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Technical Field

This invention relates to drivelines for automotive vehicles having infinitely variable torque multipli- cation characteristics with an extended effective torque multiplication range.

Background Art and General Description

My invention is adapted to be used in an infinitely variable belt drive of the kind shown, for example, in U.S. Patent No. 3.115,049, which is assigned to the assignee of my invention. The belt drive shown in that patent is used with an internal combustion engine when the driving pulleys are located on the engine crankshaft axis and the output shaft is located on a second axis arranged in parallel relationship with respect to the crankshaft axis. A belt loading servo controls the driving pulleys to cause a clamping force on the belt, which is related to the torque transmitting capacity of the drive. A ratio control servo located on the parallel axis controls the effective pitch diameter of both pul-leys. As the speed of the vehicle increases relative to the engine speed, the pitch diameter of the driving pulleys increases and the pitch diameter of the driven pulleys decreases. The effective torque ratio range that can be achieved with a drive of the kind disclosed in this patent is determined by the relative pitch diameters of the pulleys. This is true also of other prior art patents of which I am aware, such as U.S. Patent No.
2.829.533; 3,081,642; 3.017.785 and Reissue Patent No. 31.361.

Some of the prior art patents mentioned here use* a forward and reverse gear on the output shaft axis.
Reference may be made, for example, to Figure 1 of Patent No. 2,829,533 and Figure 3 of Reissue Patent No. 31.361 for a teaching of a forward and reverse gear mechanism between the driven pulley and the final drive.
Other examples of an infinitely variable belt drives for an automotive vehicle wherein provision is made for achieving reverse and forward drive shown in U.S. Patent Nos. 4.060.012; 4.317.389; 4,304,150 and
For a description of commercial embodiment of the Fiat/Van Doorne transmatic transmission employing a Van Doorne steel belt, reference may be made to Road a d Track magazine for August 1983, pgs. 78, 79 and 80.

Brief Description of the Invention

My invention comprises a dual range, continuously variable transmission with an automatic ratio change feature capable of providing an expanded ratio range to achieve maximum fuel economy during operation of the vehicle under road load conditions while maintaining the necessary startup speed ratio for high acceleration from a standing start. Provision is made in my invention for making a nonsynchronous, overdrive-to-low range shift. The shift is characterized by relatively close ratios, spaced apart about 25 percent, thereby improving the quality of the shift by reducing harshness due to transient inertia forces.
My invention comprises also a hydrokinetic coupling located between the driving pulleys and the engine to provide smoothness during acceleration. A fluid coupling bypass friction clutch is used after the vehicle achieves a steady state road load condition so that the hydrokinetic efficiency losses inherent in a fluid coupling are eliminated.

Description of the Figures of the Drawings

Figure 1A is a partial, cross-sectional assembly- view of my infinitely variable, dual range transmission showing a torque converter and a driving pulley mechanism.
Figure IB is a partial, cross-sectional view of the transmission illustrated in Figure 1A including the ratio changing clutches, the torque multiplication gearing and the driven pulley mechanism.
Figure 1C is a partial, cross-sectional view of the transmission illustrated in Figures 1A and IB showing the differential mechanism situated on the torque output side of the transmission structure.
Figure 2 is a schematic representation of the transmission assembly illustrated in Figures 1A. IB and 1C. the overall cross-sectional view of the transmission assembly being a composite of the partial cross-sectional views of Figures 1A, IB and 1C.

Best Mode for Carrying Out the Invention

In Figure 1A, numeral 10 designates one end of the crankshaft of an internal combustion engine. Numeral 12 designates a hydrokinetic coupling having a bladed impeller 14 and a bladed turbine 16. Coupling 12 is located within a coupling housing 18 which is secured at the housing margin 20 by bolts, not shown, to the engine block of the internal combustion engine.
A pulley housing 22 is secured at its right end 24 to the housing 18. The left end of the housing 18 receives a closure plate 26.

Housing 18 is provided with an end wall 28 having an impeller support sleeve 30. The hub 32 of the impeller 14 is journalled by bushing 34 on the support sleeve 30. The interior of the coupling is isolated from the exterior of the impeller housing by a running fluid seal 36.
The impeller housing is designated by reference character 38. It includes a vertical wall 40 between the engine and the coupling. This wall is secured by bolts 42 to drive plate 44, which in turn is connected drivably to the crankshaft 10 as shown at 46. The wall 40 has a hub 48 which is splined at 50 to pump driveshaft 52, the latter extending transversely in coaxial relationship with respect to the impeller. It is splined at 54 to pump rotor 56 of a positive displacement pump located in a pump housing formed in the end plate 26. Shaft 52 is journalled in boss 58 by bearing 60. A control valve body can be formed in the housing 22 at the upward portion of the transmission assembly.
The bearing boss 58 forms a part of a transmission casing end wall 62 which is provided with a bearing opening 64 for bearing assembly 66. Hub 68 of cone pulley disc 70 for a pulley assembly, generally indicated at 72. is journalled by bearing assembly 66. A com-panion cone disc 74 for the assembly 72 is internally splined at 76 to external splines 78 formed on pulley shaft 80. the latter being connected to, or forming a part of, the cone disc 70.
The cone pulley disc 74 is provided with a disc hub 82 formed with internal drive grooves 84 which receive drive balls 86. the latter being received also in ball grooves 88 formed in shaft 80.
The left margin of cylinder 90 is secured by welding at 92 to the periphery of disc 74. A closure plate 94 is situated within the cylinder 90 and is secured at its inner margin 96 to the shaft 80. the - a - latter being journalled by tapered roller bearings 98 in a bearing opening formed in a support wall 100, which forms a part of the housing 18. Piston plate 102 is situated in the cylinder 90 between the end wall 94 and the piston 104, the latter including a longitudinally extending sleeve 106 that is secured to the shaft 80.
Piston 102 and wall 94 define a working chamber 108 for control pressure, and piston 104 cooperates with the pulley disc 74 to define a working pressure chamber 110. Chamber 108 communicates with chamber 110 through port 112 and through the annular space between piston portion 106 and the shaft 82. Control pressure is supplied to the chambers 108 and 110 through port 114 which communicates with control pressure passage 116 formed in the shaft 52. When the chambers 108 and 110 are pressurized, the pulley discs 70 and 74 move toward each other, which increases the effective pulley pitch diameter. The balls 86 cause the pulley discs to rotate in unison although they permit axial adjustment of disc 74 relative to disc 70.
The right hand end of shaft 80 is splined at 118 to the hub 120 of turbine 16. A torsional damper
assembly 122 connects the hub 120 to clutch disc 124. Damper assembly 122 includes a drive plate 126 which is splined to the hub 120.
Clutch disc 124 carries friction material 128 which frictionally engages the radial face 130 of the impeller housing 38. Fluid pressure can be admitted selectively to the space 132 between the wall 40 of the housing 38 and the piston plate 124 through annular passage 134 defined by sleeve shaft 80 and the pump shaft 82. This passage in turn communicates with supply passages 136 formed in the housing 18.

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The piston disc 124 is connected to damper plate 126 through damper springs 136 which are situated between damper plates 140 joined to the piston plate 124 and the damper disc 126.
As seen in Figure 1A, a flexible drive chain 142 is located in the V-shaped pulley groove defined by pulley discs 70 and 74. Steel belt 142 has side friction surfaces 144 formed on an angle to match the angle of the surfaces of the groove defined by the pulley discs. Belt 142 registers also with the pulley groove defined by pulley discs 146 and 148 shown in Figure IB. These discs, like the discs 70 and 74. define a V-groove that receives the belt 142. It also has cone surfaces that match the angle of the side friction surfaces of the chain.
Disc 148 is formed on. or forms a part of, a pulley shaft 150 which is journalled at its right hand end by tapered roller bearings 152 located in a bearing recess 154 in the housing 28.
The pulley sheave 148 includes a hub 156 which is journalled by a tapered roller bearing 158 located in bearing opening 160 formed in the forward wall portion 62 of the housing. A cover plate 162 covers a bearing opening. Sheave 146 is supported on the hub 156. - It includes internal ball grooves that register with external ball grooves formed in the hub 156. A driving connection is established between the sheave 146 and the hub 156 by balls 164 located in the registering grooves.
Sheave 146 carries at its outer margin a cylin- er 166. Piston 168 is received within the cylinder 166, the latter being movable with the sheave 146 as the piston 168 remains stationary. A series of springs 170 is positioned between the sheave 146 and the stationary piston 168. A pressure chamber is defined by the pres-sure cavity located on the right hand side of the piston 168 as shown at 172. Fluid pressure is admitted to
chamber 172 through port 174 and communicating radial
passage 176 formed in the hub 156.
A radial wall 178 is carried by the cylinder 166 and is fixed thereto by snap ring 180. Wall 178 creates a static fluid pressure cavity between it and the piston 168 so that any centrifugal pressure that is developed due to the rotation of the sheave 146. Piston 168 and cylinder 166 will be opposed by a counterbalancing pres- sure developed due to the rotation of the fluid within the space between wall 178 and piston 168.
Piston 168 is adapted to rotate with the hub 156 since it is secured by staking at 182 to the hub 156, but it is held axially fast as the sheave 146 is adjusted.
In Figure IB a first drive range clutch 184, is located adjacent sheave 148, and a second high ratio
drive range clutch assembly is shown. at 186. Clutch
assembly 184 includes a clutch disc carrier 188 and an
overrunning coupling 190, the latter having an outer race 192 that is externally splined to internally splined
clutch discs of the clutch assembly 184. An annular •
cylinder 194 is formed in the sheave 148, and an annular cylinder 196 is received in the annular cylinder 194.
Piston 196 includes a pressure ring 198 that applies
pressure to the clutch discs, the reaction force being
taken on a thrust ring 200 carried by the disc carrier
Overrunning coupling 190 includes an inner race
202 that is splined to shaft 150. It is connected
directly to low range drive gear 204 supported on shaft
Clutch assembly 186 includes a clutch hub 206 to which is drivably connected a clutch cylinder 208. The outer margin of the cylinder 208 carries externally
splined clutch discs of the clutch assembly 186, and
cooperating internally splined clutch discs of the clutch

*__ WIPO _ assembly 186 are splined to externally splined member 210, which is connected directly to second range drive gear 212 supported on the shaft 150.
Cylinder 208 defines an annular pressure chamber 214 which receives annular piston 216. the latter having a pressure ring 218 that distributes a clutch engaging force to the disc of the clutch assembly 186. The reaction force is accommodated by a thrust ring 220 carried by the cylinder 208.
Pressure is distributed to the annular pressure chamber 214 through radial ports 222 and 224 located in shaft 150 and in hub 206. the latter being splined to shaft 150 at 226.
Reverse drive gear 228 is connected to the gear 212 and together with the gear 212 is supported by the shaft 150. It engages drivably reverse drive pinion 230 which engages in turn reverse gear 232.
Gear 232 is rotatably supported on countershaft 234, one end of which is journalled by a tapered roller bearing 236 mounted in bearing opening 238 of housing portion 240 which forms a part of the previously described housing 18 and housing portion 28. The other end of the countershaft 234 is supported by a tapered roller bearing 242 positioned in a bearing pocket 244 in housing portion 246, the latter forming a part of housing portion 62 previously described.
Second drive range gear 248 is rotatably supported on the countershaft 234. It meshes with gear 212. Low drive range gear 250 is drivably connected to countershaft 234, and it meshes with low range gear 204. Final drive gear 252 drivably engages differential drive gear 254 carried by a differential carrier 256. A spline clutch hub 258 is connected drivably to the countershaft 234, and an internally spline clutch sleeve 260 is slid-ably positioned on the hub 258. Gear 232 has external clutch teeth 262 located directly adjacent corresponding external clutch teeth on the hub 258, and gear 248 has external clutch teeth 264 located directly adjacent the external teeth of the hub 258 on the left hand side thereof. Sleeve 260 can be shifted from a left hand position to a right hand position, and when it is in the left hand position, countershaft 234 becomes con- nected directly to the gear 248. When the clutch sleeve 260 is in the right hand position, countershaft 234 becomes directly connected to gear 232.
Carrier 256 is end supported by a pair of bearings 266 and 268. Carrier 256 contains a pair of side gears 270 and 272 which are connected drivably to output shafts 274 and 276 respectively. Shaft 274 is connected to one end of a half axle shaft not shown through a universal joint 278. Similarly, shaft 276 is connected to one end of a companion axle half shaft through a universal joint 280. Pinions, one of which is shown at 282, are mounted on pinion shaft 284; and they engage side gears 270 and 272. Pinion shaft 284 is fixed to the carrier 258 during operation of the transmission structure. Torque is distributed to the first drive sheave 72 through the fluid coupling 12. Ratio changes are
achieved by controlling the pressure distribution across the member 104. This causes the effective pitch diameter of the belt and pulley arrangement to change. Load on the belt can be controlled by regulating the pressure in pressure chamber 172 of the lower sheave assembly.
During the initial stages of acceleration from a standing start, torque is distributed through the fluid coupling to the input sheave assembly 72; and after a steady state road load condition is achieved, the lockup clutch, shown at 124. is applied thereby eliminating the hydrokinetic losses in the coupling. The lockup clutch is applied by decreasing the pressure in chamber 132 behind the piston 124.

When the low range clutch assembly 184 is
applied, the output sheave assembly delivers torque to low range gear 204 through the overrunning coupling 190.
This drives gear 250, which in turn drives the differen-tial carrier. A drive range change can be achieved by engaging clutch 186 without a synchronous disengagement of the clutch assembly 184. When clutch assembly 186 is engaged, the output shift assembly drives gear 212
through the clutch 186; and overrunning coupling 190
freewheels. Clutch assembly 184 may remain applied
during operation in this fashion in the high speed drive range. Gear 212 drives gear 248, which in turn drives the gear 252 which drives the differential carrier. This requires the spline sleeve clutch 260 to be shifted in a left hand direction to establish a driving connection between clutch hub 258 and the external teeth 264.
During the initial drive-away condition as the vehicle is accelerating, the fluid coupling is operative; and power is transmitted through the one-way clutch to the low range gear pair 204 and 250. At a low vehicle speed the spline clutch sleeve 260 is shifted in a left hand direction, thus coupling the high range gear pair and uncoupling the reverse gearing. A low-to-high range shift can be scheduled by the control system to take
place at the overdriving belt ratios. The range upshift is accomplished by the modulated application of the
clutch 186 after the spline clutch sleeve 260 is in
place. The nonsynchronous shift smoothness is due to the close ratios and the application of the single clutch without the need of timed synchronous engagement or
disengagement of a companion friction element.
A downshift is accomplished by the modulated release of the high clutch. As the vehicle speed
decreases, the downshift would be scheduled at an appro-priate speed; and then the spline coupler would be
switched back to the reverse position in preparation for

O PI reverse drive. Rocking the vehicle forward and rearward can be accomplished by alternately applying and releasing the two clutch assemblies, the low clutch assembly 184 for forward movement and the high clutch assembly 186 for reverse movement.
The dual range feature of this transmission combines two discrete ratio changing features with a continuously variable belt drive function, thus expanding the ratio range and improving fuel economy while at the same time maintaining the necessary performance ratios for maximum acceleration. The coupling bypass clutch, shown in part at 124, also may be modulated for smooth engagement. The dual range feature of the transmission provides the transmission designer with the options to expand the underdrive ratio for increased torque capacity and efficiency, to expand the ratio range, or design a combination of both features for underdrive ratio reduction and expanded ratio range.

Industrial Applicability

This improved power transmission mechanism is intended for automobiles using small or medium displacement internal combustion engine.