Based on the acceleration DWOTGT determined in step S14, the first predetermined time TC1SLP1, and the turbine blow amount WTSLPTGT, coefficients a and b and an initial value TT ₀ of the turbine torque in the following formula are calculated (step S15). That is, the control amount of the turbine torque and the torque capacity of the first clutch is calculated. The turbine torque corresponds to “torque input to the automatic transmission” in the present invention, and step S15 corresponds to the first torque calculating means in the present invention. That is, the “torque input to the automatic transmission” in the present invention is not limited to the configuration of the power source, but may be any torque input to the transmission provided with the one-way clutch F1.

In the above equation, TT is the turbine torque, TC1 is the torque capacity of the first clutch C1, and t is the elapsed time. In the present invention, the turbine torque TT and the torque capacity TC1 of the first clutch C1 are determined by an arithmetic expression, and the turbine torque TT and the torque capacity TC1 of the first clutch C1 are set by a simple function. In addition, it is a linear function as in the above equation.

なお、上式におけるωtslpはタービン吹き量、ωoは出力軸回転数である。また、α、β、γは車両Ｖｅおよび自動変速機３の諸元から決まる係数である。具体的には、各軸に作用する慣性モーメントやギヤの歯数から算出することができる係数である。 Here, the calculation in step S15 will be specifically described. First, from the equation of motion inside the automatic transmission 3, the following equation is established when the first clutch C1 is slipping.

In the above equation, ωtslp is the turbine blow amount, and ωo is the output shaft speed. Α, β, and γ are coefficients determined from the specifications of the vehicle Ve and the automatic transmission 3. Specifically, it is a coefficient that can be calculated from the moment of inertia acting on each axis and the number of gear teeth.

From the above formulas, the turbine blow amount ωtslp and the angular acceleration dωo of the output shaft can be a function of time t shown below. That is, the function of the turbine blow amount ωtslp can be determined by substituting Equation (1) into Equation (3) and integrating, and by substituting Equation (2) into Equation (4), the angular acceleration of the output shaft can be determined. dωo can be determined.

すなわち、出力軸の角加速度ｄωoが目標加速度に到達する際もしくは同時にタービン吹き量ωtslpが０（ゼロ）に収束し、かつ第１所定時間ＴＣ１ＳＬＰ１の半分でタービン吹き量ωtslpが最大となるようにタービン吹き量ωtslpの目標値を設定する。 It can be seen from Equation (5) that the function indicating the turbine blow amount ωtslp is a quadratic function. Therefore, target values for the turbine blow amount ωtslp and the angular acceleration dωo of the output shaft are set as follows.

That is, when the angular acceleration dωo of the output shaft reaches the target acceleration or at the same time, the turbine blow amount ωtslp converges to 0 (zero), and the turbine blow amount ωtslp is maximized in half the first predetermined time TC1SLP1. Set the target value for the blow rate ωtslp.

ついで、式（５）および式（７）からタービントルクＴＴの初期値ＴＴ_０の演算式を求め、式（５）および式（８）から係数ａの演算式を求める。

そして、式（１１）および式（１２）に、式（１０）で算出した係数ｂを代入し、かつ式（１１）および式（１２）から、係数ａおよびタービントルクＴＴの初期値ＴＴ_０を求める。 Then, the coefficients a and b and the initial value TT ₀ of the turbine torque TT are obtained from the above-described equations. Specifically, first, the coefficient b can be obtained from the equations (6) and (9).

Next, an arithmetic expression of the initial value TT ₀ of the turbine torque TT is obtained from the expressions (5) and (7), and an arithmetic expression for the coefficient a is obtained from the expressions (5) and (8).

Then, the coefficient b calculated by the expression (10) is substituted into the expressions (11) and (12), and the coefficient a and the initial value TT ₀ of the turbine torque TT are obtained from the expressions (11) and (12). Ask.

Then, by substituting the coefficients a and b calculated as described above and the initial value TT ₀ of the turbine torque TT into the equations (1) and (2), the turbine torque TT and the torque capacity TC1 of the first clutch C1 are obtained. The turbine torque TT and the torque capacity TC1 of the first clutch C1 corresponding to the elapsed time in the function are output (step S16). Next, it is determined whether or not the slip of the first clutch C1 has converged (step S17). That is, it is determined whether or not the turbine blow amount ωtslp is 0 (zero). If the determination in step S17 is affirmative, it means that the first clutch C1 is engaged and the first speed is set, so this control is temporarily terminated.

  On the other hand, if a negative determination is made in step S17, it is determined whether or not the time has elapsed since the first predetermined time TC1SLP1 (step S18). If a negative determination is made in step S18, the process returns to step S16 and the turbine torque TT and the torque capacity TC1 of the first clutch C1 corresponding to the elapsed time are output again. On the other hand, if a positive determination is made in step S18, since the slip of the first clutch C1 has not converged by the assumed control, backup control such as increasing the hydraulic pressure is executed (step S19). That is, control is performed so that the first clutch C1 is forcibly engaged. Then, it is determined again whether or not the slip has converged by performing the backup control (step S20). If a negative determination is made in step S20, the process returns to step S19 to execute backup control again. On the other hand, if a positive determination is made in step S20, this control is temporarily terminated.

  As described above, the driving force control apparatus according to the present invention can calculate the turbine torque TT and the torque capacity TC1 of the first clutch C1 from the acceleration required for the vehicle Ve and the turbine blow amount ωtslp. Therefore, it is not necessary to prepare a map in advance for determining the turbine torque TT and the torque capacity TC1 of the first clutch C1, and the number of steps for determining the map can be reduced. Furthermore, since the one-way clutch F1 is engaged in a state where the torque capacity TC1 of the first clutch C1 is reduced, even if torque fluctuation temporarily occurs due to the engagement of the one-way clutch F1, the first clutch C1 It is possible to suppress or prevent the change in the acceleration of the vehicle Ve, that is, the occurrence of a pulse-like shift shock due to the slippage.

On the other hand, in the control example described above, the first clutch C1 passes after the first predetermined time TC1SLP1 has elapsed.
When the engagement is completed, a pulse-shaped shift shock may occur. Therefore, the driving force control apparatus according to the present invention performs smoothing control before the engagement of the first clutch is completed as shown in the flowchart of FIG. 7 in order to further suppress or prevent the shift shock. The control example shown in FIG. 7 is executed after step S17 of the control example shown in FIG. 6. Therefore, in FIG. 7, steps S11 to S15 shown in FIG. 6 are omitted. If a negative determination is made in step S17, it is determined whether or not a time has elapsed since the first predetermined time TC1SLP1 minus a certain time T1 (step S21). The fixed time T1 in step S21 is a time for executing the annealing control, and is, for example, about one-eighth of the first predetermined time TC1SLP1. The fixed time T1 may be a fixed value, or may be a time arbitrarily set according to the target acceleration DWOTGT, the vehicle speed, or the like. If a negative determination is made in step S21, the process shown in FIG. 6 is being executed, so the process returns to step S16 and again the turbine torque TT and the first clutch of the controlled variable corresponding to the elapsed time. The torque capacity TC1 of C1 is output. On the other hand, if a positive determination is made in step S21, a function of the turbine blow amount ωtslp is determined so that the turbine blow amount ωtslp becomes 0 (zero) in the second predetermined time TC1SLP2 (step S22). Specifically, a function of the turbine blow amount ωtslp is set to a parabola, that is, a quadratic function, and a coefficient P in the quadratic function is calculated. The function in step S22 is not limited to a quadratic function, but may be any function determined so that the turbine blow amount ωtslp converges to 0 (zero).

Step S22 will be specifically described. First, the fixed time T1 is set to 1/8 of the first predetermined time TC1SLP1. Therefore, the initial value of the turbine blow amount ωtslp in step S22 can be calculated from the above equations (5) and (11).

On the other hand, since the function determined in step S22 is a quadratic function in which the turbine blow amount ωtslp becomes 0 (zero) in the second predetermined time TC1SLP2, it can be assumed to be the following expression.

Therefore, the coefficient P of the quadratic function can be calculated from the equations (13) and (14), and the function in step S22 can be determined as the following equation.

Then, a function of the turbine torque TT and the torque capacity TC1 of the first clutch C1, that is, a control amount is determined based on the function determined in step S22 (step S23). Step S23 will be specifically described. First, functions of the turbine torque TT and the torque capacity TC1 of the first clutch C1 are set in the same manner as in the expressions (1) and (2). Step S23 corresponds to the second torque calculation means in this invention.

Further, the turbine blow amount ωtslp is a function shown below from the equation of motion. Specifically, as shown in Equation (3), the turbine blow amount ωtslp is obtained from the turbine torque TT and the torque capacity TC1 of the first clutch C1, in other words, the turbine blow amount ωtslp is determined by the turbine torque TT and the first clutch C1 in Equation (3). Is a function of integrating the torque capacity TC1. The turbine torque TT and the torque capacity TC1 of the first clutch C1 are the above formula (17) and formula (18). Therefore, the function of the turbine blow amount ωtslp after the time obtained by subtracting the predetermined time T1 from the first predetermined time TC1SLP1 is expressed by the following equation.

なお、式（１８）におけるｂ×（ＴＣ１ＳＬＰ１−Ｔ１）は、ワンウェイクラッチＦ１の変速ショックを抑制する制御における第１クラッチＣ１のトルク容量ＴＣ１の関数により求めることができる。したがって、式（２０）では、ＴＣ１_１と示している。 On the other hand, as shown in Expression (14) or Expression (16), the turbine blow amount ωtslp becomes 0 (zero) in the second predetermined time TC1SLP2. Therefore, the final term in equation (19) is:

Note that b × (TC1SLP1-T1) in the equation (18) can be obtained by a function of the torque capacity TC1 of the first clutch C1 in the control for suppressing the shift shock of the one-way clutch F1. Therefore, we show in Formula (20), TC1 ₁ and.

となる。 Here, when the coefficients of the respective terms in Expression (16) and Expression (19) are compared,

It becomes.

の関係が成り立つ。 That is, from equation (21)

The relationship holds.

Therefore, by substituting Equation (24) into Equation (23) and arranging it, the initial value TT ₁ of the turbine torque TT in the control for converging the turbine blow amount ωtslp can be calculated.

Here, the turbine torque TT is returned to the torque TTfwd determined from the accelerator opening from the time when switching to the control for converging the turbine blowing amount ωtslp to the lapse of the second predetermined time TC1SLP2, that is, the first When the clutch C1 is completely engaged, the rate of change of the turbine torque TT is set so that the control amount of the turbine torque TT becomes the torque TTfwd determined from the accelerator opening, that is, the torque according to the target acceleration. calculate. That is, the coefficient a2 is as follows.

Further, from the equation (21), the rate of change of the torque capacity TC1 of the first clutch C1, that is, the coefficient b2 is as follows. That is, the coefficient b2 is a value corresponding to the coefficient a2, and the control amount of the torque capacity TC1 of the first clutch C1 is determined according to the control amount of the turbine torque TT.

  Therefore, the functions of the turbine torque TT and the torque capacity TC1 of the first clutch C1 after switching to the control for converging the turbine blowing amount ωtslp can be obtained from the equations (25) to (27).

  And based on the function of the turbine torque TT calculated | required by step S23, and the torque capacity TC1 of the 1st clutch C1, the turbine torque TT and the torque capacity TC1 of the 1st clutch C1 are output (step S24). Next, it is determined again whether or not the slip of the first clutch C1 has converged (step S25). When a negative determination is made in step S25, the process returns to step S24, and the turbine torque TT and the torque capacity TC1 of the first clutch C1 are output again. On the contrary, if a positive determination is made in step S25, this control is temporarily terminated.

  As described above, by converging the slip of the first clutch C1, it is possible to suppress or prevent the occurrence of pulse-like torque fluctuations, that is, shift shocks caused by the engagement of the first clutch C1.

  Here, the turbine torque TT and the torque capacity TC1 of the first clutch C1 when the control for suppressing the shift shock due to the engagement of the one-way clutch F1 and the shift shock due to the engagement of the first clutch C1 are executed. A change in the control amount and a change in the number of revolutions of each member by performing the control will be described based on the time chart shown in FIG. In the graph showing changes in the control amount of the turbine torque TT and the torque capacity TC1 of the first clutch C1, the solid line indicates the control amount of the turbine torque TT, and the broken line indicates the control amount of the torque capacity TC1 of the first clutch C1. The time chart shown in FIG. 1 is based on the premise that the second speed or the third speed is set and the accelerator is in an OFF state, that is, a coasting state. First, the accelerator is depressed by the driver to accelerate, and the accelerator opening is switched from OFF to ON. Then, the engine speed and the engine torque start to increase based on the accelerator opening. In other words, the rotational speed and torque (hereinafter referred to as input torque) input to the automatic transmission 3 begin to rise. Then, the input shaft rotational speed increases to the rotational speed obtained by integrating the output shaft rotational speed and the speed ratio after the shift, and at the same time, the input torque decreases and the first clutch C1 starts to be engaged. This time is defined as t0. When the first clutch C1 is engaged at time t0, the one-way clutch F1 is engaged. On the other hand, when the one-way clutch F1 is engaged, torque fluctuation may occur. However, since the engagement pressure of the first clutch C1 is low, that is, the torque capacity TC1 of the first clutch C1 is low, the one-way clutch F1 is engaged. Torque fluctuations caused by this are absorbed by the first clutch C1, in other words, the torque is released by the first clutch C1, so that it is possible to suppress or prevent a change in the driving force of the output shaft. That is, the shift shock can be suppressed or prevented.

  Then, based on the function of the turbine torque TT calculated in step S15 and the torque capacity TC1 of the first clutch C1, the turbine torque TT and the torque capacity TC1 of the first clutch C1 start to increase proportionally. Specifically, the turbine torque TT starts to increase so that the gradient a and the torque capacity of the first clutch C1 become TC1 gradient b. Further, since the turbine torque TT and the torque capacity TC1 of the first clutch C1 are changed according to a function on the assumption that the angular acceleration dωo of the output shaft changes proportionally, eventually the angle of the output shaft The acceleration dωo increases proportionally. Further, the turbine blow amount ωtslp changes based on a quadratic function because the turbine torque TT and the torque capacity TC1 of the first clutch C1 are changed proportionally. That is, the locus of the turbine blow amount ωtslp becomes a parabola.

  Then, the control is switched to the control for suppressing the shift shock caused by the engagement of the first clutch C1 at a time point (hereinafter referred to as t1 time point) before the predetermined first time TC1SLP1 and a predetermined time T1. . Specifically, the turbine torque TT is increased or decreased so that the gradient a2 and the torque capacity TC1 of the first clutch C1 become the gradient b2. In FIG. 1, the torque capacity TC1 of the first clutch C1 is controlled to decrease. However, the gradient of the torque capacity TC1 of the first clutch C1 is calculated by the equation (27) and decreases. It is not always the direction to go. Further, although the turbine torque TT rapidly increases at the time point t1, the increase in the turbine torque TT is due to the switching of the control, and is the torque calculated by the equation (25). Further, at the time t1, the function of the turbine blow amount ωtslp is changed, and the direction from the time point t0 to the time point t1 is inverted up and down. Note that the coefficient in the function of the turbine blow amount ωtslp after the time point t1 is calculated by the equation (15), and is different from the function from the time point t0 to the time point t1.

  When the first predetermined time TC1SLP1 is reached, the angular acceleration dωo of the output shaft coincides with the target angular acceleration DWOTGT, that is, the target acceleration of the vehicle Ve (time t2). On the other hand, at time t2, the input shaft rotational speed does not match the rotational speed obtained by integrating the output shaft rotational speed with the speed ratio after the shift. Accordingly, the turbine torque TT is changed to the gradient a2 and the torque capacity TC1 of the first clutch C1 is changed to the gradient b2 even after the time t2. When the second predetermined time TC1SLP2 is reached, the input shaft rotational speed and the rotational speed obtained by adding the speed ratio after the shift to the output shaft rotational speed coincide with each other (at time t3). Next, the first clutch C1 is completely engaged. Therefore, in order to suppress or prevent slipping of the first clutch C1, the torque capacity TC1 of the first clutch C1 is increased at time t3. The torque capacity TC1 of the first clutch C1 rises to about the torque capacity based on the line pressure, for example.

  By controlling the turbine torque TT and the torque capacity TC1 of the first clutch C1 in this way, it is possible to suppress or prevent a shift shock due to the engagement of the one-way clutch F1 at time t0. Further, since the turbine torque TT and the torque capacity TC1 of the first clutch C1 are controlled based on the acceleration required for the vehicle Ve, the acceleration performance can be maintained. Furthermore, the input clutch rotational speed can be brought close to the rotational speed obtained by integrating the output shaft rotational speed and the speed ratio after the shift, and the first clutch C1 can be engaged, so that the first clutch C1 is engaged. Shift shock can be suppressed or prevented.

  By performing the control as described above, it is possible to suppress a shift shock caused by the engagement of the one-way clutch F1 and the first clutch C1. On the other hand, there is a possibility that the controllability is lowered due to variations due to characteristics of the hydraulic clutch and engine torque, deterioration with time, and the like. Therefore, the present invention is configured to execute feedback control. In other words, the feedback control is executed based on the deviation between the target turbine blow amount ωtslp (t) and the turbine blow amount ωtslp_real calculated from the actual turbine speed. In general, the feedback control is configured to be executed based on the deviation between the target value and the actual measurement value at a specific time, but there is an inevitable response delay until the turbine speed changes. Since an inevitable response delay occurs in the hydraulic pressure that changes the engine torque or the torque capacity TC1 of the first clutch C1, the present invention executes the feedback control in consideration of the response delay. It is configured. Specifically, the target turbine blowing amount before the time when the turbine rotational speed is actually measured is used. That is, the target turbine blowing amount used here is the target turbine blowing amount at the time obtained by subtracting the response delay from the time when the turbine rotational speed is actually measured, as a comparison target in the feedback control. That is, the target turbine blow amount used in the feedback control is calculated by a function of ωtslp (t-TLAG). In addition, TLAG here has shown the response delay time of the turbine blowing amount.

なお、式（２８）におけるωtslp_realは実タービン吹き量、ｇは制御ゲインである。また、実タービン吹き量ωtslp_realは、タービン回転数と出力軸回転数との差から算出することができる。さらに、第１クラッチＣ１の入力側の部材にセンサを設け、出力側の回転数との差を検出して第１クラッチの滑り量を検出し、その滑り量から実タービン吹き量を算出しても良い。つまり、タービン回転数や出力軸回転数あるいは第１クラッチＣ１の入力側もしくは出力側の部材の回転数を、または回転数差を検出する手段が、この発明における滑り量検出手段に相当する。 This feedback control will be specifically described based on the flowchart shown in FIG. In the flowchart shown in FIG. 8, description and description of each step are omitted for the same parts as those up to step S23 described above. Further, in the flowchart shown in FIG. 8, the feedback control is configured to be executed at the time of control for suppressing the shift shock due to the engagement of the first clutch. The feedback control may be executed simultaneously with the start of engagement. That is, the start time of feedback control is not particularly limited. When the function of the turbine torque TT and the torque capacity TC1 of the first clutch C1 is determined in step S23, a feedback correction term is added to one of the functions of the turbine torque TT and the torque capacity TC1 of the first clutch C1 ( Step S31). That is, the feedback correction term shown in the following equation is added.

In Equation (28), ωtslp_real is the actual turbine blow amount, and g is the control gain. Further, the actual turbine blow amount ωtslp_real can be calculated from the difference between the turbine rotational speed and the output shaft rotational speed. Further, a sensor is provided on the input side member of the first clutch C1, the difference between the rotational speed on the output side is detected to detect the slip amount of the first clutch, and the actual turbine blow amount is calculated from the slip amount. Also good. That is, the means for detecting the turbine rotational speed, the output shaft rotational speed, the rotational speed of the input or output side member of the first clutch C1, or the rotational speed difference corresponds to the slip amount detecting means in this invention.

  That is, the controllability is improved by using a deviation between the actual turbine blow amount ωtslp_real and the target turbine blow amount when the response delay time of the turbine blow amount is subtracted from the elapsed time. Note that the target turbine blowing rate is not limited to the point of time when the response delay time of the turbine blowing amount is subtracted from the elapsed time of the actual turbine blowing amount ωtslp_real, as long as the response delay of the turbine blowing amount can be taken into consideration. On the other hand, the turbine blow amount is a value based on the torque capacity TC1 of the first clutch C1 as shown in Expression (4), and is not affected by the input torque. Therefore, when the input torque is large, the deviation between the actual turbine blow amount ωtslp_real and the target turbine blow amount may increase. That is, there is a possibility that the actual turbine blow amount ωtslp_real is difficult to converge on the target turbine blow amount. Therefore, in the present invention, the control gain g is made variable by setting the control gain g large according to the input torque. In addition to the input torque, for example, the control gain g may be set according to the accelerator opening or torque change.

  Then, the turbine torque TT and the torque capacity TC1 of the first clutch C1 are output in accordance with the functions of the turbine torque TT and the torque capacity TC1 of the first clutch C1 set in step S31 (step S32). It is determined whether or not the slip of the first clutch C1 has converged (step S33). If a negative determination is made in step S33, the process returns to step S31. If a positive determination is made in step S33, this control is temporarily terminated.

  As described above, the controllability of the turbine blow amount ωtslp can be improved by adding the feedback correction term to the function of the turbine torque TT or the torque capacity TC1 of the first clutch C1.

  Changes in the engine torque and the control amount of the torque capacity TC1 of the first clutch C1 and changes in the rotational speed of each member will be described based on the time chart shown in FIG. That is, FIG. 9 shows the response delay of the engine torque and the torque capacity TC1 of the first clutch C1 described above. The graph showing the change in the control amount between the input torque and the torque capacity TC1 of the first clutch C1 in FIG. 9 is the same as the time chart in FIG. As shown in FIG. 9, the input torque and the torque capacity of the first clutch C1 start to increase from time t0. On the other hand, since the engine torque and the torque capacity of the first clutch C1 have an inevitable response delay as described above, when the input shaft rotational speed matches the rotational speed obtained by adding the gear ratio to the output rotational speed, the output shaft The time point at which the angular acceleration starts to increase is delayed by a predetermined time TLAG from the time point t0. Similarly, the turbine blow amount ωtslp also starts to rise after a predetermined time TLAG from the time t0.

  The control switching time point (time point t1) and the time point when the slip of the first clutch C1 converges (time point t3) are also delayed by a predetermined time TLAG for the reason described above. Therefore, in the feedback control in the present invention, for example, at the time t4 in FIG. 9, the target turbine blow amount calculated in the feedback control is the target turbine blow amount at the time t1, specifically, the point A in FIG. By performing feedback control in this way, controllability can be improved without being affected by inevitable response delays such as engine torque and torque capacity TC1 of the first clutch C1.

  In the above-described example, in order to improve the controllability of the turbine blow amount ωtslp, the feedback control is performed in consideration of the response delay such as the engine torque and the torque capacity TC1 of the first clutch C1. For example, in order to make the acceleration of the vehicle Ve reach the target acceleration DWOTGT within the first predetermined time TC1SLP1, the target turbine blowing amount in the feedback control is set to the target turbine blowing amount at time t4 in FIG. 9, specifically, FIG. Or a correction term based on the deviation between the actual turbine blow amount and the target turbine blow amount at point B may be added to the correction term in the feedback control.

  The vehicle that can be the subject of the present invention only needs to have a configuration in which power is transmitted by the engagement of the clutch and the one-way clutch. As described above, the transmission includes the clutch and the one-way clutch. It is not limited to what is provided.

  9 ... Ravigneaux type planetary gear mechanism, 10, 11 ... sensor, S20 ... sun gear, R20 ... ring gear, C20 ... carrier, Ve ... vehicle, C1, C2 ... clutch, F1 ... one-way clutch.

Claims (5)

A power transmission device that transmits power by engaging a one-way clutch that transmits only power in one direction in a rotational direction and a clutch that can selectively switch engagement and disengagement; In a vehicle driving force control device configured to slip the clutch when switching from a state where power is not transmitted to a state where power is transmitted,
A first torque for calculating a control amount between a torque input to the power transmission device and a torque capacity of the clutch based on a target angular acceleration of the output shaft of the power transmission device and a target slip amount for slipping the clutch. A driving force control apparatus for a vehicle, comprising a calculating means.   The first torque calculation means includes a torque input to the power transmission device and the clutch so that the clutch is completely engaged when the angular acceleration of the output shaft of the power transmission device reaches the target angular acceleration. The vehicle driving force control device according to claim 1, further comprising means for calculating a control amount with respect to the torque capacity of the vehicle. A second torque calculating means for calculating a control amount between a torque input to the power transmission device and a torque capacity of the clutch so that a change rate of the target slip amount before and after the clutch is completely engaged is reduced; In addition,
Calculated by the second torque calculating means from the control amount of the torque input to the power transmission device calculated by the first torque calculating means and the torque capacity of the clutch before the clutch is completely engaged. The torque input to the power transmission device and the torque capacity of the clutch are switched to be controlled, and the torque input to the power transmission device and the torque capacity of the clutch are output. The driving force control apparatus for a vehicle according to claim 1 or 2.   The second torque calculation means calculates the control amount of the torque input to the power transmission device when the clutch is completely engaged, so that the torque corresponding to the target angular acceleration is calculated. 4. The vehicle driving force control device according to claim 3, further comprising means for calculating a control amount of the torque capacity of the clutch according to the control amount. A slip amount detecting means for detecting the slip amount of the clutch;
The slippage calculated based on the control amount between the torque input to the power transmission device and the torque capacity of the clutch, and the response between the torque input to the power transmission device and the torque capacity of the clutch is delayed. Based on the deviation from the actual slip amount detected by the amount detection means, the control amount of either the torque input to the power transmission device or the torque capacity of the clutch is feedback-controlled. The driving force control apparatus for a vehicle according to any one of claims 1 to 4.

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