Patent Publication Number: US-2006006042-A1

Title: Dual clutch transmission clutch cooling circuit

Description:
BACKGROUND OF THE INVENTION  
      1. Field of the Invention  
      The present invention relates, generally to a cooling circuit for a dual clutch transmission and, more specifically, to a hydraulic circuit used to controlling the flow of cooling fluid provided to each of the two clutches of a dual clutch transmission.  
      2. Description of the Related Art  
      Generally speaking, land vehicles require a powertrain consisting of three basic components. These components include a power plant (such as an internal combustion engine), a power transmission, and wheels. The power transmission component is typically referred to simply as the “transmission.” Engine torque and speed are converted in the transmission in accordance with the tractive-power demand of the vehicle. Presently, there are two typical transmissions widely available for use in conventional motor vehicles. The first, and oldest type is the manually operated transmission. These transmissions include a foot-operated start-up or launch clutch that engages and disengages the driveline with the power plant and a gearshift lever to selectively change the gear ratios within the transmission. When driving a vehicle having a manual transmission, the driver must coordinate the operation of the clutch pedal, the gearshift lever and the accelerator pedal to achieve a smooth and efficient shift from one gear to the next. The structure of a manual transmission is simple and robust and provides good fuel economy by having a direct power connection from the engine to the final drive wheels of the vehicle. Additionally, since the operator is given complete control over the timing of the shifts, the operator is able to dynamically adjust the shifting process so that the vehicle can be driven most efficiently. One disadvantage of the manual transmission is that there is an interruption in the drive connection during gear shifting. This results in losses in efficiency. In addition, there is a great deal of physical interaction required on the part of the operator to shift gears in a vehicle that employs a manual transmission.  
      The second, and newer choice for the transmission of power in a conventional motor vehicle is an automatic transmission. Automatic transmissions offer ease of operation. The driver of a vehicle having an automatic transmission is not required to use both hands, one for the steering wheel and one for the gearshift, and both feet, one for the clutch and one for the accelerator and brake pedal in order to safely operate the vehicle. In addition, an automatic transmission provides greater convenience in stop and go situations, because the driver is not concerned about continuously shifting gears to adjust to the ever-changing speed of traffic. Although conventional automatic transmissions avoid an interruption in the drive connection during gear shifting, they suffer from the disadvantage of reduced efficiency because of the need for hydrokinetic devices, such as torque converters, interposed between the output of the engine and the input of the transmission for transferring kinetic energy therebetween. In addition, automatic transmissions are typically more mechanically complex and therefore more expensive than manual transmissions.  
      For example, torque converters typically include impeller assemblies that are operatively connected for rotation with the torque input from an internal combustion engine, a turbine assembly that is fluidly connected in driven relationship with the impeller assembly and a stator or reactor assembly. These assemblies together form a substantially toroidal flow passage for kinetic fluid in the torque converter. Each assembly includes a plurality of blades or vanes that act to convert mechanical energy to hydrokinetic energy and back to mechanical energy. The stator assembly of a conventional torque converter is locked against rotation in one direction but is free to spin about an axis in the direction of rotation of the impeller assembly and turbine assembly. When the stator assembly is locked against rotation, the torque is multiplied by the torque converter. During torque multiplication, the output torque is greater than the input torque for the torque converter. However, when there is no torque multiplication, the torque converter becomes a fluid coupling. Fluid couplings have inherent slip. Torque converter slip exists when the speed ratio is less than 1.0 (RPM input&gt; than RPM output of the torque converter). The inherent slip reduces the efficiency of the torque converter.  
      While torque converters provide a smooth coupling between the engine and the transmission, the slippage of the torque converter results in a parasitic loss, thereby decreasing the efficiency of the entire powertrain. Further, the torque converter itself requires pressurized hydraulic fluid in addition to any pressurized fluid requirements for the actuation of the gear shifting operations. This means that an automatic transmission must have a large capacity pump to provide the necessary hydraulic pressure for both converter engagement and shift changes. The power required to drive the pump and pressurize the fluid introduces additional parasitic losses of efficiency in the automatic transmission.  
      In an ongoing attempt to provide a vehicle transmission that has the advantages of both types of transmissions with fewer of the drawbacks, combinations of the traditional “manual” and “automatic” transmissions have evolved. Most recently, “automated” variants of conventional manual transmissions have been developed which shift automatically without any input from the vehicle operator. Such automated manual transmissions typically include a plurality of power-operated actuators that are controlled by a transmission controller or some type of electronic control unit (ECU) to automatically shift synchronized clutches that control the engagement of meshed gear wheels traditionally found in manual transmissions. The design variants have included either electrically or hydraulically powered actuators to affect the gear changes. However, even with the inherent improvements of these newer automated transmissions, they still have the disadvantage of a power interruption in the drive connection between the input shaft and the output shaft during sequential gear shifting. Power interrupted shifting results in a harsh shift feel that is generally considered to be unacceptable when compared to smooth shift feel associated with most conventional automatic transmissions.  
      To overcome this problem, other automated manual type transmissions have been developed that can be power-shifted to permit gearshifts to be made under load. Examples of such power-shifted automated manual transmissions are shown in U.S. Pat. No. 5,711,409 issued on Jan. 27, 1998 to Murata for a Twin-Clutch Type Transmission, and U.S. Pat. No. 5,966,989 issued on Apr. 04, 2000 to Reed, Jr. et al for an Electro-mechanical Automatic Transmission having Dual Input Shafts. These particular types of automated manual transmissions have two clutches and are generally referred to simply as dual, or twin, clutch transmissions. The dual clutch structure is most often coaxially and co-centrically configured so as to derive power input from a single engine flywheel arrangement. However, some designs have a dual clutch assembly that is coaxial but with the clutches located on opposite sides of the transmissions body and having different input sources. Other designs are known in which the two clutches are non-coaxial with differing input soucres. Regardless, the layout is the equivalent of having two transmissions in one housing, namely one power transmission assembly on each of two input shafts concomitantly driving one output shaft. Each transmission can be shifted and clutched independently. In this manner, uninterrupted power upshifting and downshifting between gears, along with the high mechanical efficiency of a manual transmission is available in an automatic transmission form. Thus, significant increases in fuel economy and vehicle performance may be achieved through the effective use of certain automated manual transmissions.  
      The dual clutch transmission structure may include two dry disc clutches each with their own clutch actuator to control the engagement and disengagement of the two-clutch discs independently. While the clutch actuators may be of the electromechanical type, since a lubrication system within the transmission requires a pump, some dual clutch transmissions utilize hydraulic shifting and clutch control. These pumps are most often gerotor types, and are much smaller than those used in conventional automatic transmissions because they typically do not have to supply a torque converter. Thus, any parasitic losses are kept small. Shifts are accomplished by engaging the desired gear prior to a shift event and subsequently engaging the corresponding clutch. With two clutches and two inputs shafts, at certain times, the dual clutch transmission may be in two different gear ratios at once, but only one clutch will be engaged and transmitting power at any given moment. To shift to the next higher gear, first the desired gears on the input shaft of the non-driven clutch assembly are engaged, then the driven clutch is released and the non-driven clutch is engaged.  
      This requires that the dual clutch transmission be configured to have the forward gear ratios alternatingly arranged on their respective input shafts. In other words, to perform up-shifts from first to second gear, the first and second gears must be on different input shafts. Therefore, the odd gears will be associated with one input shaft and the even gears will be associated with the other input shaft. In view of this convention, the input shafts are generally referred to as the odd and even shafts. Typically, the input shafts transfer the applied torque to a single counter shaft, which includes mating gears to the input shaft gears. The mating gears of the counter shaft are in constant mesh with the gears on the input shafts. The counter shaft also includes an output gear that is meshingly engaged to a gear on the output shaft. Thus, the input torque from the engine is transferred from one of the clutches to an input shaft, through a gear set to the counter shaft and from the counter shaft to the output shaft.  
      Gear engagement in a dual clutch transmission is similar to that in a conventional manual transmission. One of the gears in each of the gear sets is disposed on its respective shaft in such a manner so that it can freewheel about the shaft. A synchronizer is also disposed on the shaft next to the freewheeling gear so that the synchronizer can selectively engage the gear to the shaft. To automate the transmission, the mechanical selection of each of the gear sets is typically performed by some type of actuator that moves the synchronizers. A reverse gear set includes a gear on one of the input shafts, a gear on the counter shaft, and an intermediate gear mounted on a separate counter shaft meshingly disposed between the two so that reverse movement of the output shaft may be achieved.  
      While these power-shift dual clutch transmissions have overcome several drawbacks associated with conventional transmissions and the newer automated manual transmissions, it has been found that controlling and regulating the automatically actuated dual clutch transmission to achieve the desired vehicle occupant comfort goals is a complicated matter. There are a large number of events to properly time and execute within the transmission for each shift to occur smoothly and efficiently. In addition, the clutch and complex gear mechanisms, working within the close confines of the dual clutch transmission case, generate a considerable amount of heat. The heat build-up is aggravated by the nature of the clutch mechanisms themselves, each of which are typically constructed of two series of plates, or discs, one set connected in some manner to the output of the engine and the second attached to an input shaft of the transmission. Each of the set of plates include friction material. The clutch plates and discs are pressed together under pressure to a point at which the plates and discs make a direct physical connection. The clutch may be designed for a full “lock-up” of the plates and discs, or may be designed with a certain amount of “limited slip”. Regardless, the slipping of the friction plates within a friction type clutch, whether from a designed limited slip or the normal uncontrolled slipping that occurs during clutch engagement and disengagement, generates heat that needs to be dissipated. A considerable amount of heat can be generated in the typical dual clutch transmission utilizing a combined coaxial clutch assembly wherein the one clutch fits within the second clutch.  
      In order to provide sufficient cooling to the clutch assemblies of the conventional dual clutch transmission, the clutch assemblies are usually bathed in transmission fluid in a generally uncontrolled manner. While this approach has generally worked for its intended purpose, disadvantages remain. Specifically, these types of conventional clutch cooling hydraulic circuits have failed either to adequately provide for proper cooling and heat reduction of the clutches of the dual clutch transmission or have resulted in producing large efficiency losses by excessively flooding of the clutch assemblies with fluid.  
      Of late, newer approaches in the structure of hydraulic circuits for clutch cooling have been proposed in the related art that offer improvements, but are still limited in their cooling capacity. In one example, a hydraulic circuit is employed that provides pressurized cooling oil directly from the pump to the clutches to maintain adequate flow and pressure and then passes the cooling fluid through a cooler device only on the return line to the sump. However, since the cooling fluid is not cooled on its way to the clutches, the cooling capacity of the applied cooling fluid is somewhat limited and larger quantities of fluid must be employed to adequately cool the clutch. This generally requires a larger pump and larger supply lines. In another example, the pressurized cooling fluid is pumped through a cooling unit prior to delivery to the clutches. However, this causes a distinct pressure drop and the system must contend with the flow limitations of the cooling unit.  
      These conventional clutch cooling approaches also use a single hydraulic circuit to supply cooling oil or fluid from the cooler device to the clutches. This causes the clutches to suffer inadequate and inefficient heat removal. Furthermore, the inadequacy of these conventional hydraulic circuits is also exaggerated under clutch high loading conditions where excessively high heat is built up rapidly in the active clutch. These inherently inadequate cooling circuit strategies lead to shortened component life and ultimate failure of the clutch assemblies within the dual clutch transmission. Similarly, inadequate cooling is responsible for rapid breakdown of the physical properties of the transmission fluid, which can cause failure of the other components within the transmission. Further, the conventional hydraulic circuits that excessively flood the clutch assemblies with cooling fluid also cause unnecessary clutch drag and put excessive demands on the pump resulting in poor clutch life and lower fuel efficiencies.  
      Accordingly, there remains a need in the related art for an improved hydraulic circuit to provide cooling fluid to the clutch assemblies of the dual clutch transmissions. Specifically, there is a need for a cooling circuit that provides cooling fluid from a cooling unit to the clutches normally and supplementally provides high flow cooling fluid from the pump under high load conditions. In this manner, clutch heat is dissipated by cooled cooling fluid normally and a supplemental flow of cooling fluid that is not limited by the flow restrictions of the cooling unit is further provided when the clutch comes under high load.  
     SUMMARY OF THE INVENTION  
      The disadvantages of the related art are overcome by the hydraulic circuit of the present invention for controlling the application of pressurized cooling fluid to the clutches of a dual clutch transmission. The circuit includes a cooling unit in fluid communication with a source of the pressurized cooling fluid and adapted to exchange heat from the cooling fluid with another media. The circuit also includes at least one regulator in fluid communication with the source of the pressurized cooling fluid and separately in fluid communication with the cooling unit, and further in fluid communication with the clutches. The regulator is adapted to operatively provide the cooling fluid to the clutches. The circuit also includes at least one control actuator adapted to selectively control the fluid regulator to provide a first variable predetermined amount of cooling fluid from the cooling unit to the clutches as primary cooling, and to provide a second variable predetermined amount of cooling fluid from the source to the clutches thereby supplementing the cooling fluid from the cooling unit.  
      Thus, the present invention overcomes the limitations of the current hydraulic circuits for clutch cooling in a dual clutch transmission by providing a controlled flow of cooling fluid from the cooling unit normally and providing a supplemental flow of cooling fluid directly from the pump when needed. In this manner, the use of directly pumped supplemental cooling fluid allows the size of the pump and cooling unit to remain relatively small. This provides efficiency and cost savings while properly protecting the clutches from heat damage. The present invention also allows immediate cooling flow to the clutches during periods of high loading when clutch temperatures rapidly increase. Furthermore, the present invention is adaptable to provide primary and supplemental clutch cooling when the structure of the transmission employs either co-centric or parallel clutch assemblies.  
      Other objects, features, and advantages of the present invention will be readily appreciated, as the same becomes better understood after reading the subsequent description taken in connection with the accompanying drawings. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       FIG. 1  is a schematic illustration of a dual clutch transmission of the type that may employ the clutch cooling circuit of the present invention;  
       FIG. 2  is a schematic illustration of the hydraulic cooling circuit of the present invention for cooling the clutches of a dual clutch transmission; and  
       FIG. 3  is a schematic illustration of another example of the hydraulic cooling circuit of the present invention for controlling a separate and independent flow of cooling fluid to each of the clutches of a dual clutch transmission. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT(S)  
      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, co-centric 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 .  
      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 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 . It should be appreciated that, although not illustrated herein, the first and second clutch mechanisms  32 ,  34  and the first and second input shafts  14 ,  16  may be physically arranged within the transmission in a parallel, rather than co-centric structure.  
      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 second input gear  40 , a fourth input gear  44 , a sixth input gear  46 , and a reverse input gear  48 . As shown in  FIG. 1 , the second 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 sixth 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.  
      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 second counter gear  54 , a third counter gear  56 , a fourth counter gear  58 , a sixth counter gear  60 , and a reverse counter gear  62 . The counter shaft  18  fixedly retains the fourth counter gear  58  and sixth counter gear  60 , while first, second, 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.  
      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 , there are four synchronizers  74 ,  76 ,  78 , and  80  that 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 of a center neutralized position to the right and engage another gear to its respective shaft when moved to the left. Specifically with reference to the example illustrated in  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 second 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 sixth 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 . It should be appreciated that this example of the dual clutch transmission is representative and that other gear set, synchronizer, and shift actuator arrangements are possible within the dual clutch transmission  10  as long as the even and odd gear sets are disposed on opposite input shafts.  
      It should be further appreciated that the operation of the dual clutch transmission  10  is managed by some type of control device such as an electronic control unit (ECU) that oversees the functioning of the transmission  10 , or by an electronic control unit for the vehicle in which the dual clutch transmission  10  maybe installed. Regardless, there exists a control device, beyond the scope of this invention, that controls and operates the dual clutch transmission through a stored control scheme or series of control schemes of which the present invention is merely a part. The control device having the capability of providing the proper voltages, signals, and/or hydraulic pressures to operate the transmission  10  and particularly the clutch engagement functions. Thus, the control method of the present invention as described below is merely a portion, such as a sub-routine, or series of sub-routines, of a larger control scheme within the ECU.  
      The first and second clutch mechanisms  32  and  34  of the dual, coaxial clutch assembly  12  are operatively engaged and disengaged in a coordinated manner relative to the actuator of the various gear sets by the synchronizer  24  to selectively transfer torque to the output shaft  20 . By way of example, if torque is being transferred to the drive wheels of the vehicle to initiate movement from a standing start, the lowest, or first, gear ratio of the dual clutch transmission  10  will likely be engaged. Therefore, as seen in  FIG. 1 , synchronizer  78  will be driven to the left to engage the first counter gear  52  to the counter shaft  18  and the first clutch mechanism  32  will be engaged to transfer torque from the engine to the output shaft  20  through the first gear set. When vehicle speed increases and the ECU determines that the conditions require a shift to the second gear set, synchronizer  80  will first be driven to the right to engage the second counter gear  54  to the counter shaft  18 . Then the second clutch mechanism  34  will be engaged as the first clutch mechanism  32  is disengaged. In this manner, a powershift, where no power interruption occurs, is affected. This powershift changeover of the clutches  32  and  34  occurs for each shift change of the dual clutch transmission  10 . As the inactive clutch (now the on-coming clutch) is engaged, the load applied causes a surge of power to be transferred across the clutch with an accompanying generation of heat from the slip that occurs across the clutch. The temperature of the on-coming clutch rapidly increases, or spikes, to a point where the clutch plates or the friction material could be damaged if proper cooling is not provided. Additionally, the heat build-up, if not properly dissipated, will greatly increase the overall temperature of the dual clutch transmission  10  and may cause the damaging effects mentioned above. Simultaneously, while the temperature of the on-coming clutch is sharply rising, the disengaging, or off-going, clutch will cease transmitting torque. With the removal of the load, the disengaged clutch will stop generating heat, thus sharply lowering its cooling requirement.  
      The hydraulic circuit of the present invention is generally indicated at  220  in  FIG. 2 . The hydraulic circuit  220  includes a cooling unit  222  in fluid communication with a source of pressurized cooling fluid, through line  223  that is adapted to exchange heat from said cooling fluid with another media. As shown, pump  94  provides a source of pressurized cooling fluid, which is drawn from the sump  90 . The hydraulic circuit  220  also includes at least one clutch cooling fluid regulator generally indicated at  224  and at least one control actuator  226 . The clutch cooling fluid regulator  224  is in fluid communication with the source of the pressurized cooling fluid from the pump  94  through hydraulic line  130 . More specifically, as illustrated in  FIG. 2 , the bulk of the cooling fluid is maintained in a sump  90 . A pump  94  is used to provide positive pressure to the cooling fluid as it is drawn from the sump  90  through a filter  92 . The pump output charges a main pressure line  96  that feeds the various components of the hydraulic circuit  220 .  
      A pump pressure relief valve  100  is operatively connected in fluid communication with the main pressure line  96  to provide a maximum upper limit for the positive pressure provided by the pump  94 . The pressure relief valve  100  is moved to its closed position, as shown in  FIG. 2 , by a biasing member  102 . The biasing member  102  has a pre-determined spring force that corresponds to the desired maximum system pressure. In the advent that the pressure in the main pressure line  96  exceeds the pre-determined maximum, the excessive pressure, as applied to the right side of the valve, will move the valve member  104  of the pressure relief valve  100  to the left, overcoming the spring force of biasing member  102 . In this manner, the previously blocked relief passage  106  is opened to the sump  90  allowing the excessive pressure to bleed off and dropping the pressure in the main pressure line  96  until the biasing member  102  can force the valve member  104  of the relief valve  100  back to its closed position.  
      The main pressure line  96  also feeds the main pressure regulator  110 . The main pressure regulator  110  maintains the pressure in the main pressure line  96  at a pre-determined operating pressure, or setpoint. The main pressure regulator  110  is shown in  FIG. 2  in its closed position and includes a biasing member  112  and a main valve member, schematically indicated at  114  with internal flow passages, generally indicated at  116 . The flow passages  116  are shown in left  118 , middle  120 , and right  122  positions of the valve member  114 . The pressure in the main pressure line  96  is supplied to the right side of the main regulator valve through a flow restrictor  124  that reduces the flow volume but maintains the applied pressure. With the pump  94  operating, the pressure delivered to the right side of the main pressure regulator  110  overcomes the spring force of the biasing member  112  and moves the valve member  114  of the regulator  110  to the right from the closed left position  118  to the middle operable position  120 . Here, the internal flow passages  116  of the middle operable position  120  allow the flow of cooling fluid in the main pressure line  96  to flow into the regulated line  130 . A regulating control line  132 , shown as a dotted line in  FIG. 2 , provides a controllable biasing force to the left side of the main pressure regulator  110 . The regulating control line  132  delivers a portion of the pressure from the main pressure line  96  to the left side of the regulator  110  under the control of the line pressure solenoid  136 .  
      The line pressure solenoid  136  is electrically operated by and engine control unit (ECU) to set the regulated pressure setpoint within the regulating circuit  82  and then to maintain the desired pressure by regulating the output pressure to the setpoint. The line pressure solenoid  136  supplies a varying portion of the available main pressure through the regulating line  132  to the main pressure regulator  110  by bleeding off some portion of the main pressure as supplied through flow restrictors  138  and filter  140  to the sump  90 . In this manner, the line pressure solenoid  136  sets the desired output pressure setpoint for the main pressure regulator  110 . The line pressure solenoid  136  then varies the pressure in the regulating line  132  to maintain the output pressure delivered from the main pressure regulator  110  about the desired output pressure setpoint while accounting for fluctuations in the output pressure due to downstream pressure changes.  
      The main pressure regulator  110  also provides control over rapid increases, or surges, in the main pressure line  96  that exceeds the immediate correction ability of the line pressure solenoid  136 . The right position  122  of the valve member  114  opens additional flow passages  116  that not only allow for the continued flow of fluid through the regulator  110  to the regulated line  130 , but also allow a portion of the increased flow to pass to the suction line  144 . The suction line  144  normally remains closed off by the left and middle positions  118 ,  120  of the valve member  114 . However, when a sharp or rapid increase of pressure in the main pressure line  96  drives the valve member  114  all the way to the left, a corrective portion of the flow is fed back to the suction side of the pump  94 . As the suction line  144  bleeds off the surge of excessive pressure flow, the regulator valve member  114  moves back to the middle operative position  120 .  
      Thus, a source of the pressurized cooling fluid is provided to the clutch cooling fluid regulator  224 . The clutch cooling fluid regulator  224  is also in fluid communication with the cooling unit  222  through line  225 . Further, clutch cooling fluid regulator  224  is in fluid communication with the clutches  32 ,  34  such that the clutch cooling fluid regulator  224  is adapted to operatively provide cooling fluid to the clutches. Those having ordinary skill in the art will appreciate from the following discussion that the actual means of delivering the cooling fluid to the clutch mechanisms  32  and  34  is unimportant. Any of the various approaches for delivering fluid to clutch disks and plates, such as internal fluid passages extending through the input shafts or appropriately placed spray orifices as commonly known in the art may be employed in connection with the present invention illustrated in  FIG. 2 .  
      The control actuator  226  is adapted to selectively control the clutch cooling fluid regulator  224  to cause the clutch cooling fluid regulator  224  to provide a first variable predetermined amount of cooling fluid from the cooling unit  222  to the clutches as primary cooling. Further, the control actuator  226  is also adapted to cause the clutch cooling fluid regulator  224  to provide a second variable predetermined amount of cooling fluid from the source to the clutches thereby bypassing the cooling unit  222  as supplemental cooling.  
      More particularly, as shown in  FIG. 2  the clutch cooling fluid regulator  224  provides a controlled flow of cooling fluid to the first and second clutches  32  and  34  of the dual clutch transmission  10  through output line  228 . The clutch cooling fluid regulator  224  includes a biasing member  230  and a main valve body  232  having internal flow passages and a valve member  234  that is moveable between predetermined positions within the valve body  232  to open or close the valve flow passages. The valve member  234  has a first valve flow section  236  and a second valve flow section  238 . The regulator control line  240  (shown as a dotted line) provides an actuating force to the right sides of the clutch cooling fluid regulator  224 . The regulator control line  240  delivers a portion of the pressure from the main pressure line  96  under the control of the control actuator  226 .  
      The control actuator  226  is electrically operated by the ECU to control the delivery of cooling fluid to the clutches by regulating the cooling fluid flow through the clutch cooling fluid regulator  224 . The control actuator  226  supplies a variable portion of the available main pressure through the regulator control lines  240  to the clutch cooling fluid regulator  224  by operatively bleeding off some portion of the main pressure through flow restrictor  242  and filter  244  to the sump  90 . The pressure supplied to the right side of the control actuator  226  moves the valve member  234  to the left. Under clutch low load or low stress conditions, the control actuator  226  causes the cooling fluid regulator  224  to provide a regulated supply of pressurized cooling fluid from the cooling unit  222  as the primary cooling source through the first valve flow section  236  of valve member  234 . Under clutch high stress or heavy load conditions, the control actuator  226  causes the cooling fluid regulator  224  to provide a regulated supply of cooling fluid from the cooling unit  222  as the primary cooing source and additionally provide cooling fluid directly from the pump  94  through the second valve flow section  238 .  
      The decision of whether the cooling fluid flow should be delivered from either the first flow section  236  or the second valve flow section  238  is predetermined and under a higher level control through the ECU. The predetermined threshold of when it is necessary to supply supplemental cooling directly from the pump  94  (through second valve flow section  238 ) is a design consideration beyond the scope of the present invention. However, it should be appreciated that the maximum pressure and flow available for the primary source of cooling through the cooling unit  222  is held to the pressure and flow limitations of the cooling unit itself. Thus, any cooling requirements that exceed the maximum capacity of the cooling unit  222  will likely require a supplemental flow directly from the pump  94 .  
      Flow restrictors  250  stabilize the applied pressure and prevent surges of cooling fluid to the clutches as the supply flow is regulated. A biasing pressure from the supply of the cooling unit  222  is applied to the right side of the valve member  234  through a restrictor  252  to supplement the biasing force of the biasing member  230 . This assists the biasing member  230  in returning the valve member  234  to the closed position when the applied controlling pressure through the regulator control line  240  is removed or drops. It should be noted that line  225  from the cooling unit  222  may also provide a portion of the pressurized fluid to other parts of the dual clutch transmission  10  for any of a variety of purposes such as cooling and lubrication of additional components, as indicated by  254 .  
      It should be appreciated that other routing arrangements may also be employed without departing from the scope of the present invention as long as the cooling unit  222  precedes the primary input line  225  to the fluid regulator  224  and thereby the clutches  32 ,  34 . For example, the input pressure to the cooling unit  222  may be separately regulated from the line pressure, or the input pressure to the fluid regulator  222  may be unregulated. Furthermore, the cooing unit  222  may be a heat exchanger physically disposed outside of the transmission and exposed to an air stream to allow heat to transfer from the cooling fluid to the air stream. The cooling unit may also be outside of the transmission and physically disposed within another heat exchanger within the vehicle, such as the vehicle&#39;s main radiator so that the cooling unit is exposed to the liquid media of the radiator to allow heat to transfer from said cooling fluid to the liquid media.  
      Depending on the physical structure and particular operative design considerations of the dual clutch transmission, the hydraulic circuit of the present invention may also include a second clutch cooling fluid regulator and a second control actuator. More specifically, as previously discussed, dual clutch transmissions other than the type generally described with regard to  FIG. 1  are known to have non-co-centric clutch configurations. In these cases, proper cooling of the separate clutch mechanisms may be better served by supplying and controlling the flow of cooling fluid to each of the clutches independently.  
      As such and referring now to  FIG. 3 , where like numerals incremented by  100  are used to designate like structure, another embodiment of the hydraulic circuit of the present invention is generally indicated at  320 . The hydraulic circuit  320  is adapted to separately and independently provide a flow of cooling fluid to the two clutches of a dual clutch transmission. The hydraulic circuit  320  includes a first clutch cooling fluid regulator  324 , a first control actuator  326 , a second clutch cooling fluid regulator  424 , and a second control actuator  426 . The fluid regulators  324  and  424  are each in fluid communication with the source of the pressurized cooling fluid from the pump  94  and are each separately in fluid communication with the cooling unit  322 . Further, the first fluid regulator  324  is in fluid communication with one of the clutches ( 32 ) and the second fluid regulator  424  is in fluid communication with the second of the clutches ( 34 ) such that the first and second fluid regulators  324 ,  424  are adapted to operatively and separately provide cooling fluid to each of the clutches.  
      The control actuators  326  and  426  are adapted to selectively control their respective fluid regulators  324  and  424  to cause the fluid regulators to separately provide a first variable predetermined amount of cooling fluid from the cooling unit  222  to the respective clutches as primary cooling, and also to cause the fluid regulators to separately provide a second variable predetermined amount of cooling fluid from the source to the respective clutches thereby bypassing the cooling unit  322  as a supplemental cooling source.  
      The clutch cooling fluid regulators  324  and  424  are substantially similar to each other and to the single clutch cooling fluid regulator  224  described in reference to  FIG. 2 . Each clutch cooling fluid regulator provides a controlled flow of cooling fluid to one of the first and second clutches  32  and  34  of the dual clutch transmission  10  through output lines  328  and  428 , respectively. The fluid regulators  324  and  424  each respectively include a biasing member  330 ,  430  and a main valve body  332 ,  432  having internal flow passages and a valve member  334 ,  434  that is moveable between predetermined positions within the valve body  332 ,  432  to open or close the valve flow passages. The valve members  334 ,  434  have a first valve flow section  336 ,  436  and a second valve flow section  338 ,  438 , respectively. The regulator control lines  340  and  440  (shown as a dotted line) provide an actuating force to the right sides of the clutch cooling fluid regulators  324 ,  424 . The regulator control lines  340  and  440  deliver a portion of the pressure from the main pressure line  96  under the control of the control actuators  326  and  426 , respectively.  
      The control actuators  326  and  426  are electrically operated by the ECU to control the delivery of cooling fluid to each of the clutches independently by regulating the cooling flow through the fluid regulators  324  and  424 . Each of the control actuators  326  and  426  supply a variable portion of the available main pressure through the regulator control lines  340  and  440  to the fluid regulators  324  and  424  by operatively bleeding off some portion of the main pressure through flow restrictors  342 ,  442  and filters  344 ,  444  to the sump  90 . The pressure supplied to the right side of the fluid regulators  324  and  424  moves the valve members  334  and  434  to the left to allow the cooling fluid in the regulated line  130  to pass from the output lines  328  and  428  to the clutches  32 ,  34 . Under low clutch load or low stress conditions, the control actuators  326 , 426  cause the cooling fluid regulators  324 ,  424  to provide a regulated supply of pressurized cooling fluid from the cooling unit  322  as the primary cooling source through the first valve flow sections  336 ,  436  of valve members  334 ,  434 . Under clutch high stress or heavy load conditions, the control actuators  326 ,  426  cause the cooling fluid regulator  324 ,  424  to provide a regulated supply of cooling fluid from the cooling unit  322  as the primary cooling source and additionally provide cooling fluid directly from the pump  94  through the second valve flow sections  338 ,  438 .  
      The decision of whether the cooling fluid flow should be delivered from either the first valve flow sections  336 ,  436  or the second valve flow sections  338 ,  438  of the valve members  334 ,  434  to the respective clutches  32 ,  34  is predetermined and under a higher level control through the ECU. The predetermined threshold of when it is necessary to supply supplemental cooling directly from the pump  94  (through second valve flow sections  338 ,  438 ) is a design consideration that is beyond the scope of the present invention. However, it should be appreciated that the maximum pressure and flow available for the primary source of cooling through the cooling unit  322  is held to the pressure and flow limitations of the cooling unit itself. Thus, any cooling requirements that exceed the maximum capacity of the cooling unit  322  will likely require a supplemental flow directly from the pump  94 . It should be further appreciated that since the control the separate fluid regulators are independent from each other, the cooling fluid flow for each may be separately directed as necessary depending on the differing operating conditions for each clutch.  
      Flow restrictors  350 ,  450  stabilize the applied pressure and prevent surges of cooling fluid to the clutches as the supply flow is regulated. A biasing pressure from the supply of the cooling unit  322  is applied to the right side of the valve members  334 ,  434  through a restrictor  352 ,  452  to supplement the biasing force of the biasing members  330 ,  430 . This assists the biasing member  330 ,  430  in returning the valve member  334 ,  434  to the closed position when the applied controlling pressure through the regulator control line  340 ,  440  is removed or drops. It should be noted that line  254  from the cooling unit  222  may also provide a portion of the pressurized fluid to other parts of the dual clutch transmission  10  for any of a variety of purposes such as cooling and lubrication of additional components.  
      Thus, the present invention overcomes the limitations of the current hydraulic circuits for clutch cooling in a dual clutch transmission by providing a controlled flow of cooling fluid from the cooling unit normally and providing a supplemental flow of cooling fluid directly from the pump when needed. In this manner, the use of directly pumped supplemental cooling fluid allows the size of the pump and cooling unit to remain relatively small. This provides efficiency and cost savings while properly protecting the clutches from heat damage. The present invention also allows immediate cooling flow to the clutches during periods of high loading when clutch temperatures rapidly increase. Furthermore, the present invention is adaptable to provide primary and supplemental clutch cooling when the structure of the transmission employs either co-centric or parallel clutch assemblies.  
      The invention has been described in an illustrative manner. It is to be understood that the terminology that has been used is intended to be in the nature of words of description rather than of limitation. Many modifications and variations of the invention are possible in light of the above teachings. Therefore, within the scope of the claims, the invention may be practiced other than as specifically described.