Ratio change control for transmission

Abstract: A hybrid continuously variable transmission includes a gearing mechanism (38, 50, 42, 48) and a continuously variable transmission mechanism (16, 24, 26) which alternatively take over a drive. According to a ratio change control system, the continuously variable transmission mechanism (38, 50, 42, 48) is prevented from shifting until the transition from drive over the gearing mechanism to drive over the continuously variable transmission mechanism is completed.

Description: BACKGROUND OF THE INVENTION The present invention relates to a ratio change control for a transmission, and more particularly to a ratio change control for a hybrid continuously variable transmission including a gearing mechanism and a continuously variable transmission mechanism which are selectively rendered operable to take over a drive from a transmission input shaft to a transmission output shaft. US-A 4,735,113 discloses a V-belt type continuously variable transmission including a driver pulley, a follower pulley, and a V-belt drivingly interconnecting these pulleys. JP-A 63-176862 as well as EP-A 319044 disclose a hybrid continuously variable transmission as defined in the preamble of claim 1 and 4 respectively in which the above-mentioned continuously variable transmission mechanism is combined with a gearing mechanism such that the gearing mechanism is put into operation to provide a reduction ratio between a transmission input shaft and a transmission output shaft for start-up operation. The setting is such that this reduction ratio provided by the gearing mechanism is larger than the maximum or largest reduction ratio provided by the continuously variable transmission mechanism. For high speed operation where a relatively small drive force is required, the continuously variable transmission mechanism is put into operation to take over a drive from the transmission input shaft to the transmission output shaft owing to engagement of a power interruption device, such as a clutch. Referring to Fig. 7, let us now consider the case where the control strategy shown in U.S. Patent No. 4,735,113 is applied to the hybrid continuously variable transmission of the type mentioned above. Fig. 7 is a graph wherein a fully drawn line G shows target driver pulley revolution speed values versus varying vehicle speed values when a gearing mechanism is put into operation to take over a drive. In Fig. 7, a fully drawn line L shows target driver pulley revolution speed values versus vehicle speed values when a continuously variable transmission mechanism provides its maximum or largest reduction ratio, while a fully drawn line S shows target driver pulley revolution speed values versus vehicle speed values when the continuously variable transmission mechanism provides its minimum or smallest reduction ratio. Let it be assumed that a vehicle starts moving with a 3/8 throttle opening degree. In this case, the driver pulley revolution speed increases along the line G from a point O to a point A with the reduction ratio provided by the gearing mechanism. Upon arriving at the point A, a transition is made from the drive owing to the gearing mechanism to a drive owing to the continuously variable transmission. This causes a drop in driver pulley revolution speed without any substantial change in vehicle speed, resulting in a transfer from the point A to a point B on the line L. Then, with the maximum reduction radio provided by the continuously variable transmission mechanism, the vehicle increases its speed till a point C on the line L. As the vehicle increases its speed further from a vehicle speed value corresponding to the point C, the reduction ratio decreases continuously toward the minimum reduction ratio provided by the continuously variable transmission mechanism along an operation line I. According to this control strategy, the continuously variable transmission mechanism effects a ratio change from the maximum reduction ratio to a relative smaller reduction ratio on a broken line M since a target driver pulley revolution speed corresponding to the point C′ is set if the vehicle speed increases upto a vehicle speed value corresponding to the point C′. A problem arises if, during transition from the drive owing to the gearing mechanism to the drive owing to the continuously variable transmission mechanism, the vehicle speed increases beyond the vehicle speed value corresponding to the point C and the continuously variable transmission has shifted off the maximum reduction ratio thereof. The transition in drive takes a relatively long time since there is a delay in hydraulic system for activating a clutch contributing to this transition. In this case, the vehicle starts running through the drive owing to the continuously variable transmission mechanism with a reduction ratio smaller than the maximum reduction ratio thereof. This transition is inherent with a substantial shock. This may cause an insufficient driving force and thus poor acceleration. Another hybrid continuously variable transmission is known. This known hybrid continuously variable transmission has a gearing mechanism which provides a reduction radio smaller than the minimum or smallest reduction ratio provided by a continuously variable transmission mechanism. Such a hybrid continuously variable transmission is disclosed in Laid-open Japanese Patent Application No. 58-156764. If the continuously variable transmission mechanism is controlled according to the above-mentioned control strategy, there occurs a case where the continuously variable transmission mechanism shifts down from the minimum reduction ratio thereof during a transition from a drive owing to the gearing mechanism to a drive owing to the continuously variable transmission mechanism. This causes a substantial increase in engine speed, inducing a substantial shock. A ratio change control according to the preamble of claim 1 and a method of a ratio change control for transmission according to the preamble of claim 4 are also disclosed in EP-A-0,336,323. This document which is a document in the sense of Article 54(3) EPC, describes a control system for transmission which includes a gearing mechanism combined with a V-belt type of continuously variable transmission mechanism. A low clutch and a high clutch are provided, which are arranged in such a way that when the low clutch is engaged, torque is transmitted by the gearing mechanism, while when the high clutch is engaged, torque is transmitted by the V-belt CVT. A valve is provided for supplying hydraulic fluid to the low clutch when the maximum reduction ratio of the V-belt type CVT is achieved. It is the object of the present invention to provide a ratio change control and a method for a ratio change control for transmission which assures smooth and shockless transition between a drive owing to a gearing mechanism and a drive owing to a continuously variable transmission mechanism. This object is solved according to the present invention by the subject-matter of claim 1. The method according to the invention is the subject-matter of claim 4. Preferred embodiments of the invention are the subject-matter of the dependent claims. BRIEF DESCRIPTION OF THE DRAWINGS Referring to the accompanying drawings, Fig. 1 is a schematic diagram of a hybrid transmission including a gearing mechanism which is combined with a continuously variable transmission mechanism such that the gearing mechanism takes over a drive to provide a reduction ratio larger than the maximum reduction ratio provided by the continuously variable transmission mechanism; Fig. 2 is a diagram illustrating arrangement of a variety of different shafts of the transmission; Fig. 3 is a hydraulic circuit of a control system for the hybrid transmission; Fig. 4 is a block diagram of a control unit associated with sensors and a shift motor; Fig. 5 is a flowchart of a program contained in the control unit; Fig. 6 is a functional block diagram; Fig. 7 a graphical representation of a control strategy discussed above; Fig. 8 is a similar view to Fig. 1 showing a hybrid transmission including a gearing mechanism which is combined with a continuously variable transmission such that the gearing mechanism takes over a drive to provide a reduction ratio smaller than the minimum reduction ratio provided by the continuously variable transmission; Fig. 9 is a similar view to Fig. 2 illustrating in the fully drawn line rotary elements which play a role when the continuously variable transmission mechanism takes over a forward drive in the hybrid transmission shown in Fig. 8; Fig. 10 is a similar view to Fig. 9 illustrating in the fully drawn line rotary elements which play a role when the gearing mechanism takes over a forward drive in the hybrid transmission shown in Fig. 8; Fig. 11 is a similar view to Fig. 9 illustrating in the fully drawn line rotary elements which play a role when the gearing mechanism takes over a reverse drive in in the hybrid continuously variable transmission shown in Fig. 8; and Fig. 12 is a flowchart similar to Fig. 5. DETAILED DESCRIPTION OF THE EMBODIMENT Referring to the accompanying drawings, a preferred embodiment of a control system according to the present invention is described. Referring now to Fig. 1, a power train of a motor vehicle is described. In Fig. 1, an engine 10 is shown as having an output shaft 10a which a torque converter 12 is coupled with in the conventional manner. The torque converter 12 includes, as usual, a pump impeller 12a, a turbine runner 12b, and a stator 12c. It also includes a lock-up clutch 12d which the pump impeller 12a and turbine runner 12b are selectively interconnected with. The turbine runner 12b of the torque converter 12 is drivingly connected to a turbine shaft or a driver shaft 14. On the driver shaft 14 is a driver pulley 16. The driver pulley 16 includes an axially stationary conical member 18 fixedly connected to the driver shaft 14, and an axially moveable conical member 22 connected to the driver shaft 14 in opposed spaced relationship with the stationary conical member 18. The conical members 18 and 22 define therebetween a V-shaped pulley groove. The driver pulley 16 includes a driver pulley cylinder chamber 20. The moveable conical member 22 is axially moveable toward the stationary conical member 18 in response to hydraulic pressure developed in the driver pulley cylinder chamber 20, causing the V-shaped pulley groove to decrease its width. The driver pulley 16 is drivingly connected via a V-belt 24 to a follower pulley 26. The follower pulley 26 includes an axially stationary conical member 30 fixedly connected to a follower shaft 28, and an axially moveable conical member 34 connected to the follower shaft 28 for axial movement. The conical members 30 and 34 define a V-shaped pulley groove therebetween. The follower pulley 26 includes a follower pulley cylinder chamber 32. The moveable conical member 34 is axially moveable toward the stationary conical member 30 in response to hydraulic pressure developed in the follower pulley cylinder chamber 32. The driver pulley 16, V-belt 24, and follower pulley 26 form a continuously variable transmission mechanism. The setting is such that the maximum and largest reduction radio provided by this continuously variable transmission mechanism only is smaller than a reduction ratio provided by a gearing mechanism including a driver shaft side forward gear 42 and an output shaft side forward gear 48, which are described later. The driver shaft 14 extends through a hollow shaft 36. The hollow shaft 36 is rotatably supported on the driver shaft 14. Rotatably supported on the hollow shaft 36 are a reverse gear 38 and a forward gear 42. The forward gear 42 is connectable to the hollow shaft 36 by means of a hydraulic fluid operated forward clutch 52, while the reverse gear 38 is connectable to the hollow shaft 36 by means of a hydraulic fluid operated reverse clutch 53. With a hydraulic fluid operated low clutch 44 engaged, the driver shaft 14 is connected to the hollow shaft 36. Extending in parallel to the driver shaft 14 is an output shaft 46. The forward gear 14 is mounted via a one-way clutch 40 to the output shaft 46, while a reverse gear 50 is mounted for unitary rotation with the output shaft 46. The forward gear 48 is in constant mesh with the forward gear 42. The reverse gear 50 is in constant mesh with a reverse idler gear 56 that is rotatable with an idler shaft 54. The reverse idler gear 56 is in constant mesh with the reverse gear 38, too. In Fig. 1, since it is impossible to illustrate them in the cross sectional plane, the reverse idler shaft 54 and reverse idler gear 56 are illustrated by the broken line. Actually, they are arranged as illustrated in Fig. 2. In Fig. 1, the distance between the shafts and the diameter of each of the gears do not reflect the actual dimension. Thus, in order to known the actual relationship, reference should be made to Fig. 2. The follower shaft 28 has a forward gear 58. Via a hydraulic fluid operated high clutch 60, the forward gear 58 is connectable to the follower shaft 28. As best seen in Fig. 2, the forward gear 58 is in constant mesh with the reverse gear 50. The forward gear 58 and reverse gear 50 have the same diameter. The output shaft 46 has a reduction gear 62 for rotation therewith. The reduction gear 62 is in constant mesh with a final gear 64 of a differential 66. The differential 66 includes a pair of pinion gears 68 and 70 which are rotatable with the final gear 64. A pair of side gears 72 and 74 mesh with the pinion gears 68 and 70. The side gears 72 and 74 are coupled with drive axles 76 and 78, respectively, for rotation therewith. The neutral state is provided when the low clutch 44 and the high clutch 60 are both released. In this state, the transmission of rotational power from the driver shaft 14 to the output shaft 46 is interrupted. On start-up or hill-climbing where a relatively large driving force is required, the forward clutch 52 is engaged and the low clutch 44 engaged. The high clutch 60 is released. In this state, the rotational power of the output shaft 10a of the engine 10 is transmitted via the torque converter 12 to the driver shaft 14, and further to the hollow shaft 36 via the low clutch 44 that is engaged. The torque of the hollow shaft 36 is transmitted via the forward clutch 52 to the forward gear 42, and further to the forward gear 48 which the gear 42 meshes with. Owing to the fact that the forward gear 48 is drivingly connected via the one-way clutch 40 to the output shaft 46, the rotational power is transmitted to the output shaft 46. Thereafter, the rotational power is transmitted via the reduction gear 62 and the final gear 64 to the differential 66 where it is distributed between the drive axles 76 and 78, causing road wheels of the vehicle, not illustrated, to rotate. During the transmission of rotational power mentioned above, the rotational power is not transmitted through the continuously variable transmission mechanism, but through the gearing mechanism. With the reduction ratio provided by the intermeshed forward gears 42 and 48, the rotational power is transmitted to the output shaft 46, thus providing a relatively large driving force. When the operating condition progresses and now demands a less driving force, the high clutch 60 is engaged with the above described state maintained. This causes the rotational power to be transmitted through the continuously variable transmission. The rotational power of the driver shaft 14 is transmitted, via the V-belt 24 and the follower pulley 26, to the follower shaft 28, and further to the forward gear 58 via the high clutch 60 that is engaged. Since the forward clutch 58 meshes with the reverse gear 50, the rotational power is transmitted to the output shaft 46, and further to the drive axles 76 and 78 via the same power delivery path as previously described. In this case, the output shaft 46 rotates at a higher speed that the forward gear 48 does, and thus the one-way clutch 40 idles. This allows the low clutch 44 to be kept engaged. In the manner as described above, the rotational power is transmitted through the continuously variable transmission mechanism. Thus, the reduction ratio can be varied continuously by varying the width of the V-groove of the driver pulley 26 which in turn induces variation in the width of the V-shaped groove of the follower pulley 26. For reverse drive, the reverse clutch 53 is engaged, the low clutch 44 is engaged, and the high clutch 60 is released. The engagement of the reverse clutch 53 causes the reverse gear 38 to be connected to the hollow shaft 36 for unitary rotation. In this state, the rotational power of the driver shaft 14 is transmitted via the low clutch 44, the hollow shaft 36, the reverse clutch 53, the reverse gear 38, the reverse idler gear 56 and the reverse gear 50 to the output shaft 46. Since the reverse idler gear 56 is operatively disposed in the power delivery path, the direction of rotation of the output shaft 46 is the opposite to the direction of rotation of the output shaft 46. Thus, the vehicle can move in the reverse direction. A hydraulic control system for the hybrid continuously variable transmission is now described. As shown in Fig. 3, the control system comprises an oil pump 101, a line pressure regulator valve 102, a manual valve 104, a shift control valve 105, a direction control valve 108 (which may be called as a shift command valve) for controlling direction of adjustment pressure within a hydraulic fluid line 190, a shift operating mechanism 112, a throttle valve 114, a constant pressure regulating valve 116, a solenoid valve 118, a torque converter pressure regulating valve 120, and a lock-up control valve 122. The shift control valve 106 has a valve bore 172 provided with five ports 172a, 172b, 172c, 172d and 172e, a spool 174 having three axially spaced lands 174a, 174b, and 174c slidably fit in the valve bore 172, and a spring 175 biasing the spool 174 to the left as viewed in Fig. 3. The port 172b communicates via a hydraulic fluid conduit 176 with the driver pulley cylinder chamber 20, and this conduit 176 communicates with the high clutch 60 at its servo chamber. The port 172a and the port 172e are drain ports, respectively. An orifice 177 is provided at the drain port 172a. The port 172d communicates via a hydraulic fluid conduit 179 with the follower pulley cylinder chamber 32. The port 172c communicates with a hydraulic fluid conduit 132 that serves as a line pressure circuit and thus is supplied with the line pressure. The spool 174 has a lefthand end, as viewed in Fig. 3, rotatably linked via a pin 181 to a middle portion of a lever 178 of the shift operating mechanism 112 which is later described in detail. The land 174b has an axial section with a curved contour. This allows a portion of hydraulic fluid supplied from the line pressure port 172c to flow into the port 172a. Thus, the pressure at the port 172b is determined by a ratio of the amount of hydraulic fluid flowing from the port 172c toward the port 172b to the amount of hydraulic fluid discharged out of the drain port 172a. If the spool 174 moves to the left as viewed in Fig. 3, this leftward movement of the spool 174 causes the degree of opening of a clearance on the line pressure side of the port 172b to increase, and the degree of opening of a clearance on the discharge side of the port 172b to decrease. This results in an increase in pressure at the sport 172b. The port 172d is always supplied with the line pressure from the port 172c. The hydraulic presure developed at the port 172b is supplied via the conduit 176 to the driver pulley cylinder chamber 20, while the hydraulic pressure developed at the port 172d is supplied to the follower pulley cylinder chamber 32. Therefore, the leftward movement of the spool 174, as viewed in Fig. 3, causes an increase in the hydraulic pressure developed in the driver pulley cylinder chamber 20, resulting in a decrease in the width of the V-shaped pulley groove of the driver pulley 16. This also results in an increase in the width of the V-shaped pulley groove of the follower pulley 26 since the V-belt 26 is wedged into the V-shaped groove of the follower pulley 26. Therefore, the reduction ratio becomes small since the radius of the running diameter of the V-belt on the driver pulley 16 increases, but the radius of the running diameter of the V-belt 24 on the follower pulley 26 decreases. The reduction ratio becomes large when the spool 174 is urged to move to the right as viewed in Fig. 3. The lever 178 of the shift operating mechanism 112 has its middle portion linked via a pin pin 181 to the spool 174 of the shift control valve 106. The lever 178 has one or lower end, as viewed in Fig. 3, linked via a pin 183 to a reduction ratio transmission member 158 and the opposite or an upper end linked via a pin 185 to the rod 182 of the direction control valve 108. The rod 182 is formed with a rack 182c which a pinion gear 110a of a shift motor 110 in the form of a stepper motor meshes with. According to this shift operating mechanism 112, rotating the pinion gear 110a of the shift motor 110 in such a direction as to displace the rod 182 to the right, as viewed in Fig. 3, causes the lever 178 to swing clockwise about the pin 183. This clockwise movement of the lever 178 causes the spool 174 of the shift control valve 106 to move to the right, as viewed in Fig. 3. The rightward movement of the spool 174 of the shift control valve 106 causes a reduction in hydraulic pressure within the driver pulley chamber 20, causing the axially moveable conical member 22 of the driver pulley 16 in such a direction as to increase the width of the V-shaped pulley groove. Viewing in Fig. 3, the conical member 22 moves to the left. Since the end of the lever 178 is connected via the pin 183 to the reduction ratio transmission member 158, this leftward movement of the conical member 22, as viewed in Fig. 3, causes the lever 178 to swing clockwise about the pin 185. This clockwise movement of the lever 178 displaces the spool 174 to the left, as viewed in Fig. 3, causing an increase in hydraulic pressure within the driver pulley chamber 20, thus causing the reduction ratio to become small. As a result, the driver pulley 16 and the follower pulley 26 assume one of stable states, providing one reduction ratio. Each stable state corresponds to one rotary position which the shift motor 110 takes. If the shift motor 110 rotates in the opposite or reverse direction, the above-mentioned process progresses. The rod 182 is moveable to the right, as viewed in Fig. 3, beyond a position at which the maximum and largest reduction ratio is induced toward an overstroke position. When the rod 182 has moved toward and stayed at the overstroke position, a change-over switch 298 is activated. The output signal of this switch 298 is fed to a control unit. It will now be recognized that when the shift motor 110 is operated in a predetermined shift pattern, the variation of the reduction ratio follows this pattern. Therefore, the variation in reduction ratio in the continuously variable transmission mechanism is controlled by controlling the shift motor 110. The rotary position which the shift motor 110 takes is determined by a number of pulses supplied to the shift motor 110 by a control unit 300 shown in Fig. 4. The control unit 300 stores a plurality of shift patterns and generates the number of pulses in accordance with one shift pattern selected out of all. The direction control valve 108 includes a valve bore 186 provided with ports 186a, 186b, 186c and 186d, and a rod 182 with lands 182a and 182b received in the valve bore 186. The port 186a communicates with a hydraulic fluid conduit 188. The port 186b communicates via a hydraulic fluid conduit 190 with the solenoid valve 118. The port 186c communicates with a hydraulic fluid conduit 189. The port 186d is a drain port. Normally, the ports 186a and 186b communicate with each other via a space defined between the lands 182a and 182b. When the rod 182 moves beyond the position corresponding to the maximum reduction ratio toward the overstroke position, the port 186a is covered by the land 182a, while the port 186b is allowed to communicate with the port 186c. The above-mentioned hydraulic fluid conduit 189 communicates with the low clutch 44. The other valves illustrated in Fig. 3 are substantially the same as their counterparts disclosed in JP 61-105351. The hydraulic circuit except the low clutch 44 and the high clutch 60 is substantially the same as a hydraulic control system disclosed in European Patent Application published under publication number 0180209 on May 7, 1986 or United States Patent No. 4,735,113 issued to Yamamuro et al. on April 5, 1988. Referring to Fig. 4, the control unit 300 comprises an input interface 311, a reference pulse generator 312, a central processor unit (CPU) 313, a read only memory (ROM) 314, a random access memory (RAM) 315 and an output interface 316. They are operatively interconnected by an address bus 319 and a data bus 320. Output signals of an engine speed sensor 305, a vehicle speed sensor 302, a throttle sensor 303, a shift position switch 304, a turbine speed sensor 305, an engine coolant temperature sensor 306, a brake sensor 307, and a change-over switch 298. These output signals are supplied directly or indirectly via wave shapers 308, 309, and 322, and an AD converter 310. Output signals of the control unit 300 are supplied via an amplifier and leads 317a, 317b, 317c and 317d to the shift motor 110. The solenoid 224 is also under the control of the control unit 300. For further understanding of the control unit 300, reference should be made to U.S. Patent No. 4,735,113 mentioned before. The operation of this embodiment is described. When the rotation of motor 110 toward the maximum and largest reduction ratio causes the rod 182 to move beyond the predetermined position corresponding to the maximum reduction ratio to the overstroke position, the direction control valve 108 takes a position as illustrated by the lower half thereof viewing in Fig. 3. In this state of the direction control valve 108, the conduit 190 is allowed to communicate with the conduit 189, establishing a state where the hydraulic fluid pressure adjusted by the solenoid valve 118 is supplied to the low clutch 44. Thus, the torque capacity of the low clutch 44 becomes adjustable by the solenoid valve 118. Upon the direction control valve 108 staying in the overstroke position, the spool 174 of the shift control valve 106 which is linked via the lever 178 to the rod 182 is displaced to the right as viewed in Fig. 3 and thus the port 172b is allowed to communicate with the drain port 172a. Thus, the driver pulley cylinder chamber 20 and high clutch 60 are drained. As a result, the high clutch 60 is released with the low clutch 44 left engaged and thus the gearing mechanism takes over a drive between the driver shaft 14 and the output shaft 46. The torque is therefore multiplied at the reduction ratio determined by the forward gears 42 and 48. When the operating condition progresses and demands a less driving force, the shift motor 110 rotates toward the minimum and smallest reduction ratio, the rod 182 moves toward the minimum reduction ratio beyond the predetermined position corresponding to the maximum reduction ratio. This movement of the rod 182 causes the lever 178 to displace the spool 174 to the left as viewed in Fig. 3. This causes the land 174b of the spool 174 to close communication of the port 172b with the drain port 172a and open communication of the port 172b with the line pressure port 172c, allowing the supply of hydraulic fluid to the port 172b from the line pressure port 172c. Thus, the driver pulley cylinder chamber 20 is supplied with the hydraulic fluid pressure and the high clutch 60 is also supplied with the hydraulic fluid pressure and engaged. Engagement of the high clutch 60 causes a transition from a drive owing to the gearing mechanism to a drive owing to the continuously variable transmission mechanism. Upon completion of the transition, the continuously variable transmission takes over a drive between the shafts 14 and 46. Thereafter, the rotational power is transmitted by the continuously variable transmission mechanism and the reduction ratio is allowed to continuously vary. Under this operating condition, although the low clutch 40 is kept engaged, since the one-way clutch 40 idles, the rotational power is transmitted via the continuously variable transmission mechanism. The direction control valve 108 allows the conduit 188 to communicate with the conduit 190, rendering the solenoid valve 118 to control the lock-up control valve 122. Thus, the engagement of the lock-up clutch 12d is controlled by the solenoid valve 118. The valve 108 assumes the overstroke position when the spool 136 of the manual valve 104 is placed at P or R or N position. Since the valve 108 is in the overstroke position and the solenoid valve 118 drains the conduit 190 when the spool 136 of the manual valve 104 is placed at P or N position, no hydraulic fluid pressure is applied to the low clutch 44, leaving same released. Referring to the flowchart shown in Fig. 5, it is now explained how the control unit 300 operates during transition from the drive owing to the gearing mechanism to the drive owing to the continuously variable transmission mechanism. The ROM 314 of the control unit 300 stores a program as illustrated by the flowchart in Fig. 5. In Fig. 5, at a step 400, a reading operation is performed based on the output signal of the throttle sensor 303 to store the result as a throttle opening degree TH. At the subsequent step 402, a reading operation is performed based on the output signal of the vehicle speed sensor 302 to store the result as a vehicle speed V. At a step 404, a reading operation is performed based on the output signal of the turbine speed sensor 305 to store the result as an actual driver pulley revolution speed N a . At a step 406, a table look-up operation of a shift point mapping data stored in the ROM 314 is performed using the stored throttle opening degree TH and vehicle speed V in order to determine and store the result as a target driver pulley revolution speed N t . At a step 408, using the data N t and V stored at steps 406 and 402, an equation I t = N t /V is calculated to store the result as a target reduction ratio I t . At a step 410, using the data N a and V stored at steps 404 and 402, an equation I r = N a /V is calculated in order to determine and store the result as an actual reduction ratio I r . At a step 412, it is determined whether or not the actual reduction ratio I r is less than or equal to a reference value I c corresponding to the maximum or largest reduction ratio provided by the continuously variable transmission mechanism. If the gearing mechanism still takes over a drive or the transition in drive from the gearing mechanism to the continuously variable transmission mechanism has not been completed yet, the actual reduction ratio I r is greater than the maximum reduction ratio I c , an answer to the inquiry at the step 412 is NO, and thus the program proceeds to a step 414. At the step 414, the target reduction ratio I t is set equal to the maximum reduction ratio I c . At the subsequent step 416, a number of steps along which a stepper motor drive signal should be moved to establish the target reduction ratio I t , namely the maximum reduction ratio I c in this instance, is determined and the result is stored as a step number M. Then, at a step 418, the stepper motor drive signal is moved along the number of steps M. Thus, in this instance, the continuously variable transmission mechanism is kept at the maximum reduction ratio thereof. In other words, a shift operation in the continuously variable transmission mechanism is prevented. Upon completion of the transition into the drive owing to the continuously variable transmission mechanism, the actual reduction ratio I r becomes equal to the maximum reduction ratio I c , an answer to the inquiry at the step 412 becomes YES, and the program proceeds from the step 412 directly to the step 416. In this instance, the data determined at the step 408 is used as it is, and the continuously variable transmission is now allowed to shift. From this explanation along with the flowchart shown in Fig. 5, it is now appreciated that the continuously variable transmission mechanism is prevented from shifting from the maximum reduction ratio thereof until the transition from the drive owing to the gearing mechanism to the drive owing to the continuously variable transmission mechanism is completed. Referring to Fig. 6, a control concept according to the present invention is described. For this control, the outputs of the throttle sensor 303 and vehicle speed sensor 302 are supplied to a functional block 500 where the target driver pulley revolution speed N t for a throttle opening degree TH indicated by the output of the throttle sensor 303 and a vehicle speed V indicated by the output of the vehicle speed sensor 302 is determined. The output N t of this block 500 is supplied to a block 502 where a target reduction ratio I t is determined by calculating an equation I t = N t /V . The output I t passes through a gate 504 unmodulated to a block 506 where a stepper motor pulse number M is determined. The output of this block 506 is supplied to the shift motor 110. The output of the vehicle speed sensor 302 and the output of the turbine revolution speed sensor 305 are supplied to a block 508. The output of the turbine revolution speed sensor 305 is indicative of a driver pulley revolution speed N a . At the block 508, an actual reduction ratio I r is determined by calculating an equation I r = N a /V . A reference value I c corresponding to the maximum or largest reduction ratio provided by the continuously variable transmission mechanism is set at a reference setting block 510. The output of the block 510 indicative of the reference value I c and the output of the block 508 indicative of the actual reduction ratio I r are supplied to a comparator 512 where it is determined whether I r is less than or equal to the reference value I c or not. If the actual reduction ration I r is still larger than the reference value I c and thus the transition from the drive owing to the gearing mechanism to the drive owing to the continuously variable transmission mechanism has not been completed, the gate 504 shifts in response to the result S of comparison in the block 512 to a state where the target reduction ratio I t is set equal to I c . Thus, the reference value I c is always supplied to the block 506, causing the shift motor 110 to keep the continuously variable transmission mechanism at the maximum reduction ratio, thus preventing the shifting operation of the continuously variable transmission mechanism. Of course, if the actual reduction ratio I r becomes equal to or less than the reference value I c , the gate 504 allows the passage of the target reduction ratio I t to the block 506, thus allowing the continuously variable transmission mechanism to shift to the target reduction ratio I t . The present invention is applicable to a hybrid transmission including a gearing mechanism which is combined with a continuously variable transmission mechanism such that the gearing mechanism takes over a drive to provide a reduction ratio smaller than the minimum reduction ratio provided by the continuously variable transmission mechanism. One example of such a hybrid transmission is illustrated in Figs. 8 through 11. A hybrid transmission shown in Fig. 8 is substantially the same as the previously described hybrid transmission shown in Fig. 1 except the fact that the sizes of gears 58A, 50A, 42A, 48A, 38A, and 54A are different from their counterparts 58, 50, 42, 48, 38, and 54. Fig. 9 illustrates in the fully drawn line that the pulley unit 16, 24, and 26 and the gears 58A, 50A, 62, and 64 play a role when the continuously variable transmission takes over a forward drive in the hybrid transmission shown in Fig. 8. Fig. 10 illustrates in the fully drawn line that the gears 42A, 48A, 62, and 64 play a role when the gearing mechanism takes over a forward drive in the hybrid transmission shown in Fig. 8. When it takes over the drive, the gearing mechanism provides a reduction ratio smaller than the minimum or smallest reduction ratio provided by the continuously variable transmission. Fig. 11 illustrates in the fully drawn line that the gears 38A, 54A, 50A, 62, and 64 play a role when the gearing mechanism takes over a reverse drive in the hybrid transmission shown in Fig. 8. The present invention is embodied in this hybrid transmission such that the continuously variable transmission mechanism is prevented from shifting from the minimum reduction ratio until a transition from the forward drive owing to the gearing mechanism to the drive owing to the continuously variable transmission mechanism is completed. This control is illustrated in the flowchart shown in Fig. 12. Referring to Fig. 12, the flowchart is substantially the same as the flowchart shown in Fig. 5 except the provision of steps 412A and 414A in lieu of the before mentioned steps 412 and 414. In Fig. 12, it is determined at the step 412A whether an actual reduction ratio I r is greater than or equal to a reference value I c , corresponding to the minimum or smallest reduction ratio provided by the continuously variable transmission mechanism. If an answer to the inquiry at the step 412A is NO, that is the gearing mechanism still takes over the drive or the transition to the consinuously variable transmission mechanism has not been completed yet, a target reduction ratio I t is set equal to the above-mentioned reference value I c ′ at the step 414A. Thus, the continuously variable transmission mechanism is kept at the minimum reduction ratio until the transition from the drive owing to the gearing mechanism to the drive owing to the continuously variable transmission mechanism is completed.

Claims: A ratio change control for a transmission, the transmission including a gearing mechanism and a continuously variable transmission mechanism, the gearing mechanism and continuously variable transmission mechanism being operatively combined with each other such that they alternatively take over a drive, characterized in that a comparator (512) is provided for comparing an actual reduction ratio (Ir) with a reference value (Ic or Ic') corresponding to a predetermined reduction ratio which limits a range of reduction ratios available by the continuously variable transmission mechanism (step 412 or 412A); means are provided for detecting whether a transition from a drive owing to the gearing mechanism to a drive owing to the continuously variable transmission mechanism is completed or not in response to said result of comparison of the actual reduction ratio (Ir) with the reference value (Ic or Ic'); and means for preventing shifting operation is provided to keep the continuously variable transmission mechanism at the predetermined reduction ratio until the transition is completed (steps 412, 414 and 416 or steps 412A, 414 and 416). A ratio change control as claimed in claim 1, characterized in that the predetermined reduction ratio is the maximum of reduction ratios which are established by the continuously variable transmission mechanism if the gearing mechanism provides a reduction ratio larger than the maximum reduction ratio. A ratio change control as claimed in claim 1, characterized in that the predetermined reduction ratio is the minimum of reduction ratios which are established by the continuously variable transmission mechanism if the gearing mechanism provides a reduction ratio smaller than the minimum reduction ratio. A method of a ratio change control for a transmission, the transmission including a gearing mechanism and a continuously variable transmission mechanism, the gearing mechanism and continuously variable transmission mechanism being operatively combined with each other such that they alternatively take over a drive, the ratio change control, characterized by: comparing an actual reduction ratio (Ir) with a reference value (Ic or Ic') corresponding to a predetermined reduction ratio which limits a range of reduction ratios available by the continuously variable transmission mechanism (step 412 or 412A); determining whether a transition from a drive owing to the gearing mechanism to a drive owing to the continuously variable transmission mechanism is completed or not in response to result of comparison of the actual reduction ratio (Ir) with the reference value (Ic or Ic'); and keeping the continuously variable transmission mechanism at the predetermined reduction ratio until the transition is completed (steps 412, 414 and 416 or steps 412A, 414 and 416). A method as claimed in claim 4, characterized in that the predetermined reduction ratio is the maximum of reduction ratios which are established by the continuously variable transmission mechanism if the gearing mechanism provides a reduction ratio larger than the maximum reduction ratio. A method as claimed in claim 4, characterized in that the predetermined reduction ratio is the minimum of reduction ratios which are established by the continuously variable transmission mechanism if the gearing mechanism provides a reduction ratio smaller than the minimum reduction ratio.

Inventors: HIBI TOSHIFUMI, HIBI, TOSHIFUMI
Applicants: NISSAN MOTOR, NISSAN MOTOR CO., LTD.
Application year: 1990
Main classification: F16H61
IPC codes: F16H61, F16H37