JPH04331868A - Slip controller for fluid type transmission device equipped with direct connection clutch for vehicle - Google Patents

Abstract

PURPOSE:To eliminate the speed change shock by speedily lowering the engine revolution speed by setting the difference of the revolution speed between on the input side and output side of a direct connection clutch in the initial stage of the selection to a slip control state, to a prescribed value, and by generating the aimed slip revolution speed which gently lowers. CONSTITUTION:A fluid type transmission device (a) equipped with a direct connection clutch for vehicle is interposed between an engine (b) and an automatic transmission (c). The slip quantity of a direct connection clutch is adjusted by a means (d) so that the aimed value and the actual value of the slip revolution speed accord. In this case, a means (e) judges that the direct connection clutch is switched from a release state to a slip control state. When this selection state is judged, the difference of the revolution speed between on the input side and output side of the direct connection clutch in the initial stage of the selection judgment is set to a prescribed value, and the aimed slip revolution speed which gently reduces afterwards is generated by a means (f). Accordingly, the revolution speed of the engine (b) can be gently reduced, and the generation of speed change shock is prevented.

Description

Description: TECHNICAL FIELD The present invention relates to a slip control device for a hydrodynamic transmission device with a direct coupling clutch for a vehicle. [0002] In a vehicle equipped with a fluid transmission device having a direct coupling clutch such as a torque converter with a lock-up clutch or a fluid coupling with a lock-up clutch, for example, the cycle of the engine while the vehicle is running at low speed is In order to absorb torque fluctuations, a control device has been proposed that causes a direct coupling clutch to slip. For example, there is a control device for a vehicle automatic transmission described in Japanese Patent Publication No. 13060/1983. According to this type of control device, for example, when it is determined that the running state of the vehicle represented by the manifold negative pressure and the engine rotational speed is in a pre-stored slip region, 1.
0~20r. p. m. The actual slip rotation speed of the direct coupling clutch is controlled so as to match the target slip rotation speed set to a certain degree. [0003] However, in the conventional slip control device for a hydrodynamic transmission device with a direct coupling clutch for a vehicle as described above, when the direct coupling clutch is switched from the disengaged state to the slip control, the The rotational speed difference between the input side rotational speed (usually the engine rotational speed) and the output side rotational speed (usually the input shaft rotational speed of the automatic transmission) of the direct coupling clutch, which had formed a large value, is suddenly reduced. As a result, the sudden drop in engine rotational speed causes the inertia of the engine to be released, causing a shock and causing an inconvenient driving experience. Such inconvenience becomes more pronounced in driving conditions where the amount of depression of the accelerator pedal is greater. The present invention has been made against the background of the above-mentioned circumstances, and its object is to provide a vehicle for use in which the shock associated with the switching of a direct coupling clutch from a disengaged state to slip control is suitably eliminated. An object of the present invention is to provide a slip control device for a hydrodynamic transmission device with a direct coupling clutch. [Means for Solving the Problems] The gist of the present invention to achieve the above object is to provide a hydrodynamic transmission device with a direct coupling clutch for a vehicle, as shown in the diagram corresponding to the claims in FIG. , a slip control device comprising a slip adjustment means for adjusting the slip amount of the direct coupling clutch so that the target slip rotational speed and the actual slip rotational speed match, wherein (a) the direct coupling clutch is in its released state; (b) when the switching determining means determines that the direct coupling clutch has been switched from its disengaged state to its slip control state; The present invention includes target slip rotational speed generating means for generating a target slip rotational speed that takes as an initial value the difference between the input side rotational speed and the output side rotational speed of the direct coupling clutch at the beginning of the switching determination, and then gradually decreases thereafter. [0006] With this arrangement, when the switching determining means determines that the direct coupling clutch has been switched from its released state to the slip control state, the target slip rotational speed generating means controls the switching. A target slip rotational speed is generated which uses the difference between the input side rotational speed and the output side rotational speed of the direct coupling clutch at the beginning of the determination as an initial value and gradually decreases thereafter. [0007] Therefore, as described above, the target slip rotation is set to the difference between the input side rotation speed and the output side rotation speed of the direct coupling clutch at the beginning of the determination of switching to the slip control state as an initial value, and gradually decreases thereafter. The slip adjustment means adjusts the slip amount of the direct coupling clutch so that the actual slip rotational speed matches the engine speed, so the engine rotational speed is gradually reduced, the shock is suitably eliminated, and discomfort during driving is prevented. It will be done. [Embodiment] An embodiment of the present invention will be described in detail below with reference to the drawings. FIG. 2 is a schematic diagram of a vehicle power transmission device to which an embodiment of the present invention is applied. In the figure, power from an engine 10 is transmitted to the drive wheels via a torque converter 12 with a lock-up clutch, a stepped automatic transmission 14 composed of three planetary gear units, and a differential gear device (not shown). It has become so. The torque converter 12 is connected to the engine 1.
a pump impeller 18 connected to the crankshaft 16 of the automatic transmission 14; a turbine impeller 22 fixed to the input shaft 20 of the automatic transmission 14 and rotated by receiving oil from the pump impeller 18; a stator wheel 28 fixed to a housing 26, which is a non-rotating member, via a clutch 24;
A lock-up clutch 32 is connected to the input shaft 20 via a damper 30. When the hydraulic pressure in the disengagement side oil chamber 33 is higher than that in the engagement side oil chamber 35 in the torque converter 12, the lockup clutch 32 is brought into a disengaged state, so that the input/output rotational speed ratio of the torque converter 12 is changed. Torque is transmitted with a corresponding amplification factor. However, when the oil pressure in the engagement side oil chamber 35 is higher than that in the release side oil chamber 33, the lockup clutch 32 is brought into an engaged state. The shaft 20 is directly connected. [0010] The automatic transmission 14 includes three
Single pinion type planetary gear set 34, 36, 38
and a counter shaft (output shaft) 40 that transmits power between the input shaft 20, the output gear 39 that rotates together with the ring gear of the planetary gear set 38, and the differential gear set. Some of the components of the planetary gear systems 34, 36, 38 are not only integrally connected to each other, but also selectively connected to each other by three clutches C0, C1, C2. In addition, the planetary gear device 34, 3
Some of the components of 6,38 are the four brakes B0,
selectively connected to the housing 26 by B1, B2, and B3, and further, some of the components
The one-way clutches F0, F1, F2 can be engaged with each other or with the housing 26 depending on the direction of rotation. [0011] The clutches C0, C1, C2 and brakes B0, B1, B2, B3 are composed of, for example, a multi-plate clutch or a band brake having one band or two bands wound in opposite directions. , are each actuated by a hydraulic actuator, and the operation of these hydraulic actuators is controlled by an electronic control device 42, which will be described later.
As shown in FIG. 3, four forward speeds and one reverse speed are obtained with different speed ratios I (=rotational speed of input shaft 20/rotational speed of counter shaft 40). In FIG. 3, "1st", "2nd", "3rd", "O/
"D (Overdrive)" represents the first gear, second gear, third gear, and fourth gear on the forward side, respectively, and the above gear ratios are from the first gear to the forward gear. The value gradually decreases toward the fourth gear. Note that since the torque converter 12 and automatic transmission 14 are constructed symmetrically with respect to the axis, the input shaft 2 is shown in FIG.
The lower side of the rotation axis of 0 and the upper side of the rotation axis of the counter shaft 40 are omitted from illustration. [0012] A hydraulic control circuit 44 for shift control for controlling the gear stage of the automatic transmission 14 and a hydraulic control circuit 46 for engagement control for controlling engagement of the lock-up clutch 32 are provided. It is being As is well known, the hydraulic control circuit 44 for speed change control uses solenoid No.
．． 1 and solenoid no. 2, the first solenoid valve 48 and the second solenoid valve 50 are respectively turned on and off.
As shown in FIG. 3, the clutch and the brake are selectively operated by the combination of the above operations to establish one of the first to fourth gears. The engagement control hydraulic control circuit 46 also includes a solenoid N, which is a linear solenoid, as shown in FIG.
o. 3, a switching valve 54 that can be switched between a release side position where the lockup clutch 32 is released and an engagement side position where the lockup clutch 32 is engaged; The slip control valve 56 has a slip control valve 56 whose source pressure is a regulator pressure Pcl generated by a clutch pressure regulating valve (not shown) in the hydraulic control circuit 42 in accordance with the throttle valve opening. The linear solenoid valve 52 has a constant modulator pressure Pm generated within the hydraulic control circuit 42 for speed change control.
odu as the source pressure, continuously generates an output pressure Plin of a magnitude corresponding to the magnitude of the drive current Isol from the electronic control device 42, and applies this output pressure to the switching valve 54 and the slip control. act on valve 56. The switching valve 54 includes a spring 58 that biases a spool valve element (not shown) toward a release side position;
a first port 60 to which the regulator pressure Pcl is supplied;
, a second port 62 to which the output pressure of the slip control valve 56 is supplied, and a third port 64 connected to the release side oil chamber 35.
, a fourth port 66 connected to the engagement side oil chamber 33, and a fifth port 68 connected to the drain. When the output pressure Plin of the linear solenoid valve 52 supplied thereto falls below a predetermined constant value, the spool valve element of the switching valve 54 is positioned at the release side position according to the biasing force of the spring 58. The second port 62 is closed, and the first port 60 and the third port 64 and the fourth port 66 and the fifth port 68 are communicated with each other. Therefore, when the output pressure Plin of the linear solenoid valve 52 acting on the spool valve element of the switching valve 54 falls below a predetermined constant value, the spool valve element of the switching valve 54 moves toward the release side according to the biasing force of the spring 58. position, the oil pressure Poff in the disengagement side oil chamber 35 becomes the regulator pressure Pcl, and at the same time the engagement side oil chamber 33
The hydraulic pressure Pon inside is set to atmospheric pressure, and the lock-up clutch 32 is released. However, the output pressure Plin of the linear solenoid valve 52 applied to the spool valve element of the switching valve 54
exceeds a predetermined value, the spool valve element of the switching valve 54 is switched to the engagement side position against the biasing force of the spring 58, closing the fifth port 68 and closing the first port. 60 and the fourth port 66, and between the second port 62 and the third port 64, respectively. Therefore, at the same time as the oil pressure Pon in the engagement side oil chamber 33 is set to the regulator pressure Pcl, the oil pressure Poff in the release side oil chamber 35 is pressure controlled by the slip control valve 56, and the lockup clutch 32 is controlled to slip. Or be released. The slip control valve 56 includes a spring 70 for biasing a spool valve element (not shown) toward an output pressure increase side. Hydraulic pressure Pon in the engagement side oil chamber 33 is applied to this spool valve in order to generate a thrust toward the output pressure increasing side, and it is released to generate a thrust toward the output pressure decreasing side. The oil pressure Poff in the side oil chamber 35 and the output pressure Plin of the linear solenoid valve 52 are respectively applied. For this reason,
The slip control valve 56 operates so that the differential pressure ΔP (=Pon-Poff) corresponding to the slip amount becomes a value corresponding to the output pressure Plin of the linear solenoid valve 52, as shown in Equation 1. Here, in Equation 1, F is the biasing force of the spring 70,
A1 is the pressure receiving area of the hydraulic pressure Pon in the spool valve,
A2 (where A1 = A2) is the pressure receiving area of the oil pressure Poff, and A3 is the pressure receiving area of the output pressure Plin. ##EQU1## Therefore, in the engagement control hydraulic control circuit 46 configured as described above, the hydraulic pressure Pon in the engagement side oil chamber 33 and the hydraulic pressure Poff in the disengagement side oil chamber 35 are As shown in FIG. 4, is changed according to the output pressure Plin of the linear solenoid valve 52, so that the output pressure Plin of the linear solenoid valve 52 controls the switching of the switching valve 54, and the switching valve 54 is in the engaged position. Slip control of the lock-up clutch 32 after switching to the lock-up clutch 32 can be performed respectively. The electronic control unit 42 includes a CPU 82, a ROM
84, a RAM 86, an interface (not shown), etc., which is a so-called microcomputer.
A throttle sensor 88 that detects the throttle valve opening provided in the intake pipe of the engine 10, an engine rotation speed sensor 90 that detects the rotation speed of the engine 10, and an input shaft that detects the rotation speed of the input shaft 20 of the automatic transmission 14. A rotation sensor 92, a counter shaft rotation sensor 94 that detects the rotation speed of the counter shaft 40 of the automatic transmission 14, detects the operation position of the shift operation lever 96, that is, one of the L, S, D, N, R, and P ranges. A signal representing the throttle valve opening θth, a signal representing the engine rotational speed Ne (pump impeller rotational speed NP, that is, the input side rotational speed Np of the lock-up clutch 32), and an input shaft rotational speed are received from the operation position sensor 98 for Nin (Turbine wheel rotation speed N
A signal representing the output shaft rotational speed Nout, and a signal representing the operating position Ps of the shift operation lever 96 are supplied. The CPU 82 of the electronic control unit 42 processes input signals according to a program stored in advance in the ROM 84 while utilizing the temporary storage function of the RAM 86, and executes shift control of the automatic transmission 14 and engagement control of the lock-up clutch 32. In order to do this, the first solenoid valve 48, the second solenoid valve 50, and the linear solenoid valve 52 are respectively controlled. In the above-mentioned shift control, a shift diagram corresponding to an actual gear is selected from a plurality of types of shift diagrams stored in advance in the ROM 84, and the running state of the vehicle, such as throttle valve opening, is determined from the shift diagram. The transmission gear stage is determined based on the vehicle speed calculated from the rotational speed θth and the output shaft rotational speed Nout, and the first solenoid valve 48 and the second solenoid valve 50 are driven to obtain the determined transmission gear stage. , the operations of the clutches C0, C1, C2 and the brakes B0, B1, B2, B3 of the automatic transmission 14 are controlled to establish one of the four forward gears. Note that FIG. 3 shows the gears in each shift range of the shift operation lever 96 and the operating states of the solenoid, clutch, brake, and one-way clutch when establishing the gears. "
The ``○'' and ``x'' marks represent an energized state and a de-energized state, respectively. Moreover, the "○" mark in the column of clutches and brakes represents an engaged state, and the blank mark represents a non-engaged state. Furthermore, the "○" mark in the one-way clutch column indicates that the clutch is in an engaged state during forward drive, and the blank mark indicates a non-engaged state. Hereinafter, the engagement control operation of the lock-up clutch 32 will be explained in detail using the flowcharts of FIGS. 5 and 6. In the figure, in step S1, the various input signals, which are the state quantities of the running vehicle, are read, and in step S2, whether or not the engagement condition of the lock-up clutch 32 is satisfied is stored in the ROM 84 in advance.
The determination is made based on the relationship stored in the vehicle running state, for example, the output shaft rotational speed (vehicle speed) Nout and the throttle valve opening θth. This relationship is selected from a plurality of pre-stored relationships depending on the actual gear position, and for example, the relationship shown in FIG. 7 is used. In Fig. 7, the coupling effect is maintained on the lower throttle valve opening side, which is closer to the release area than the boundary line between the engagement area and the release area, in order to improve fuel efficiency as much as possible without impairing drivability. At the same time, a slip control region is provided to absorb torque fluctuations of the engine 10. If the judgment in step S2 is affirmative because the actual output shaft rotational speed Nout and throttle valve opening θth are within the engagement range shown in FIG. "0"
cleared. This flag XLC has the content “1
”, this indicates that the release area in FIG. 7 has been passed once. Next, in step S4, the drive current Isol for the linear solenoid valve 52 is set to 100%.
That is, after the maximum drive current (rated value) is set, the drive current Isol is output in step S5. As a result, as shown in FIG. 4, the oil pressure Pon in the engagement side oil chamber 33 reaches its maximum value, the regulator pressure Pcl.
At the same time, the oil pressure Poff in the release side oil chamber 35 is set to its minimum value, that is, atmospheric pressure, so the lockup clutch 32 is brought into an engaged state. However, the actual output shaft rotational speed Nout
If the throttle valve opening θth is not within the engagement range shown in FIG. 7, the determination in step S3 is negative, and therefore, in step S6, it is determined whether the slip control condition is satisfied. This slip control condition is that the actual output shaft rotational speed Nout and throttle valve opening θth are within the slip control region shown in FIG. If the determination in step S6 is negative, the actual output shaft rotational speed Nout and throttle valve opening θth are within the release range shown in FIG.
In step S7, the contents of the flag XLC are set to "1", and in step S8, the actual slip rotational speed NslipT is calculated from Equation 2 based on the engine rotational speed Ne and the turbine rotational speed NT. ##EQU00002## Then, in the following step S9, the drive current Isol for the linear solenoid valve 52 is set to 0%,
That is, after the minimum drive current is set, the drive current Isol is output in step S5. As a result, as shown in FIG. 4, the oil pressure Pof in the release side oil chamber 35
At the same time, the hydraulic pressure Pon in the engagement side oil chamber 33 is set to the minimum value, ie, the regulator pressure Pcl, which is the maximum value.
In other words, since it is atmospheric pressure, the lock-up clutch 3
2 is in the released state. If the determination in step S6 is affirmative, the vehicle condition is within the slip control region shown in FIG. 7, so slip control of the lock-up clutch 32 is executed in steps S10 to S21. . First, in step S10, for example, FIG.
A steady-state target slip rotational speed NslipT c is determined based on the throttle valve opening θth and the turbine rotational speed NT from the relationship shown in FIG. The relationship shown in FIG. 8 has been determined in advance in order to maintain the connected state within the slip control region during steady driving, while suitably absorbing torque fluctuations of the engine 10 and improving fuel efficiency as much as possible. be. Then, when the steady state target slip rotational speed NslipT c is determined, step S11
It is determined whether the content of the flag XLC is "1" or not. If step S11 is executed without going through the release area in FIG. 7, the content of the flag The steady-state target slip rotational speed NslipT c determined in S10 is temporarily stored in register A;
In step S18, the target slip rotation speed NslipT in the steady state, which is the content of the register A, is set as the target slip rotation speed NslipT. Then, in step S19, the actual slip rotational speed Nslip is calculated from Equation 3 based on the engine rotational speed Ne and the turbine wheel rotational speed NT, and in step S20, the control deviation ΔN is calculated using Equation 3.
is calculated based on the target slip rotational speed NslipT and the actual slip rotational speed Nslip. [0027] [0028] [0029] Then, in step S21, the drive current Isol for the linear solenoid valve 52 is calculated from the control equation shown in equation 5, and in step S5, the drive current Isol is Output. This control formula is for executing PID feedback control to eliminate the control deviation ΔN. ##EQU00005## However, if step S11 is executed after passing through the release area in FIG. 7, the content of the flag XLC is "1" and the determination in step S11 is affirmed. Therefore, in step S12, the target slip rotational speed NslipT during transition is determined based on the actual throttle valve opening θth from the pre-stored relationship shown in FIG.
A decrease value ΔNs is calculated. This relationship is such that the target slip rotational speed NslipT, whose initial value is the rotational speed difference between the engine rotational speed Ne and the turbine rotational speed NT immediately before the start of slip control of the lockup clutch 32, is sufficient for supercharging in a range that does not cause problems due to shock. This has been experimentally determined in advance so as to quickly return to the target slip rotational speed NslipT c in the steady state. Then, in the subsequent step S13, the target slip rotational speed NslipT obtained in the step S8 is
is set in register A, and step S14
Then, the decrease value ΔNs determined in step S12 is subtracted from the contents of register A. In the following step S15, the register A
The content of is the steady state target slip rotational speed NslipT
It is determined whether or not the value is less than or equal to c. Within a predetermined period after this step S15 is first executed, the contents of register A are still at the target slip rotational speed Ns in the steady state.
Since it is larger than lipT c , the determination in step S15 is denied, and steps from step S18 onwards are executed. Therefore, by repeatedly performing the above steps, the engine rotational speed Ne and the turbine rotational speed N at the beginning of the slip control of the lockup clutch 32 are
The target slip rotational speed NslipT, whose initial value is the rotational speed difference with T, is gradually decreased as shown in FIG. Then, when the contents of the register A, that is, the target slip rotational speed NslipT reaches the target slip rotational speed NslipT c in the steady state, the judgment in step S15 is affirmed, so in step S16 the contents of the flag XLC become "0". ” after being reset to
Step S17 and subsequent steps are executed. Therefore, from now on, the steady state target slip rotational speed NslipT
Target slip rotational speed NslipT with content c
Slip control of the lockup clutch 32 is performed so that the actual slip rotational speed ΔN matches the actual slip rotational speed ΔN. As described above, in this embodiment, when it is determined in step S11 corresponding to the switching determining means that the lock-up clutch 32 has been switched from its released state to the slip control state, the target slip Steps S12 to S18 corresponding to the rotational speed generating means determine the engine rotational speed Ne (=input side rotational speed Np of the lockup clutch 32) and the turbine rotational speed NT (=output side of the lockup clutch 32) at the beginning of the switching determination. A target slip rotational speed NslipT is generated which has the rotational speed difference from the rotational speed (rotational speed) as an initial value and gradually decreases thereafter. Then, the slip control of the lock-up clutch 32 is executed so that the target slip rotational speed NslipT and the actual slip rotational speed ΔN match, so that the engine rotational speed Ne is gradually reduced as shown in FIG. This effectively eliminates the shock caused by the shift to slip control, and prevents the driver from feeling uncomfortable while driving. Incidentally, in the conventional control device, as shown in FIG. 11, when the lock-up clutch 32 shifts from the disengaged state to the slip control, the slip control is started, and the engine Rotational speed N
Since the rotational speed difference between e and the turbine rotational speed NT is rapidly reduced, the engine rotational speed Ne is similarly reduced, causing a shock. Furthermore, according to this embodiment, in connection with the switching from the disengaged state of the lock-up clutch 32 to the slip control, the engine rotational speed N at the time of the switching is changed.
The torque converter 12, the hydraulic control system, the engine 1
Regardless of individual differences in characteristics such as 0 or changes over time, the same switching feeling can be obtained. Although one embodiment of the present invention has been described above based on the drawings, the present invention can also be applied to other embodiments. For example, in the above-described embodiment, when the lock-up clutch 32 is switched from the disengaged state to the slip control, the target slip rotation speed NslipT is the engine rotation speed Ne and the turbine rotation speed N at the time of the switching.
The initial value was the rotational speed difference with T (=Ne - NT), but the initial value is k (Ne - NT), which is multiplied by a coefficient k such as 0.8, 0.9, etc. You can. Furthermore, although the reduction value ΔNs in the above embodiment was a function of the throttle valve opening θth, it may be a constant value. Further, in the above-described embodiment, the stepped planetary gear type automatic transmission 14 was provided downstream of the torque converter 12, but a continuously variable transmission may also be used. Furthermore, in the above embodiment, the torque converter 12 with a direct coupling clutch was described;
A fluid coupling with a direct clutch may also be used. In short, any hydrodynamic transmission device that has a direct coupling clutch will suffice. Furthermore, in the engagement hydraulic control circuit 46 of the embodiment described above, the differential pressure ΔP (=Pon−
Poff ) is the output pressure Pli of the linear solenoid valve 52
A slip control valve 56 was used that operates so that the differential pressure ΔP becomes a value corresponding to the output pressure Plin, but a normal flow control valve that does not operate so that the differential pressure ΔP becomes a value corresponding to the output pressure Plin may be used. The above-mentioned embodiment is merely one embodiment of the present invention, and various modifications may be made to the present invention without departing from the spirit thereof.

[Brief explanation of the drawing]

FIG. 1 is a claim correspondence diagram of the present invention.

FIG. 2 is a schematic diagram of a control device according to an embodiment of the present invention and a vehicle power transmission device to which the control device is applied.

3 is a chart illustrating the relationship between the combination of operations of the first solenoid valve and the second solenoid valve and the resulting gear stage in the automatic transmission included in the hydrodynamic transmission device of FIG. 2; FIG.

4 is a diagram illustrating the relationship between the output pressure Plin of the linear solenoid valve and the oil pressure Pon in the engagement side oil chamber and the oil pressure Poff in the release side oil chamber of the lock-up clutch in the engagement control hydraulic control circuit of FIG. 2; be.

5 is a part of a flowchart illustrating an engagement control operation of a lock-up clutch provided in the torque converter of FIG. 2; FIG.

6 is a part of a flowchart illustrating an engagement control operation of a lock-up clutch provided in the torque converter of FIG. 2; FIG.

FIG. 7 is a diagram showing relationships used in the flowcharts of FIGS. 5 and 6;

FIG. 8 is a diagram showing relationships used in the flowcharts of FIGS. 5 and 6;

FIG. 9 is a diagram showing relationships used in the flowcharts of FIGS. 5 and 6;

FIG. 10 is a time chart illustrating operations obtained by the flowcharts of FIGS. 5 and 6;

FIG. 11 is a time chart illustrating operations in conventional engagement control.

[Explanation of symbols]

12 Torque converter (hydraulic transmission device) 32
Lock-up clutch (direct clutch) 46 Hydraulic control circuit for engagement control (slip adjustment means) Step S11
Switching determination means steps S12 to S18 Target slip rotational speed generation means

Claims (1)

[Claims] 1. A slip control device in a hydrodynamic transmission device with a direct coupling clutch for a vehicle, comprising a slip adjustment means for adjusting the amount of slip of the direct coupling clutch so that a target slip rotational speed and an actual slip rotational speed match. a switching determination means for determining that the direct coupling clutch has been switched from its disengaged state to a slip control state; If the determination is made, target slip rotation speed generating means generates a target slip rotation speed that takes as an initial value the difference between the input side rotation speed and the output side rotation speed of the direct coupling clutch at the beginning of the switching determination, and that gradually decreases thereafter. A slip control device for a hydrodynamic transmission device with a direct coupling clutch for a vehicle, comprising: