JP2778033B2 - Fluid coupling for vehicles - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a vehicle fluid coupling interposed between an engine and a transmission.

2. Description of the Related Art Conventionally, for example, the following technology is disclosed as a technology for improving the driving performance at the time of starting and the driving feeling at the time of deceleration of a vehicle equipped with a continuously variable transmission. For example, when starting, the speed ratio of the continuously variable transmission is increased to improve start acceleration, while at the time of deceleration, the speed ratio is reduced,
There is disclosed a technique for preventing a driver from feeling excessive deceleration (see Japanese Patent Application Laid-Open No. 59-176563).

[Problems to be Solved by the Invention] However, for example, Japanese Patent Application Laid-Open
By using the technology disclosed in Japanese Patent No. -175663, driving is performed with a reduced feeling of good acceleration at the time of starting acceleration and an excessive feeling of deceleration (engine braking effect) when the engine is driven during deceleration. Although it was possible to obtain the effect of improving the feeling in a predetermined region, the driving feeling at the time of deceleration was sometimes deteriorated due to the following problems.

That is, it is necessary to increase the gear ratio before the vehicle stops in preparation for the next start in order to improve the driving feeling during deceleration. There is a problem that the feeling of deceleration increases and the driver feels excessive deceleration and jerky feeling.

An object of the present invention is to improve the driving feeling when the vehicle is decelerated by solving the above problems.

[Means for Solving the Problems] A fluid coupling for a vehicle according to the present invention, which has been made to achieve the above object, is provided between a vehicle engine and a transmission as illustrated in FIG. In the fluid coupling for a vehicle, when the engine is in a predetermined driven state, the flow direction of the hydraulic oil flowing from the pump to the turbine is in a forward direction with respect to the rotation of the turbine runner, and the engine is driven at a predetermined driving state. By adopting the shapes of the pump impeller and the turbine runner, the flow direction of the hydraulic oil flowing from the turbine to the pump when in the state is opposite to the rotation of the pump impeller. When the engine is in the driving state, the capacity coefficient is made smaller than when the engine is in the driving state.

The driven state of the engine is, for example, a state in which the engine is driven by a power transmission system of the vehicle.

The capacity coefficient of the fluid coupling is one of the requirements for determining the transmission torque of the fluid coupling.For example, reducing the internal pressure of the fluid coupling, reducing the amount of oil, and obstructing the outlet of the turbine runner. It is known to reduce the capacity coefficient by providing a plate or the like.

[Operation and Effect of the Invention] In the fluid coupling for a vehicle according to the present invention, the flow direction of the hydraulic oil flowing from the pump to the turbine when the engine is in a predetermined driven state is forward with respect to the rotation of the turbine runner. The pump impeller and the turbine runner have a shape in which the flow direction of hydraulic oil flowing from the turbine to the pump when the engine is in a predetermined driving state is opposite to the rotation of the pump impeller. As an example of such a shape of the pump impeller and the turbine runner, as shown in FIG. 3, the pump impeller has a shape in which the rotation destination side is concave, and the turbine runner has a shape in which the rotation destination side is convex.

In order to make it easier to understand the above operation and effects, an example in which the present invention is applied to a fluid clutch of a vehicle with a continuously variable transmission will be described with reference to FIGS. Will be described.

(1) When the engine is running, oil flows out from the outer peripheral side of the pump impeller and flows in from the outer peripheral side of the turbine runner toward the center, as shown by the solid line in FIG.

When the engine is driven, oil flows out from the outer peripheral side of the turbine runner and flows from the outer peripheral side of the pump impeller toward the center, as shown by the two-dot chain line in FIG. From the turbine runner. In this case, the direction of the oil flow from the pump impeller to the turbine runner is in the direction indicated by VP in FIG. 3 (forward direction with respect to the rotation of the turbine runner), and it becomes the drive side. The capacity coefficient is reduced because it helps the turbine runner to rotate.

On the other hand, if the engine is running, the flow direction of the oil flowing from the center of the turbine runner to the pump impeller is opposite to the above-mentioned VP (“outflow direction in case of driving” in FIG. 3). Since the direction is opposite to the rotation of the impeller, the pump driving torque increases, that is, the capacity coefficient increases.

(2) When the driving state changes from the driving state of (1) to the driven state, the fourth indicating the relationship between the engine torque and the capacity coefficient.
As shown in the figure, for example, the driving torque (T) of the engine at the point A shifts to the driven torque (-T) at the point B. That is, when the power is turned off in the state of the point A, the turbine rotation speed (the rotation speed of the turbine runner) Nt is calculated from the engine rotation speed Ne.
(The number of revolutions corresponding to the vehicle speed V of the continuously variable transmission), and if the capacity coefficient does not change between when the engine is in the driven state and when the engine is in the driven state, the turbine speed and the capacity coefficient As shown in FIG. 5, the speed is balanced by the speed ratio (here, e'1) with the pump impeller that matches the driven torque (-T) of the engine. That is, the turbine runner rotates at Nt1, and the engine rotates at Ne1 at point B in FIG.

(3) However, if the capacity coefficient becomes smaller when the engine is in the driven state than in the driven state, for example, the capacity shown by the solid line in FIGS. 4 and 5 is changed to the capacity shown by the two-dot chain line. As shown in FIG. 5, since the vehicle speed is constant, that is, the turbine speed Nt1 is constant, the speed ratio e '= Np / Nt is small (e ").
4), the engine speed drops to Ne2 at point B 'in FIG.

(4) As shown in the above (3), by decreasing the capacity coefficient, the engine speed Ne decreases and the engine reverse drive torque changes with the engine speed. The torque decreases. As shown in FIG. 4, when the capacity coefficient is reduced until the engine speed reaches the idle speed, the reverse drive torque, which is the driven torque of the engine, is not substantially generated.

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

In FIG. 6, a vehicle engine 10 includes a direct-coupled clutch.
It is connected to an input shaft 16 of a continuously variable transmission 14 via a fluid coupling 12 as a fluid clutch with 11. Note that the turbine runner 302 and the pump impeller 304 of the fluid coupling 12 have a blade shape as shown in FIG. 3, the pump impeller 304 has a concave shape on the rotation destination side, and the turbine runner 302 has a convex shape on the rotation destination side. It has a shape. A variable pulley whose input shaft 16 has a V-groove width, that is, a hanging diameter of the conduction belt 20, is changed by a hydraulic cylinder 18.
22 are provided. The output shaft 24 is provided with a variable pulley 28 whose V groove width is changed by a hydraulic cylinder 26. Accordingly, the rotational force transmitted to the input shaft 16 is transmitted to the output shaft 24 via the transmission belt 20 wound around the variable pulleys 22 and 28, and is also transmitted to the sub-transmission 30 at the subsequent stage. The auxiliary transmission 30 includes a Ravigneaux-type compound planetary gear device including a first sun gear 32, a second sun gear 34, a ring gear 36, and the like. A high speed clutch 38, a low speed brake 40, and a reverse brake 42 are not shown. The following Table 1 shows that the actuator is selectively operated by the hydraulic actuator.
As shown in the figure, the speed ratio Rf of the subtransmission 30 is switched,
Alternatively, normal rotation and reverse rotation can be switched.

Here, in Table 1, ρ1 is Zs1 / Zr, and ρ2 is Zs2 / Zr. Here, Zs1 is the number of teeth of the first sun gear 32, Zs2 is the number of teeth of the second sun gear 34, and Zr is the number of teeth of the ring gear 36. The output shaft 24 of the belt-type continuously variable transmission 14 constitutes an input shaft of the sub-transmission 30 and a carrier 44 that supports a planetary gear in the sub-transmission 30.
Constitutes the output shaft, the speed ratio of the auxiliary transmission 30 is a value obtained by dividing the rotation speed of the output shaft 24 by the rotation speed of the carrier 44. The rotational force transmitted to the carrier 44 is transmitted to a pair of drive wheels 52 of the vehicle via the intermediate gears 46 and 48 and the final reduction gear 51.

In the vicinity of the variable pulleys 22 and 28, a pulse signal SP1 having a frequency corresponding to the rotation speed of the variable pulleys and 28 is provided.
And an output shaft rotation speed sensor 60 for outputting SP2 and SP2 to the controller 54. In the vicinity of the intermediate gear 48, a vehicle speed sensor 61 for outputting a pulse signal SV having a frequency corresponding to the rotation speed of the intermediate gear 48 to the controller 54 is provided. A throttle valve 62 provided in an intake pipe of the engine 10 is opened and closed by operating an accelerator pedal.The throttle valve 62 is provided with a throttle sensor 64, and the throttle sensor 64 detects a throttle valve opening θ from the throttle sensor 64. Express throttle signal S
θ is supplied to the controller 54. The ignition circuit of the engine 10 is provided with an engine speed sensor 65, from which a speed signal SNE representing the engine speed Ne is supplied to the controller 54.

In this embodiment, a shift lever 66 is used as a shift switching device. From an operation position sensor 68 that detects the operation position of the shift lever 66, a signal SP indicating the shift operation position Psh of the shift lever 66 is sent to the controller 54. Supplied to This shift lever 66 is mechanically associated with a manual valve in the hydraulic circuit 70, and when operated in the neutral range, the high speed clutch 38,
This prevents the hydraulic pressure from being supplied to any of the hydraulic actuators for operating the low-speed gear brake 40 and the reverse brake 42, respectively.However, when operated in the reverse range, the hydraulic actuator that operates the reverse brake 42 is operated. Only supply hydraulic oil. Also, shift lever 66
Is operated in the normal traveling (D: drive) range of the forward range, the hydraulic oil is supplied only to the hydraulic actuator that operates the high speed clutch 38, and the high speed gear To be maintained. When the shift lever 66 is set to the automatic shift range (S
Range) or the engine brake range (L range), the operation oil is supplied to one of the hydraulic actuators that operate the high speed clutch 38 and the low speed brake 40. . To these hydraulic actuators, hydraulic pressure is alternatively supplied from a shift valve that operates in response to the operation of a shift electromagnetic valve 72 provided in the hydraulic circuit 70.

The hydraulic circuit 70 supplies the hydraulic cylinder 26 provided on the output shaft 24 with a line hydraulic pressure adjusted in accordance with the actual speed ratio of the continuously variable transmission 14 and the output torque of the engine 10,
The tension of the conduction belt 20 is necessary and sufficiently controlled. Also,
The hydraulic circuit 70 supplies or discharges hydraulic oil to the hydraulic cylinder 18 provided on the input shaft 16 in response to the operation of the shift direction switching valve 74, and also includes a shift speed switching valve.
In response to the operation of 76, the hydraulic oil inflow speed to the hydraulic cylinder 18 or the hydraulic oil discharge speed from the hydraulic cylinder 18 is changed.

Also, the hydraulic circuit 70 includes a lock-up solenoid valve 77 and a quick release solenoid valve 78, which will be described later, and switches the direction of hydraulic oil to the direct coupling clutch 11,
The internal pressure of the fluid coupling 12 is reduced to lower the capacity coefficient.

The hydraulic pump 79 is driven by the engine 10 or the like, so that the hydraulic oil in the oil tank 80 is supplied to the hydraulic circuit.
The hydraulic pressure is supplied to the hydraulic circuit 70 and functions as a hydraulic pressure source of the hydraulic circuit 70.

The controller 54 includes an input / output interface 82, a central processing unit 84, a storage unit 86, etc., and various input signals input via the input / output interface 82 according to programs and data stored in the storage unit 86 in advance. By controlling the operation of the shift solenoid valve 72 on the basis of the processing result, thereby automatically shifting the gear stage of the subtransmission 30. The shift direction switching valve 74 and the shift speed switching valve
By controlling the operation of the lock-up solenoid valve 77, the gear ratio of the continuously variable transmission 14 is changed to an optimum value.
, The direct coupling clutch 11 is locked up “on” or locked up “off”, and by controlling the release solenoid valve 78, the internal pressure of the fluid coupling 12 is reduced.

Next, the gear ratio control routine of this embodiment, which is executed every predetermined time (here, 8 msec), will be described with reference to the flowchart of FIG.

FIG. 7 shows a control routine for controlling the speed ratio of the entire transmission of the vehicle. First, the vehicle speed V, the throttle opening θ, the rotation speed Nin of the input shaft 16 and the rotation speed Nout of the output shaft 24 are shown. , Engine speed Ne, shift lever
The 66 operation positions Psh are read based on the signals SV, Sθ, SP1, SP2, SNE, and SP (step 100). Next, shift lever 66
It is determined whether the actual operation position is in the normal travel range or the automatic shift range (step 110). If it is determined that the vehicle is in the normal travel range, the speed ratio control routine in the normal travel range shown in FIG. 8 stored in advance in the storage unit 86 is executed to optimize the speed ratio γ of the continuously variable transmission 14. Control (step 120).

On the other hand, if it is determined that the shift lever 66 has been controlled to the automatic shift range (step 110), the sub-transmission 3
The shift control of 0 is executed (step 130). That is,
From the shift pattern stored in the storage unit 86 in advance, the vehicle speed V
The gear stage of the auxiliary transmission 30 is determined based on the throttle opening θ, and a drive signal is output to the shift electromagnetic valve 72 so that the determined gear stage is realized. The shift pattern is, for example, as shown in FIG. 9, and is stored in a form such as a data map. In the figure, U12 is prepared in consideration of the running performance of the vehicle, and is used to determine an upshift from a lower gear (first speed) to a higher gear (second speed). It is a shift line, and D21 in the figure is
It is prepared to form an appropriate hysteresis and considering the acceleration performance by kick down,
It is a downshift line used to determine a downshift from a higher gear to a lower gear.

Next, it is determined whether the actual gear of the subtransmission 30 is the high gear or the low gear (step 140). If it is determined that the gear is the higher gear,
For example, instead of the gear ratio control routine (step 120) in the normal traveling range shown in FIG. 8, a gear ratio control routine in a high gear stage (not shown in detail) is started to execute the gear ratio control of the continuously variable transmission 14. (Step 150).

If it is determined in step 140 that the gear stage of the auxiliary transmission 30 is the lower gear stage, for example, instead of the gear ratio control routine (step 120) in the normal traveling range shown in detail in FIG. A gear ratio control routine in a low gear stage (not shown in detail) is started, and the continuously variable transmission is started.
The gear ratio control of 14 is executed (step 160).

Next, a gear ratio control routine in the normal traveling range shown in FIG. 8 will be described. In the control routine of FIG. 8, first, the calculation of the target rotation speed Nin ^* is stored in advance in the storage unit 86, and then the target rotation speed Nin ^* data map for the normal travel range of FIG. This is performed based on the vehicle speed Vi (i = 0 to max) and the throttle opening θ read in step 100 in the figure (step 170).

After calculating the target rotation speed Nin ^* , the continuously variable transmission 1
Perform control to change the gear ratio of step 4 (steps 180 through
200). That is, first, it is determined whether or not the target rotation speed Nin ^* is equal to or less than the rotation speed Nin of the input shaft 16 (step 180).
Next, if Nin ^* > Nin, it is determined that the rotational speed Nin of the input shaft 16 is to be increased, and the continuously variable transmission is controlled by controlling the shift direction switching valve 74 and the shift speed switching valve 76.
The control (downshift control) for increasing the speed ratio γ of 14 is executed (step 190). On the other hand, when the target rotation speed Nin ^* is smaller than the rotation speed Nin of the input shaft 16, the continuously variable transmission 14
(Upshift control) for reducing the gear ratio γ of the vehicle (step 200).

By executing the steps 170 to 200, the speed ratio of the continuously variable transmission 14 when the shift lever 16 is in the normal running range is based on the target speed Nin ^* data map for the normal running range in FIG. Determined and actually controlled.

The fluid coupling 12 interposed between the continuously variable transmission 14 and the auxiliary transmission 30 and the engine 10 for which the shift control is performed based on FIGS. 7 to 10 has a capacity coefficient shown in FIG. The fluid coupling 12 is changed by the change mechanism 300.
The capacity coefficient changing mechanism 300 is controlled by a deceleration control routine shown in FIG.

However, the turbine runner 302 and the pump impeller 304
Has the above-described shape, regardless of the following control, the capacity coefficient of the fluid coupling 12 becomes smaller when the engine 10 is in the driven state, for example, during engine braking than in the driven state.

The capacity coefficient changing mechanism 300 shown in FIG. 11 controls the internal pressure of the fluid coupling 12 including the turbine runner 302, the pump impeller 304, the direct coupling clutch 11, and the like. A known lock-up relay valve 306 for controlling the direction to directly connect the clutch 11, and a lock-up solenoid valve 7 for controlling the position of the lock-up relay valve 306.
7. A quick release valve 312 for rapidly releasing the operating oil pressure in the working chamber 310 of the fluid coupling 12 and a quick release solenoid valve 78 for controlling the position of the valve 312 are provided.

The lock-up relay valve 306 whose position is controlled by the lock-up solenoid valve 77 is a solenoid valve 77
Release the line oil pressure of the oil passage 316 to the drain, the spool 317 moves in the direction of the arrow YA to supply the line oil pressure of the oil passage 318 to the working chamber 310 via the oil passages 319 and 320, and the direct coupling clutch By releasing the hydraulic oil in the chamber 322 via the oil passages 324 and 326, the direct connection clutch 11 is directly connected (lock-up “on”). On the other hand, when the solenoid valve 77 is closed, the line oil pressure is
When added to valve chamber 328 via 6, spool 317
Moves in the direction of the arrow YA (position shown in FIG. 11) to supply the line oil pressure to the direct connection clutch chamber 322 via the oil passages 319 and 324, and to supply the hydraulic oil in the operation chamber 310 via the oil passages 320 and 330,
Release to the drain via the check valve 331 and the cooler 332 causes the direct connection clutch 11 to be in the lock-up “off” state.

The quick release valve 312, the position of which is controlled by the quick release solenoid valve 78, is configured such that when the solenoid valve 78 releases the pressurized oil supplied to the valve chamber 342 via the oil passage 340 to the drain, the spool 344 moves in the direction indicated by the arrow YA. To the working chamber 310
By releasing the working oil inside the drain through the oil passages 320 and 346 and the throttle 348, the internal pressure of the fluid coupling 12 held until then is reduced, and the capacity coefficient is reduced.

Quick release solenoid valve 7 of the capacity coefficient changing mechanism 300
In the deceleration control routine of FIG. 12 for controlling “8” “on” and “off”, first, the vehicle speed V is set to a predetermined vehicle speed V ₀ (here, “zero”).
It is determined whether or not the vehicle is traveling by determining whether or not the vehicle is running (step 400). If it is determined that the vehicle is running, it is then determined whether or not the throttle valve 62 is fully closed by determining whether or not the throttle opening θ is “zero” (step 410). If it is to be fully closed by the judgment whether fully closed, then by the speed ratio of the continuously variable transmission 14 gamma determines whether exceeds a predetermined value gamma _0, an engine brake feeling it is determined whether an area exceeding a predetermined speed ratio gamma ₀ to be unpleasant (step 420). The gear ratio γ is the input shaft speed N
Calculated based on in.

If the speed ratio γ is determined to be large by the determination of the speed ratio γ, it is next determined whether or not the rate of change V of the vehicle speed V is greater than a predetermined value V ₀ to determine whether the vehicle is in a deceleration state. Is determined (step 430). Here, when the vehicle is accelerating on a downhill or the like, that is, when it is not determined that the engine brake is applied, the sudden release solenoid valve 78 is then moved to the sudden release valve 312 by the operating oil in the working chamber 310. The control for switching to the release side (actual release valve operation) is actually performed to reduce the pressure in the working chamber 310 (step 440).

On the other hand, when the vehicle is stopped (V ≦ V ₀ , step 400), the throttle is open (θ ≠ 0, step 410), and the speed ratio γ is sufficiently small (γ ≦ γ ₀ , step 420), the vehicle accelerates. (≦ ₀ , step 430), control is performed to switch the quick release valve 312 to the side where the hydraulic oil is not released (quick release valve operation stop).
The pressure in the working chamber 310 is not actually reduced (step 450).

According to the above control of the present embodiment, as shown by the solid line in FIG. 13 showing the state of the turbine speed Nt in the deceleration state when the throttle is fully closed, for example, when the throttle opening θ is fully open and the vehicle is decelerating. When the gear ratio γ becomes larger than γ ₀ , that is, when the vehicle speed V becomes Von or less, the internal pressure of the fluid coupling 12 decreases and the capacity coefficient decreases. Thus, for example, when the vehicle speed is equal to or lower than Von, the engine speed Ne decreases, and the sense of effectiveness of the engine brake is reduced. Therefore, there is an extremely excellent effect that the driving feeling during deceleration is improved. In addition, according to the present embodiment, even if the speed change control of the continuously variable transmission 14 is performed at the same shift line at the time of starting and at the time of deceleration, the effect of improving the feeling of acceleration at the time of starting and the driving And an excellent effect that both of the above-mentioned effects can be obtained.

As described above, according to the present embodiment, when the engine is driven when the vehicle is decelerated, the capacity coefficient of the fluid clutch becomes small due to the shapes of the turbine runner and the pump impeller, and depending on the conditions, By releasing the hydraulic oil to the drain and reducing the internal pressure, the capacity coefficient is further reduced, and for example, the driven torque of the engine is reduced. As a result, an extremely excellent effect of reducing excessive engine braking feeling during deceleration and improving driving feeling is achieved.

In addition, the present embodiment can improve the feeling of acceleration at the time of starting and the function of improving the feeling of driving at the time of deceleration without changing the shift control of the continuously variable transmission only by controlling the fluid clutch. It has an extremely excellent effect.

Instead of the method of reducing the internal pressure to reduce the capacity coefficient in the above embodiment, a method of providing a baffle plate on the inner diameter of the pump impeller to prevent oil from flowing from the pump impeller to the turbine runner during turbine driving. Alternatively, the baffle may be configured to block the flow of oil only when the engine is driven. Further, the present invention may be applied to a vehicle with a torque converter or a vehicle with a stepped transmission to obtain a similar effect.

[Brief description of the drawings]

FIG. 1 is a block diagram illustrating the basic configuration of the present invention, and FIGS. 2 to 5 are explanatory diagrams showing the operation and effect of the present invention by taking a fluid clutch control of a vehicle with a continuously variable transmission as an example. ,
FIG. 6 is a block diagram of a system to which one embodiment of the present invention is applied, FIG. 7 is a flowchart of a gear ratio control routine of the embodiment, FIG. 8 is a flowchart of a gear ratio control routine in the normal traveling range, 9 is a graph showing shift characteristics of the subtransmission according to the embodiment, FIG. 10 is a graph showing shift characteristics of the continuously variable transmission according to the embodiment, FIG. 11 is a configuration diagram of a capacity coefficient changing mechanism of the embodiment, FIG. FIG. 12 is a flowchart of the deceleration control routine, and FIG. 13 is an explanatory diagram of its operation. 10 Engine 12 Fluid coupling 14 Continuously variable transmission 30 Auxiliary transmission 54 Controller 300 Capacity coefficient changing mechanism

──────────────────────────────────────────────────続 き Continued on the front page (58) Field surveyed (Int.Cl. ⁶ , DB name) F16D 33/06 F16D 33/20 B60K 41/02 F16H 61/64

Claims (1)

(57) [Claims] 1. A fluid coupling for a vehicle interposed between an engine of a vehicle and a transmission, wherein a flow direction of hydraulic oil flowing from a pump to a turbine when the engine is in a predetermined driven state. A pump having a forward direction with respect to the rotation of the turbine runner, and a flow direction of hydraulic oil flowing out of the turbine to the pump when the engine is in a predetermined driving state being opposed to the rotation of the pump impeller. By adopting the shapes of the impeller and the turbine runner, a fluid coefficient is reduced when the engine is in the driven state as compared with when the engine is in the driven state. .