TRANSMISSION

CROSS-REFERENCE TO RELATED APPLICATIONS This application claims the benefit of U.S. Provisional Application No.

61/648,684, filed May 18, 2012.

FIELD OF THE INVENTION

The present invention relates to hydraulic systems and methods of utilization thereof for dual clutch transmission.

BACKGROUND OF THE INVENTION

Examples of dual clutch transmissions are described in U.S. Patent Nos. 5,71 1 ,409; 6,996,989; 6,887,184; 6,909,955; U.S. Publication Nos. 2006/0101933A1 ; 2006/0207655A1 ; and 2010/0096232A1 . A hydraulic system for a dual clutch transmission is shown in Koenig et al., U.S. Patent No. 6,898,992 (commonly assigned). It is desirable that a hydraulic system for a dual clutch transmission wherein main line pressure is provided by a valve piloted by the maximum pressure within the two main clutches of the dual clutch transmission.

SUMMARY OF THE INVENTION

To make manifest the above noted and other desires, a revelation of the present invention is brought forth. The present invention in a preferred embodiment provides a hydraulic system and method of utilization thereof for a multiple clutch transmission, the transmission. The present invention provides a hydraulic system for dual clutch transmission where a main line pressure is provided by a valve piloted by a comparison valve which provides a pilot pressure signal representative of the maximum pressure within one of the dual clutches utilized in the transmission.

Further areas of applicability of the present invention will become apparent from the detailed description provided hereinafter. It should be understood that the detailed description and specific examples, while indicating the preferred embodiment of the invention, are intended for purposes of illustration only and are not intended to limit the scope of the invention. BRIEF DESCRIPTION OF THE DRAWINGS

The present invention will become more fully understood from the detailed description and the accompanying drawings, wherein:

Figure 1 is a schematic view of a dual clutch transmission;

Figure 2 is a sectioned view of an actuator for a synchronizer utilized in the dual clutch transmission of Figure 1 ;

Figure 3 is a hydraulic schematic of a hydraulic system prior to a hydraulic system of the present invention;

Figure 4 is a hydraulic schematic of a hydraulic system of the present invention;

Figure 5 is a graphic representation of maximum clutch pressure versus main line regulated pressure; and

Figure 6 is an alternate preferred embodiment hydraulic schematic of a hydraulic system of the present invention. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

The following description of the preferred embodiment(s) is merely exemplary in nature and is in no way intended to limit the invention, its application, or uses.

A representative dual clutch transmission that may be controlled by the present invention is generally indicated at 10 in the schematic illustrated in FIG. 1 . Specifically, as shown in FIG. 1 , the dual clutch transmission 10 includes a dual, coaxial clutch assembly generally indicated at 12, a first input shaft, generally indicated at 14, a second input shaft, generally indicated at 16, that is coaxial to the first, a counter shaft, generally indicated at 18, an output shaft 20, a reverse counter shaft 22, a plurality of synchronizers, generally indicated at 24, and a plurality of shift actuators generally indicated at 26 (FIG. 2). The dual clutch transmission 10 forms a portion of a vehicle powertrain and is responsible for taking a torque input from a prime mover, such as an internal combustion engine, and transmitting the torque through selectable gear ratios to the vehicle drive wheels. The dual clutch transmission 10 operatively routes the applied torque from the engine through the dual, coaxial clutch assembly 12 to either the first input shaft 14 or the second input shaft 16. The input shafts 14 and 16 include a first series of gears, which are in constant mesh with a second series of gears disposed on the counter shaft 18. Each one of the first series of gears interacts with one of the second series of gears to provide the different gear ratios sets used for transferring torque. The counter shaft 18 also includes a first output gear that is in constant mesh with a second output gear disposed on the output shaft 20. The plurality of synchronizers 24 are disposed on the two input shafts 14, 16 and on the counter shaft 18 and are operatively controlled by the plurality of shift actuators 26 to selectively engage one of the alternate gear ratio sets. Thus, torque is transferred from the engine to the dual, coaxial clutch assembly 12, to one of the input shafts 14 or 16, to the counter shaft 18 through one of the gear ratio sets, and to the output shaft 20. The output shaft 20 further provides the output torque to the remainder of the powertrain. Additionally, the reverse counter shaft 22 includes an intermediate gear that is disposed between one of the first series of gears and one of the second series of gears, which allows for a reverse rotation of the counter shaft 18 and the output shaft 20. Each of these components will be discussed in greater detail below.

Specifically, the dual, coaxial clutch assembly 12 includes a first clutch mechanism 32 and a second clutch mechanism 34. The first clutch mechanism 32 is, in part, physically connected to a portion of the engine flywheel (not shown) and is, in part, physically attached to the first input shaft 14, such that the first clutch mechanism 32 can operatively and selectively engage or disengage the first input shaft 14 to and from the flywheel. Similarly, the second clutch mechanism 34 is, in part, physically connected to a portion of the flywheel and is, in part, physically attached to the second input shaft 16, such that the second clutch mechanism 34 can operatively and selectively engage or disengage the second input shaft 16 to and from the flywheel. As can be seen from FIG. 1 , the first and second clutch mechanisms 32, 34 are coaxial and co-centric such that the outer case 28 of the first clutch mechanism 32 fits inside of the outer case 36 of the second clutch mechanism 34. Similarly, the first and second input shafts 14, 16 are also coaxial and co-centric such that the second input shaft 16 is hollow having an inside diameter sufficient to allow the first input shaft 14 to pass through and be partially supported by the second input shaft 16. The first input shaft 14 includes a first input gear 38 and a third input gear 42. The first input shaft 14 is longer in length than the second input shaft 16 so that the first input gear 38 and a third input gear 42 are disposed on the portion of the first input shaft 14 that extends beyond the second input shaft 16. The second input shaft 16 includes a sixth input gear 40, a fourth input gear 44, a second input gear 46, and a reverse input gear 48. As shown in FIG. 1 , the sixth input gear 40 and the reverse input gear 48 are fixedly supported on the second input shaft 16 and the fourth input gear 44 and second input gear 46 are rotatably supported about the second input shaft 16 upon bearing assemblies 50 so that their rotation is unrestrained unless the accompanying synchronizer is engaged, as will be discussed in greater detail below.

In a preferred embodiment, the counter shaft 18 is a single, one-piece shaft that includes the opposing, or counter, gears to those on the inputs shafts 14, 16. As shown in FIG. 1 , the counter shaft 18 includes a first counter gear 52, a sixth counter gear 54, a third counter gear 56, a fourth counter gear 58, a second counter gear 60, and a reverse counter gear 62. The counter shaft 18 fixedly retains the fourth counter gear 58 and second counter gear 60, while first, sixth, third, and reverse counter gears 52, 54, 56, 62 are supported about the counter shaft 18 by bearing assemblies 50 so that their rotation is unrestrained unless the accompanying synchronizer is engaged as will be discussed in greater detail below. The counter shaft 18 also fixedly retains a first drive gear 64 that meshingly engages the corresponding second driven gear 66 on the output shaft 20. The second driven gear 66 is fixedly mounted on the output shaft 20. The output shaft 20 extends outward from the transmission 10 to provide an attachment for the remainder of the powertrain.

In a preferred embodiment, the reverse counter shaft 22 is a relatively short shaft having a single reverse intermediate gear 72 that is disposed between, and meshingly engaged with, the reverse input gear 48 on the second input shaft 16 and the reverse counter gear 62 on the counter shaft 18. Thus, when the reverse gears 48, 62, and 72 are engaged, the reverse intermediate gear 72 on the reverse counter shaft 22 causes the counter shaft 18 to turn in the opposite rotational direction from the forward gears thereby providing a reverse rotation of the output shaft 20. It should be appreciated that all of the shafts of the dual clutch transmission 10 are disposed and rotationally secured within the transmission 10 by some manner of bearing assembly such as roller bearings, for example, shown at 68 in FIG. 1 .

The engagement and disengagement of the various forward and reverse gears is accomplished by the actuation of the synchronizers 24 within the transmission. As shown in FIG. 1 in this example of a dual clutch transmission 10, four synchronizers 74, 76, 78, and 80 are utilized to shift through the six forward gears and reverse. It should be appreciated that there are a variety of known types of synchronizers that are capable of engaging a gear to a shaft and that the particular type employed for the purposes of this discussion is beyond the scope of the present invention. Generally speaking, any type of synchronizer that is movable by a shift fork or like device may be employed. As shown in the representative example of FIG. 1 , the synchronizers are two sided, dual actuated synchronizers, such that they engage one gear to its respective shaft when moved off a center neutralized position to the right and engage another gear to its respective shaft when moved to the left. Specifically with reference to FIG. 1 , synchronizer 78 can be actuated to the left to engage the first counter gear 52 on the counter shaft 18 or actuated to the right to engage the third counter gear 56. Synchronizer 80 can be actuated to the left to engage the reverse counter gear 62 or actuated to the right to engage the sixth counter gear 54. Likewise, synchronizer 74 can be actuated to the left to engage the fourth input gear 44 or actuated to the right to engage the second input gear 46. Synchronizer 76 is actuated to the right to directly engage the end of the first input shaft 14 to the output shaft 20 thereby providing a direct 1 :1 (one to one) drive ratio for fifth gear. There is no gear set to engage to the left of synchronizer 76.

To actuate the synchronizers 74, 76, 78, and 80, this representative example of a dual clutch transmission 10 utilizes hydraulically driven shift actuators 26 with attached shift forks to selectively move the synchronizers so that they engage or disengage (neutralize) the desired gears. As shown in FIG. 2, the shift actuators 26 are essentially two way or dual hydraulic valve assemblies that are driven back and forth linearly, in parallel to one of the input shafts 14, 16 or the counter shaft 18, to move a shift fork 96, and ultimately one of the plurality of synchronizers 24 in and out of engagement. It should be appreciated from the description that follows that other types of actuators that are capable of driving a shift fork back and forth to move a synchronizer may also be employed with the method of the present invention. These include mechanical actuators, hydro-mechanical actuators, electromechanical actuators, electrical actuators, and the like.

Referring to FIG. 2, the hydraulically operated shift actuators 26 include an outer case 86 that includes a main bore 88 having two cylindrically shaped ends 90, 92. A piston 98 is slidably disposed within the main bore 88 of the case 86. The piston 98 includes two opposing sealed heads 82 and 84. The interaction of each piston head 82 and 84 within its respective cylinder end 90, 92 forms alpha and beta pressure or expansion chambers 106, 104.

Between the piston heads 82 and 84 is a gap. Positioned within the gap 95 is the shift fork 96. To actuate the synchronize 74 to the right to actuate the second gear ratio, fluid is injected into alpha expansion chamber 106 through inlet-outlet 100 to move the piston and shift fork 96 to the right causing synchronizer 80 to engage the second input gear 46 to the shaft 16. A detent mechanism (not shown) connected with the linkage with the shift fork 96 holds the shift fork 96 in to hold its actuated position. To release the second input gear 46 from its shaft 16, the beta expansion chamber 104 is pressurized through inlet 102 and the piston 98 and shift fork 96 are shifted back to a detented neutral position. A slight pressurization of the expansion chamber 106 is temporarily maintained to prevent overtravel of the piston 98 and inadvertent engagement of fourth input gear 44 to the shaft 16.

Figure 3 illustrates a hydraulic control system for the first and second clutches 32, 34 and for synchronizers 74, 76, 78 and 80. The control system 7 has an oil sump 120. To provide a source of pressurize oil or fluid, a pump 122 is connected to the sump 120 via a suction filter 124. The pump 122 delivers pressurized fluid to lines 126 and 128. A pump relief valve 127 connected to line 126 prevents over pressurization in line 126. Fluid in line 128 passes through pressure filter 130 into lines 132 and 134. The line 134 is connected with a variable bleed solenoid (VBS.) 136. The VBS 136 controls operation of a main line pressure regulator valve 140. When at least partially actuated by VBS 136, the valve 140 connects line 126 with a lube regulator valve 142. A VBS 144 controls valve 142 to control lubrication of the clutches 32, 34 via a clutch lube line 146. Clutch lube line 146 is connected with the clutch lubrication system. Valve 140 is also fluidly connected with an oil cooler limit valve 148. An outlet of the valve 148 is looped back to in inlet side of the pump 122. Valve 142 additionally delivers fluid to oil cooler 150. Variable force solenoids (VFS) 152, 154 control the pressure within their respective clutches 32 and 34 by selectively communicating to the clutches with the line 134 or with the sump 120.

The control system for the synchronizers includes a first multiplex valve 160. The first multiplex valve 160 has a first position allowing delivery of pressurize fluid to synchronizers 74 and 76. Synchronizers 78 and 80 are diverted to the sump 120. In a second position of the first multiplex valve 160 the reverse occurs allowing delivery of pressurize fluid to synchronizers 78 and 80 with synchronizers 74 and 76 being diverted to the sump. An on/off solenoid valve 162 controls operation of the first multiplex valve 160.

A second multiplex valve 164 is fluidly connected with the first multiplex valve 160. The second multiplex valve 164 has a first position allowing pressurized fluid connection of alpha and beta chambers of the synchronizer 74 (when the first multiplex valve 160 is in the first position). The alpha and beta chambers of synchronizer 76 are diverted to the sump 120. When the second multiplex valve 164 is placed in the second position by an on/off solenoid 166, the fluid connections of the second multiplex valve 164 are reversed. The alpha chambers for the synchronizers 74, 76, 78, and 80 include pressure chambers for odd and even gear ratios.

To actuate the alpha chamber there is provided a first actuator regulator valve 170. First actuator regulator valve 170 has a biased position connecting the alpha chamber to the sump 120. In a second position, the first actuator regulator valve 170 connects the alpha chamber with the line 132. A proportional solenoid valve provided by VBS solenoid valve 174 controls the first actuator regulator valve 170. In like manner, VBS 176 controls the second actuator regulator valve 180 for the beta chamber of the synchronizer 74.

To control the synchronizer 76 the first multiplex valve 160 is in the first position and the second multiplex valve is placed in the second position. To control synchronizer 80 or synchronizer 78 the first multiplex valve 160 is placed in the second position. For synchronizer 80, the second multiplex valve 164 is in the first position. For control of the synchronizer 78, the second multiplex valve 164 is placed in the second position.

To place the second input gear 46 into engagement with the shaft 16, the first multiplex valve 160 and second multiplex valve 162 are placed in the first position. The first regulator valve 170 is turned on to pressurize the alpha expansion chamber 106 moving the piston 98 and shift fork 96 to the right. A position sensor 97 is used to inform or confirm the fact to the transmission electronic controller (not shown) that the transmission 10 is in the second gear. A feature of the control system for the synchronizers is that no two gears of the transmission can be actuated at the same time. If the second input gear 46 is being actuated, all of the pressure chambers of the synchronizers 80, 78 and 76 are diverted to the sump. If a control system failure causes the second actuator regulator valve 180 to pressurize the beta expansion chamber 104 of the shift actuator 26 for synchronizer 74, the pressure within the opposing beta 104 and alpha 106 expansion chambers act against each preventing any simultaneous gear activation (however when the alpha chamber 106 is depressurized the above noted failure causes the gear (fourth input gear 44) associated with the beta chamber to be stuck on). Another feature of the control system is that most valve failures allow at least one odd gear and at least one even gear to still operate. Failure of the first multiplex valve 160 in the first position allows operation of synchronizer 74 providing second and fourth gears. Additionally, fifth and neutral gears of synchronizer 76 are available. Upon such a failure, the transmission controller programs the transmission 10 to operate in second, fourth and fifth gear ratios dependent upon vehicle speed in a "limp" home mode of operation.

A failure of the first multiplex valve 160 in the second position still allows for operation of the reverse, six gear, third gear and first. Failure of the second multiplex valve 164 in the first position will still allow operation of the second, fourth, sixth and reverse gears. Failure of any one given actuator valve still allows for partial gear operation. Failure of the actuator regulator valve 170 in the on position will freeze (be detented) second input gear 46 with the shaft 16. To get another gear for "limp" home operation, the transmission controller opens the clutch 34. Engaged clutch 32 is utilized to rotate the shaft 14. The controller of the transmission then picks a gear ratio from a set of gear ratios associated with the shafts 14 or 20 (first, third or fifth) gear to be utilized for "limp" home mode of operation. The transmission controller then alternates between second and one gear from the set of first, third or fifth gear. As long as a forward travel gear is engaged when one of the actuator regulator valves 170, 180 fails, the transmission will have two gear ratios of forward operation in the "limp" home mode of operation.

An advantage of the hydraulic system shown in Figure 3 over other hydraulic systems as shown in Koenig et al. US Patent 6,898,992 (commonly assigned) is that only two high flow rate solenoid actuator regulator valves 170, 180 are required.

Referring to Figure 3, a variable bleed solenoid valve 136 is utilized to provide a pilot pressure for the main line regulator valve 140. It is desirable to provide a hydraulic control system for a dual clutch transmission wherein a main line supplying pressure to hydraulic system does not require a variable bleed solenoid. Variable bleed solenoids can typically comprise 2-4% of the total cost of the hydraulic control system therefore providing a less expensive alternative than the variable solenoid 136 can provide significant cost savings when considering that the control system is utilized in a highly mass produced item. Referring additionally to Figures 4 and 5, a preferred embodiment of the present invention is brought forth. In Figure 4, items providing similar functions in that to the prior art control system shown in Figure 3 are given like reference numbers. In the inventive hydraulic control system and method according to the present invention, line 193 connects solenoid 152 with odd gear ratio clutch 32. Line 192 fluidly connects solenoid valve 154 to even gear ratio clutch 34. Line 192 and 193 are also fluidly connected with a dual check valve type comparison valve 206. Comparison valve 206 is configured to connect alternately line 193 or 192 with pilot pressure line 212 dependent upon which clutch 32 or 34 has the greater pressure and is essentially two check valves connected with one another. If the transmission is being operated wherein the pressure in clutch 34 is greater, comparison valve 206 will connect line 192 with pilot line 212. In similar fashion, if the pressure within clutch 32 is greater, comparison valve 206 will connect line 193 with pilot line 212. The cost of comparison valve 206 is significantly less than the prior variable bleed solenoid valve 136 (Figure 3). Referring to Figure 5, utilizing the pressure within the clutches to provide a pilot pressure for the main line regulator valve causes the main line pressure to have generally fixed value typically one or two bars above that of the highest pressure within clutch 32 or 34.

Figure 6 is a schematic of an alternate preferred embodiment hydraulic system of the present invention. The hydraulic system of Figure 6 is similar to that of Figure 4 with a major exception in that the synchronizers (not shown) are electromechanically controlled. The hydraulic system of Figure 6 also shares the advantage of the control system of Figure 4 in that by utilizing the comparison valve to provide the clutch pilot pressure allows the engine driven pump to lower energy consumption in the transmission since the pressure that the pump is working against is held at a level slightly offset above from the pressure within the clutches.

The description of the invention is merely exemplary in nature and, thus, variations that do not depart from the gist of the invention are intended to be within the scope of the invention. Such variations are not to be regarded as a departure from the spirit and scope of the invention.