Hydraulic control systems for dual clutch transmissions

The present disclosure relates to hydraulic control systems and more particularly to hydraulic control systems and their components for dual clutch transmissions.

The statements in this section merely provide background information related to the present disclosure and may or may not constitute prior art.

In automotive transmission art, the dual clutch transmission (DCT) is a relatively new concept. A typical dual clutch transmission configuration includes a pair of mutually exclusively operating input clutches which drive a pair of input shafts. The input shafts may be disposed on opposite sides of an output shaft or may be disposed concentrically between spaced-apart output shafts. One of each of a plurality of pairs of constantly meshing gears which provide the various forward and reverse gear ratios is freely rotatably disposed on one of the shafts and the other of each pair of gears is coupled to one of the other shafts. A plurality of synchronizer clutches selectively couple the freely rotatable gears to the associated shaft to achieve forward and reverse gear ratios. After the synchronizer clutch is engaged, the input clutch associated with the input shaft having the engaged synchronizer clutch is applied to transmit power through the transmission. Reverse gear is similarly achieved except that it includes an additional (idler) gear to provide torque reversal.

Dual clutch transmissions are known for their sporty, performance oriented operating characteristics which mimic those of a conventional (manual) transmission. They also typically exhibit good fuel economy due to their good gear mesh efficiency, ratio selection flexibility, reduced clutch losses and the lack of a torque converter.

There are several design considerations unique to dual clutch transmissions. For example, because of heat generated during clutch slip, the input clutches must be of relatively large size. Furthermore, such heat generation typically requires correspondingly larger and more complex cooling components capable of dissipating relatively large quantities of heat. Finally, because such transmissions typically have many sets of axially aligned, meshing gears, their overall length may limit their use to certain vehicle designs.

Control of the input clutches and selection and engagement of a particular gear by translation of a synchronizer and associated positive clutch is typically achieved by a hydraulic control system. Such a system, itself under the control of an electronic transmission control module (TCM), includes hydraulic valves and actuators which engage the synchronizers and gear clutches. Optimum operating efficiency and thus fuel efficiency and minimal heat generation can be achieved by designing such hydraulic control systems to exhibit low leakage and positive control characteristics. The present invention is so directed.

The present invention comprehends two embodiments of a hydraulic control system for a dual clutch transmission having three countershafts, an idler shaft and five shift rails and hydraulic actuators. The hydraulic control systems include a regulated source of pressurized hydraulic fluid including a pump, a filter and an accumulator, a pair of pressure control valves and a branching hydraulic circuit including pressure or flow control valves, spool or logic valves and two position valves which collectively supply and exhaust hydraulic fluid from a plurality of shift actuators. The actuators are connected to shift rails which include shift forks and are slidable to engage synchronizers and positive clutches associated with the various gear ratios.

The embodiments define essentially independent control systems supplied with hydraulic fluid through two independently operating valves. The independent control systems are associated with respective transmission countershafts and, generally speaking, one countershaft is associated with the even-numbered gears (second, fourth, etc.) and the other countershaft is associated with the odd-numbered gears (first, third, etc.). When the transmission is operating in a normal ascending or descending gear selection sequence, this configuration generally permits pre-staging or pre-selection of a gear associated with one countershaft while a gear associated with the other countershaft is engaged and transmitting torque. Furthermore, if a component or components associated with one countershaft fail, the other countershaft and the alternating (i.e., first, third, fifth) selection of gear ratios it provides will still be fully operational—a highly desirable failure mode.

The hydraulic control systems according to the present invention are less complex and expensive relative to competing systems, provide improved control through interconnected logic valves which reduce the likelihood of engaging a wrong or multiple gears and provide reduced energy consumption by allowing shut-down of portions of the control system during steady state operation. The control systems utilize pairs of pressure, flow control valves, on/offs or a combination of components to control both sides of shift actuator pistons which provides better control and improved shifts.

Thus it is an object of the present invention to provide a hydraulic control system for a dual clutch automatic transmission.

It is a further object of the present invention to provide a hydraulic control system for a dual clutch transmission having a plurality of spool or logic valves and hydraulic actuators.

It is a still further object of the present invention to provide a hydraulic control system for a dual clutch transmission having a plurality of two position solenoid valves (on-offs), spool valves and hydraulic actuators.

It is a still further object of the present invention to provide a hydraulic control system for a dual clutch transmission having a plurality of flow or pressure control valves, two position solenoid valves, logic or spool valves and hydraulic actuators.

It is a still further object of the present invention to provide a hydraulic control system for a dual clutch transmission comprising two essentially independent hydraulic systems, each associated with a respective transmission countershaft.

It is a still further object of the present invention to provide a hydraulic control system for a dual clutch transmission having a pair of input clutches associated with a pair of concentric input shafts and a plurality of countershafts.

Further objects, advantages and areas of applicability will become apparent from the description provided herein. It should be understood that the description and specific examples are intended for purposes of illustration only and are not intended to limit the scope of the present disclosure.

The drawings described herein are for illustration purposes only and are not intended to limit the scope of the present disclosure in any way.

FIG. 1 is a pictorial view of an exemplary dual clutch automatic transmission with portions broken away incorporating a hydraulic control system according to the present invention having five shift actuator assemblies;

FIGS. 2A , 2 B and 2 C are schematic flow diagrams of a first embodiment of a hydraulic control system according to the present invention for a dual clutch automatic transmission; and

FIGS. 3A , 3 B and 3 C are schematic flow diagrams of a second embodiment of a hydraulic control system according to the present invention for a dual clutch automatic transmission; and

FIG. 4 is an enlarged view of a portion of the second embodiment of a hydraulic control system illustrating a passive pressure regulator according to the present invention.

The following description is merely exemplary in nature and is not intended to limit the present disclosure, application, or uses.

With reference to FIG. 1 , an exemplary dual clutch automatic transmission incorporating the present invention is illustrated and generally designated by the reference number 60 . The dual clutch transmission 60 includes a typically cast, metal housing 12 which encloses and protects the various components of the transmission 60 . The housing 12 includes a variety of apertures, passageways, shoulders and flanges (not illustrated) which position and support the components of the transmission 60 . The transmission 60 includes an input shaft 14 which receives motive power from a prime mover (not illustrated) such as an internal combustion gas or Diesel engine or a hybrid or electric power plant and a single or double output shaft 16 which drives a final drive assembly 18 which may include a propshaft, a differential and drive axles. The input shaft 14 is coupled to and drives a clutch housing 62 . The clutch housing 62 , in turn, drives a pair of concentrically disposed dry input clutches, a first input clutch 64 A and a second input clutch 64 B which are mutually exclusively engaged to provide drive torque to a respective pair of concentric input members, a first or inner input shaft 66 A and a second or outer hollow input shaft or quill 66 B.

Secured to and rotating with each of the input members 66 A and 66 B are a plurality of helical or spur gears (not illustrated) which are in constant mesh with helical or spur gears which are freely rotatably disposed on a first layshaft or countershaft 68 A and a parallel, second layshaft or countershaft 68 B. Adjacent and parallel to the second countershaft is a third layshaft or countershaft 68 C. A first drive gear meshes with a first driven gear 70 A on the first countershaft 68 A. A second drive gear meshes with a second driven gear 72 A on the first countershaft 68 A. A third drive gear meshes with a third driven gear 74 A on the first countershaft 68 A. A fourth drive gear meshes with a fourth driven gear 76 A on the first countershaft 68 A. A fifth driven gear 70 B on the second countershaft 68 B meshes with a fifth drive gear 70 C on the third countershaft 68 C. The second drive gear also meshes with a sixth driven gear 72 B on the second countershaft 68 B which meshes with a seventh driven gear 72 C on the third countershaft 68 C. An eighth drive gear meshes with an eighth driven gear 74 B on the second countershaft 68 B.

Disposed either adjacent certain single gears or between adjacent pairs of gears on the countershafts 68 A, 68 B and 68 C are synchronizer clutch assemblies. Each synchronizer clutch assembly, in accordance with conventional practice, includes a synchronizer assembly which, when activated, synchronizes the speed of a gear to that of the associated countershaft and a positive clutch, such as a dog or face clutch, which positively connects the gear to the shaft. Thus, between the driven gears 70 A and 72 A on the first countershaft 68 A is a first shift actuator and synchronizer clutch assembly 80 A having a double, i.e., back-to-back, first synchronizer clutch 82 A which selectively and exclusively synchronizes and engages one of the gears 70 A and 72 A to the first countershaft 68 A. The first synchronizer clutch 82 A is bi-directionally translated by a first shift rail and fork assembly 84 A which, in turn, is translated by a first shift actuator assembly 86 A. The real time position of the first synchronizer clutch 82 A and the first shift rail and fork assembly 84 A is sensed by a first linear position sensor 88 A which preferably provides a continuous, i.e., proportional, output signal to a transmission control module TCM indicating the position of the first synchronizer clutch 82 A.

Between the fifth driven gear 70 B and the sixth driven gear 72 B on the second countershaft 68 B is a second shift actuator and synchronizer clutch assembly 80 B having a single synchronizer clutch 82 B which synchronizes and couples the fifth and the sixth gears 70 B and 72 B together. The second synchronizer clutch 82 B is bi-directionally translated by a second shift rail and fork assembly 84 B which, in turn, is translated by a second shift actuator assembly 86 B. The real time position of the second synchronizer clutch 82 B and the second shift rail and fork assembly 84 B is sensed by a second linear position sensor 88 B which preferably provides a continuous, i.e., proportional, output signal to the transmission control module TCM indicating the position of the second synchronizer clutch 82 B.

Between the driven gears 74 A and 76 A on the first countershaft 68 A is a third shift actuator and synchronizer clutch assembly 90 A having a double, i.e., back-to-back, third synchronizer clutch 92 A which selectively and exclusively synchronizes and engages one of the gears 74 A and 76 A to the first countershaft 68 A. The third synchronizer clutch 92 A is bi-directionally translated by a third shift rail and fork assembly 94 A which, in turn, is translated by a third shift actuator assembly 96 A. The real time position of the third synchronizer clutch 92 A and the third shift rail and fork assembly 94 A is sensed by a third linear position sensor 98 A which preferably provides a continuous, i.e., proportional, output signal to the transmission control module TCM indicating the position of the third synchronizer clutch 92 A.

Adjacent the eighth driven gear 74 B on the second countershaft 68 B is a fourth shift actuator and synchronizer clutch assembly 90 B having a single synchronizer clutch 92 B which synchronizes and couples the eighth driven gear 74 B to the second countershaft 68 B. The fourth synchronizer clutch 92 B is bi-directionally translated by a fourth shift rail and fork assembly 94 B which, in turn, is translated by a fourth shift actuator assembly 96 B. The real time position of the fourth synchronizer clutch 92 B and the fourth shift rail and fork assembly 94 B is sensed by a fourth linear position sensor 98 B which preferably provides a continuous, i.e., proportional, output signal to the transmission control module TCM indicating the position of the fourth synchronizer clutch 92 B.

Finally, between the fifth drive gear 70 C and the seventh driven gear 72 C on the third countershaft 68 C is a fifth shift actuator and synchronizer clutch assembly 90 C having a double, i.e., back-to-back, synchronizer clutch 92 C which selectively and exclusively synchronizes and engages the driven gear 72 C to the third countershaft 68 C or couples the driven gear 72 C to the drive gear 70 C. The fifth synchronizer clutch 92 C is bi-directionally translated by a fifth shift rail and fork assembly 94 C which, in turn, is translated by a fifth shift actuator assembly 96 C. The real time position of the fifth synchronizer clutch 92 C and the fifth shift rail and fork assembly 94 C is sensed by a fifth linear position sensor 98 C which preferably provides a continuous, i.e., proportional, output signal to the transmission control module TCM indicating the position of the fifth synchronizer clutch 92 C. It should be understood that the linear position sensors 88 A, 88 B, 98 A, 98 B and 98 C may be replaced with other sensors such as two or three position switches or open loop control with system characterization.

Additionally, a detent mechanism may be employed with each of the shift assemblies to assist obtaining and maintaining a gear or speed ratio once it is selected and to assist obtaining and maintaining the synchronizer clutch in neutral, i.e., an unengaged position. Thus, a first detent assembly 89 A may be operatively associated with the first shift actuator and synchronizer clutch assembly 80 A. A second detent assembly 89 B may be operatively associated with the second shift actuator and synchronizer clutch assembly 80 B. A third detent assembly 99 A may be operatively associated with the third shift actuator and synchronizer clutch assembly 90 A. A fourth detent assembly 99 B may be operatively associated with the fourth shift actuator and synchronizer clutch assembly 90 B and a fifth detent assembly 99 C may be operatively associated with the fifth shift actuator and synchronizer clutch assembly 90 C.

It will be appreciated that the transmission 60 illustrated and described above is laid out with four forward gears on one countershaft and the remaining (three) forward gears and reverse on two other countershafts. It is thus capable of providing seven forward speeds and reverse. Similar configurations, all deemed to be within the scope of this invention may, for example, include six forward speeds (or gears) and one or two reverse speeds (or gears) or five forward speeds and one or two reverse speeds.

It should be understood that while the present invention is directed to hydraulic control systems for dual clutch transmissions, such systems are typically controlled by sensor signals and memory, software and one or more microprocessors contained in a transmission control module TCM. Thus, the transmission control module TCM includes a plurality of inputs which receive data from, for example, the linear position sensors, the speed sensors, and the pressure sensor; and a plurality of outputs which control and modulate, for example, the positions of the clutches, shift rails and logic solenoid valves.

Referring now to FIGS. 1 , 2 A, 2 B and 2 C, a first embodiment of a hydraulic control system for the dual clutch automatic transmission 60 described above is illustrated and designated by the reference number 1700 . The hydraulic control system 1700 includes a sump 102 to which hydraulic fluid returns and collects from various components and regions of the automatic transmission 60 . A suction line 104 which may include a filter 106 communicates with the inlet port 108 of an engine driven or electric pump 110 which may be, for example, a gear pump, a vane pump, a gerotor pump or other positive displacement pump. An outlet port 112 of the pump 110 provides hydraulic fluid under pressure in a supply line 114 to a spring biased blow-off safety valve 116 and to a pressure side filter 118 which is disposed in parallel with a spring biased check valve 120 . The safety valve 116 is set at a relatively high predetermined pressure and if the pressure in the supply line 114 exceeds this pressure, the safety valve 116 opens momentarily to relieve and reduce it. If pressure ahead of the filter 118 rises to a predetermined differential pressure, indicating a partial blockage or flow restriction when cold of the filter 118 and the possibility that insufficient hydraulic fluid may be provided in an outlet line 122 to the remainder of the control system 1700 , the check valve 120 opens to allow hydraulic fluid to bypass the filter 118 .

A second check valve 124 , in the outlet line 122 , is configured to maintain hydraulic pressure in a main supply line 126 and to prevent backflow through the pump 110 . The main supply line 126 supplies pressurized hydraulic fluid to an accumulator 130 having a piston 132 and a biasing compression spring 134 . The accumulator 130 may be one of many other designs including a gas filled piston accumulator. The accumulator 130 stores pressurized hydraulic fluid and supplies it to the main supply line 126 , to a main or system pressure sensor 136 and to the other components of the control system 1700 thereby eliminating the need for either an engine driven pump or an electric pump 110 to run continuously. The main pressure sensor 136 reads the delivered hydraulic system pressure in real time and provides this data to the transmission control module TCM.

It should be appreciated that the other embodiment of the hydraulic control system according to the present invention preferably includes the same hydraulic supply, filtration and control components just described. Accordingly, these components will be only briefly described in connection with the subsequent figure and embodiment, it being understood that the above description may be referenced to provide details of these components.

The main supply line 126 communicates with an inlet port 140 A of a pressure control solenoid valve 140 . An outlet port 140 B of the pressure control solenoid valve 140 communicates with a branching manifold 1702 and an exhaust port 140 C communicates with the sump 102 .

A first branch 1702 A of the branching manifold 1702 communicates with an inlet port 154 A of a first electric pressure or flow clutch control solenoid valve 154 . The first clutch control solenoid valve 154 also includes an outlet port 154 B and an exhaust port 154 C which communicates with the sump 102 . When the clutch control solenoid valve 154 is activated or energized, pressurized hydraulic fluid is provided through a flow control orifice 156 in a line 158 to a first clutch piston and cylinder assembly 160 . It should be understood that the addition or removal of flow control orifices in all the hydraulic lines of the hydraulic control system 1700 and the other embodiment is within the purview of the present invention. The locations and sizes of the flow control orifices are based on operational, software and algorithm requirements. Thus, it should be understood that the incorporation or omission of flow control orifices in all the hydraulic lines is within the scope of the present invention.

Slidably disposed within a cylinder 162 is a single acting piston 164 which translates to the right in FIG. 2A under hydraulic pressure to engage the first input clutch 64 A, illustrated in FIG. 1 . When the first clutch control solenoid valve 154 is de-energized, the inlet port 154 A is closed and hydraulic fluid from the cylinder 162 passes from the outlet port 154 B to the exhaust port 154 C and into the sump 102 . A second branch 1702 B of the branching manifold 1702 communicates with the output of a first clutch pressure limit control valve 166 . If pressure within the first clutch piston and cylinder assembly 160 exceeds a predetermined pressure determined by the pressure control solenoid 140 , the first pressure limit control valve 166 opens to relieve and reduce the pressure. It should be noted that the first pressure limit control valve 166 can be eliminated depending upon the system requirements.

A sixth branch 1702 F of the branching manifold 1702 communicates with an inlet port 204 A of a second electric pressure or flow clutch control solenoid valve 204 . The second clutch control solenoid valve 204 also includes an outlet port 204 B and an exhaust port 204 C which communicates with the sump 102 .

When the second clutch control solenoid valve 204 is activated or energized, pressurized hydraulic fluid is provided through an orifice 206 in a line 208 to a second clutch piston and cylinder assembly 210 . Slidably disposed within a cylinder 212 is a single acting piston 214 which translates under hydraulic pressure to the right in FIG. 2A , to engage the second input clutch 64 B, illustrated in FIG. 1 . When the second clutch control solenoid valve 204 is inactive or de-energized, the inlet port 204 A is closed off and hydraulic fluid from the cylinder 212 passes from the outlet port 204 B to the exhaust port 204 C and into the sump 102 . A seventh branch 1702 G of the branching manifold 1702 communicates with the output of a clutch pressure limit control valve 216 . If pressure within the second clutch piston and cylinder assembly 210 exceeds a predetermined pressure supplied by the pressure control solenoid 140 , then the pressure limit control valve 216 opens to relieve and reduce the pressure. It should be noted that the second pressure limit control valve 216 can be eliminated depending upon the system requirements.

A fourth branch 1702 D of the branching manifold 1702 communicates with the inlet port 1030 A of a first electric pressure or flow control solenoid valve 1030 . An outlet port 1030 B of the first pressure or flow control solenoid valve 1030 is connected by a first hydraulic line 1432 with a first inlet port 1040 A of the first spool or logic control valve 1040 . An exhaust port 1030 C of the first pressure or flow control solenoid valve 1030 communicates with the sump 102 . The fourth branch 1702 D of the branching manifold 1702 also communicates with an inlet port 1430 A of a second electric pressure or flow control solenoid valve 1430 . A second hydraulic line 1434 communicates between an outlet port 1430 B of the second electric pressure or flow control solenoid valve 1430 and a second inlet port 1040 B of the first spool or logic valve 1040 . An exhaust port 1430 C communicates with the sump 102 . It should be noted that instead of feeding the main manifold 1702 C and 1702 D from the control port 140 B, they could be connected to the higher pressure of the main supply line 126 for a slightly different embodiment. This will remove some flow demand from the pressure control solenoid valve 140 but also changes the failure modes.

The first spool or logic valve 1040 includes a control port 1040 C which is selectively supplied pressurized hydraulic fluid from an outlet port 1042 B of a first two position solenoid (on-off) valve 1042 which, in turn, is supplied with pressurized hydraulic fluid at an inlet port 1042 A from a third branch 1702 C of the branching manifold 1702 . When the first two position solenoid valve 1042 is energized, the spool of the first logic valve 1040 translates to the left in FIG. 2B ; when the first two position solenoid valve 1042 is de-energized, the spool of the first logic valve 1040 translates to the right and returns hydraulic fluid to the sump 102 through an exhaust port 1042 C. It should be understood that the method by which the logic valves change states are not limited to hydraulic on/offs. For example, other concepts replace the on/off hydraulic actuation of the logic valve with a solenoid armature and coil which acts directly on the logic valve spool. Additionally some of the on/off valves described below which actuate the logic valves can be multiplexed where a single on/off valve moves multiple logic valves. The first spool or logic valve 1040 also includes three exhaust ports 1040 D, 1040 E, and 1040 F disposed between and alternating with the inlet ports 1040 A and 1040 B which communicate with the sump although, for reasons of clarity, such connections are not shown. It should be understood that slight modifications can be done to the logic valves and ports without changing the function of the system or deviating from the scope of this patent.

A hydraulic line 1046 communicates between a first outlet port 1040 G of the first spool or logic valve 1040 and a first input port 1060 A of a second spool or logic valve 1060 . Another hydraulic line 1048 communicates between a third outlet port 1040 H of the first spool or logic valve 1040 and a second input port 1060 B of the second spool or logic valve 1060 . Another hydraulic line 1052 connects the fourth outlet port 1040 J of the first spool or logic valve 1040 to the first inlet port 1090 A of a third spool or logic valve 1090 . Another hydraulic line 1054 connects the second outlet port 1040 I of the first spool or logic valve 1040 to the second inlet port 1090 B of the third spool or logic valve 1090 .

The second spool or logic valve 1060 includes a control port 1060 C and three exhaust ports 1060 D, 1060 E, and 1060 F which communicate with the sump 102 , although such connections are not shown for reasons of clarity. A second two position (on-off) solenoid valve 1062 includes an inlet port 1062 A which receives hydraulic fluid through a third branch 1702 C of the branching manifold 1702 and selectively supplies it through an outlet port 1062 B to the control port 1060 C of the second logic valve 1060 . When the second two position solenoid valve 1062 is energized, the spool of the second logic valve 1060 translates to the left in FIG. 2B ; when the second two position solenoid valve 1062 is de-energized, the spool of the second logic valve 1060 translates to the right, returning hydraulic fluid through an exhaust port 1062 C to the sump 102 .

Two hydraulic lines 1064 and 1073 communicate between the first and the third outlet ports 1060 G and 1060 H of the second logic valve 1060 and two ports 1068 A and 1068 B in opposite ends of a first, preferably dual area piston and cylinder assembly 1070 which includes a cylinder 1068 and a piston 1072 . The first piston and cylinder assembly 1070 translates the first shift rail and fork assembly 84 A to select one of two gears. Two additional hydraulic lines 1074 and 1083 communicate between the second and fourth outlet ports 1060 I and 1060 J of the second logic valve 1060 and two ports 1078 A and 1078 B in opposite ends of a second piston and cylinder assembly 1080 which includes a cylinder 1078 and a piston 1082 . The second piston and cylinder assembly 1080 translates the second shift rail and fork assembly 84 B to select a single gear.

The third spool or logic valve 1090 includes a control port 1090 C and three exhaust ports 1090 D, 1090 E, and 1090 F which communicate with the sump 102 . A third two-position (on-off) solenoid valve 1092 includes an inlet port 1092 A which receives hydraulic fluid through a fifth branch 1072 E of the branching manifold 1072 . When the third two position solenoid valve 1092 is energized, the spool of the third logic valve 1090 translates to the left in FIG. 2C ; when the third two position solenoid valve 1092 is de-energized, the spool of the third logic valve 1090 translates to the right, returning hydraulic fluid to the sump 102 through an exhaust port 1092 C. A pair of hydraulic lines 1094 and 1103 communicate between the first and third outlet ports 1090 G and 1090 H of the third logic valve 1090 and two ports 1098 A and 1098 B at opposite ends of a third, preferably dual area piston and cylinder assembly 1100 which includes a cylinder 1098 and a piston 1102 . The third piston and cylinder assembly 1100 translates the third shift rail and fork assembly 94 A to select one of two gears.

A hydraulic line 1104 connected to the second outlet port 1090 I communicates with a first inlet port 1530 A of a fourth spool or logic valve 1530 and a hydraulic line 1113 connected to the fourth outlet port 1090 J communicates with a second inlet port 1530 B of the fourth logic valve 1530 . A control port 1530 C at the right end of the fourth logic valve 1530 is selectively supplied with pressurized hydraulic fluid from the outlet port 1062 B of the second two position (on-off) solenoid valve 1062 (illustrated in FIG. 2B ) in a line 1532 . Thus, the spool of the fourth logic valve 1530 translates in unison with the spool of the second logic valve 1060 . Alternately, the forth logic valve 1530 could be controlled by an additional on/off valve without changing the function of the control system 1700 . However, to reduce components a single on/off valve was used to pilot two logic valves. When the second two position (on-off) solenoid valve 1062 is energized, both spools translate to the left as viewed in FIGS. 2B and 2C . When the second two position solenoid valve 1062 is de-energized, both spools translate to the right.

The fourth logic valve 1530 includes a first outlet port 1530 G which communicates through a line 1536 to a port 1538 A at one end of a fourth piston and cylinder assembly 1540 having a cylinder 1538 and a piston 1542 which is coupled to the fourth shift rail and fork assembly 94 B and selects a single gear. A port 1538 B at the other end of the fourth piston and cylinder assembly 1542 communicates through a line 1544 to a third outlet port 1530 H. Similarly, a second outlet port 1530 I communicates through a line 1546 to a port 1548 A at one end of a fifth, preferably dual area piston and cylinder assembly 1550 having a cylinder 1548 and a piston 1552 which is coupled to the fifth shift rail and fork assembly 94 C which selects one of two gears. The other end of the fifth piston and cylinder assembly 1550 includes a port 1538 B which communicates through a line 1554 to a fourth outlet port 1530 J.

Operation of the first embodiment of the hydraulic control system 1700 essentially involves the selection of a desired gear ratio in the transmission 60 by the transmission control module TCM and selection and activation of the pressure or flow control solenoid valves 1030 and 1430 to provide controlled flow and pressure of hydraulic fluid to the logic valves 1040 , 1060 , 1090 and 1530 and activation of the two position solenoid valves 1042 , 1062 and 1092 to position the logic valve spools to direct pressurized hydraulic fluid flow to the correct sides of the piston and cylinder assemblies 1070 , 1080 , 1100 , 1540 and 1550 to translate the shift rails 84 A, 84 B, 94 A, 94 B and 94 C to engage the desired gear. Once this has occurred, the input clutch 64 A or 64 B associated with the countershaft 68 A, 68 B or 68 C of the selected gear is engaged by activation of one of the piston and cylinder assemblies 160 or 210 .

A convenient example of operation may be presented by describing same with the spools of the logic valves 1040 , 1060 , 1090 and 1530 in the positions illustrated in FIGS. 2B and 2C . Activation of the first pressure or flow control solenoid valve 1030 provides hydraulic fluid to the line 1432 , through the first logic valve 1040 , through the line 1046 to the second logic valve 1060 and through the line 1064 to the port 1068 A in the first piston and cylinder assembly 1070 . The piston 1072 and the first shift rail 84 A will then translate to the right (to the left in FIG. 1 ) and engage, for example, sixth gear. The shift is completed by engaging the appropriate input clutch. If, on the other hand, the second pressure or flow control solenoid valve 1430 is activated, hydraulic fluid flow occurs through the lines 1434 , 1048 and 1073 , either returning the first shift rail 84 A to neutral or moving the shift rail 84 A all the way to the left to the position illustrated in FIG. 2B to engage, for example, second gear. The choice of the center (neutral) or left position is commanded by the transmission control module TCM with linear position information from, for example, the first linear position sensor 88 A illustrated in FIG. 1 . It should be appreciated that all the pistons 1072 , 1082 , 1102 , 1542 and 1552 may, if desired, be dual area pistons or may be single area pistons with associated feedback and control assemblies or combinations thereof, as illustrated. A similar pattern of valve activation and logic valve spool translation provides the seven forward and reverse gears of the transmission 60 .

Referring now to FIGS. 1 , 3 A, 3 B and 3 C, a second embodiment of a hydraulic control system according to the present invention is illustrated and generally designated by the reference number 1800 . The second embodiment 1800 of the hydraulic control system, as stated, includes, in common with the other embodiment, the pump 110 , the filters 106 and 118 , the accumulator 130 and the other components of the hydraulic fluid supply and thus they will not be further described. It should be noted that the filters 106 and/or 118 can be omitted without departing from the scope of this invention.

The second embodiment 1800 of the hydraulic control system is the same as the first embodiment 1700 illustrated in FIGS. 2A , 2 B and 2 C except that a passive feed limit regulator assembly 1310 replaces the pressure control solenoid valve 140 . A simplified version can also be envisioned by removing all valves between the main supply line 126 and the branching manifold 1702 in which the clutch and gear controls and actuators are fed directly by the accumulator 130 .

Referring then briefly to FIG. 4 , the main supply line 126 communicates with an inlet port 1308 of the passive feed limit regulator assembly 1310 which includes a generally cylindrical housing 1312 which slidably receives a spool 1314 having a pair of spaced-apart lands 1314 A and 1314 B. The passive feed limit regulator assembly 1310 also includes an outlet port 1316 which is in fluid communication with the branching manifold 1702 . The outlet port 1316 is also in fluid communication with a control port 1322 through a flow restricting orifice 1324 . At the opposite end of the passive feed limit regulator assembly 1310 from the control port 1322 is a compression spring 1326 . The compression spring 1326 biases and drives the spool 1314 to the left as illustrated in FIGS. 3A and 4 , which allows communication and fluid flow between the inlet port 1308 and the outlet port 1316 . As hydraulic pressure builds in the outlet port 1316 , the manifold 1702 and the control port 1322 , the spool 1314 will be driven to the right in FIGS. 3A and 4 , which closes off fluid flow between the inlet port 1312 and the outlet port 1316 . Sufficient pressure at the control port 1322 will completely terminate the flow of hydraulic fluid through the feed limit regulator assembly 1310 .

Referring again to FIGS. 3A , 3 B and 3 C, the second embodiment 1800 includes the components associated with activation of the first clutch 64 A, such as the first electric pressure or flow clutch control solenoid valve 154 , the orifice 156 , the first clutch piston and cylinder assembly 160 and the first clutch pressure limit control valve 166 as well as the components associated with activation of the second clutch 64 B, such as the second electric pressure or flow clutch control solenoid valve 204 , the orifice 206 , the second clutch piston and cylinder assembly 210 and the second clutch pressure limit control valve 216 .

The second embodiment 1800 also includes the fourth branch 1702 D of the manifold 1702 which communicates with the inlet port 1030 A of the first electric pressure or flow control solenoid valve 1030 . The outlet port 1030 B is connected by the first hydraulic line 1432 with the first inlet port 1040 A of the first spool or logic control valve 1040 . The exhaust port 1030 C communicates with the sump 102 . The fourth branch 1702 D of the manifold 1702 also communicates with the inlet port 1430 A of the second electric pressure or flow control solenoid valve 1430 . The second hydraulic line 1434 communicates between the outlet port 1430 B and the second inlet port 1040 B of the first spool or logic valve 1040 . The exhaust port 1430 C communicates with the sump 102 .

The first spool or logic valve 1040 includes the control port 1040 C which is selectively supplied pressurized hydraulic fluid from the outlet port 1042 B of the first two position solenoid (on-off) valve 1042 which receives pressurized hydraulic fluid at the inlet port 1042 A from the third branch 1702 C of the manifold 1702 . Energizing the first two position solenoid valve 1042 translates the spool of the first logic valve 1040 to the left in FIG. 3B ; de-energizing it translates the spool of the first logic valve 1040 to the right and returns hydraulic fluid to the sump 102 through the exhaust port 1042 C. The first spool or logic valve 1040 also includes three exhaust ports 1040 D, 1040 E, and 1040 F disposed between and alternating with the inlet ports 1040 A and 1040 B which communicate with the sump 102 .

The hydraulic line 1046 communicates between the first outlet port 1040 G of the first spool or logic valve 1040 and the first input port 1060 A of the second spool or logic valve 1060 . Another hydraulic line 1048 communicates between the third outlet port 1040 H and the second input port 1060 B. Another hydraulic line 1052 connects the fourth outlet port 1040 J to the first inlet port 1090 A of the third spool or logic valve 1090 . Another hydraulic line 1054 connects the second outlet port 1040 I to the second inlet port 1090 B of the third spool or logic valve 1090 .

The second spool or logic valve 1060 includes the control port 1060 C, the second two position (on-off) solenoid valve 1062 and three exhaust ports 1060 D, 1060 E, and 1060 F which communicate with the sump 102 . The inlet port 1062 A of the second two position solenoid valve 1062 receives hydraulic fluid through the third branch 1702 C of the manifold 1702 and selectively supplies it through the outlet port 1062 B to the control port 1060 C of the second logic valve 1060 . Energizing the second two position solenoid valve 1062 translates the spool of the second logic valve 1060 to the left in FIG. 3B ; de-energizing it translates the spool of the second logic valve 1060 to the right, returning hydraulic fluid through the exhaust port 1060 C to the sump 102 .

The two hydraulic lines 1064 and 1073 communicate between the first and the third outlet ports 1060 G and 1060 H of the second logic valve 1060 and two ports 1068 A and 1068 B in opposite ends of the first, preferably dual area piston and cylinder assembly 1070 which includes the cylinder 1068 and the piston 1072 . The first piston and cylinder assembly 1070 translates the first shift rail and fork assembly 84 A. The two hydraulic lines 1074 and 1083 communicate between the second and the fourth outlet ports 1060 I and 1060 J of the second logic valve 1060 and the ports 1078 A and 1078 B at opposite ends of the second piston and cylinder assembly 1080 which includes the cylinder 1078 and the piston 1082 . The second piston and cylinder assembly 1080 translates the second shift rail and fork assembly 84 B.

The third spool or logic valve 1090 includes the control port 1090 C and three exhaust ports 1090 D, 1090 E, and 1090 F which communicate with the sump 102 . The inlet port 1092 A of the third two position solenoid valve 1092 receives hydraulic fluid through the fifth branch 1072 E of the manifold 1072 . Energizing the third two position solenoid valve 1092 translates the spool of the third logic valve 1090 to the left in FIG. 3C ; de-energizing it translates the spool of the third logic valve 1090 to the right, returning hydraulic fluid to the sump 102 through the exhaust port 1092 C. The pair of hydraulic lines 1094 and 1103 communicate between the first and the third outlet ports 1090 G and 1090 H of the third logic valve 1090 and the ports 1098 A and 1098 B at opposite ends of the third, preferably dual area piston and cylinder assembly 1100 which includes the cylinder 1098 and the piston 1102 . The third piston and cylinder assembly 1100 translates the third shift rail and fork assembly 94 A.

The hydraulic line 1104 connects the second outlet port 1090 I of the third logic valve 1090 to the first inlet port 1530 A of the fourth spool or logic valve 1530 and the hydraulic line 1113 connects the fourth outlet port 1090 J to the second inlet port 1530 B. The control port 1530 C at the right end of the fourth logic valve 1530 is selectively supplied with pressurized hydraulic fluid from the outlet port 1062 B of the second two position (on-off) solenoid valve 1062 (illustrated in FIG. 3B ) in a line 1532 . Thus, the spool of the fourth logic valve 1530 translates in unison with the spool of the second logic valve 1060 . When the second two position solenoid valve 1062 is energized, both spools translate to the left as viewed in FIGS. 3B and 3C . When the second two position solenoid valve 1062 is de-energized, both spools translate to the right.

The fourth logic valve 1530 includes the first outlet port 1530 G which communicates through the line 1536 to the port 1538 A at one end of the fourth piston and cylinder assembly 1540 having the cylinder 1538 and the piston 1542 which is coupled to the fourth shift rail and fork assembly 94 B. The port 1538 B at the other end of the fourth piston and cylinder assembly 1542 communicates through the line 1544 to the third outlet port 1530 H. Similarly, the second outlet port 1530 I communicates through the line 1546 to the port 1548 A at one end of the fifth, preferably dual area piston and cylinder assembly 1550 having the cylinder 1548 and the piston 1552 which is coupled to the fifth shift rail and fork assembly 94 C. The other end of the fifth piston and cylinder assembly 1550 includes the port 1538 B which communicates through the line 1554 to the fourth outlet port 1530 J.

It will be appreciated that the two hydraulic control systems of the present invention achieve significant improvements in reduced energy consumption and shift performance not only because of the incorporation of the dedicated electric pump and accumulator but also because of the use of pressure and flow control solenoid valves which allow the majority of the hydraulic system components to be turned off in normal, steady-state, operation. Additionally, these solenoid valves and the linear position sensors on each piston and cylinder shift actuator assembly which provide real time data to the transmission control module regarding the instantaneous positions of the actuators, shift rails and clutches, achieve gear selection and clutch operation that is rapid, positive and efficient without overshoot and wasted energy.

It should also be appreciated that the incorporation of the branching manifold and the supply of hydraulic fluid to both of the clutch operators in both embodiments provides improved operational flexibility as pressurized hydraulic fluid is essentially always available to engage the transmission clutches.

Similarly, the configurations of the two embodiments and the position feedback provided by the linear position sensors permits and facilitates rapid gear sequencing and improved, i.e., reduced, shift times.

The description of the invention is merely exemplary in nature and 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.