Patent Publication Number: US-2023151854-A1

Title: Transmission hydraulic system and method for operation of the hydraulic system

Description:
CROSS REFERENCE TO RELATED APPLICATION 
     The present application is a continuation of U.S. Non-Provisional patent application Ser. No. 17/445,814, entitled “TRANSMISSION HYDRAULIC SYSTEM AND METHOD FOR OPERATION OF THE HYDRAULIC SYSTEM”, and filed on Aug. 24, 2021. The entire contents of the above-listed application are hereby incorporated by reference for all purposes. 
    
    
     TECHNICAL FIELD 
     The present description relates generally to a hydraulic system in a transmission providing both clutch actuation and lubrication functionality. 
     BACKGROUND AND SUMMARY 
     Certain transmissions use hydraulically controlled components, such as wet friction clutches, to reduce power interruptions during shifting transients to diminish noise, vibration, and harshness (NVH). Transmission have additionally deployed lubrication systems that direct oil to bearings, clutches, such as friction clutches, and other moving components, to increase component longevity and reduce wear. The lubricant supplied to the friction clutches decreases clutch wear but may also increase drag torque due to the viscous effects of oil that occur when the clutch is disengaged. This drag torque may cause continuous undesired power losses which decreases gearbox efficiency. US 2015/0012191 A1 to Ohya et al. teaches an automatic transmission with a lubricating oil supply system that routes oil to a plurality of clutches and a motor for lubrication and cooling. In the oil supply system, an electronically adjusted flowrate control valve is positioned downstream of a relief valve. Further, in the system, the flowrate control valve is actively commanded by a controller to decrease the pressure of oil delivered to the clutches, particularly during disengagement. 
     The inventors have recognized several issues with Ohya&#39;s oil supply system and other hydraulic systems. Using a hydraulic clutch actuation system that is independent from a clutch lubrication system may demand complex hydraulic conduit routing schemes. The complex hydraulic conduit routing schemes may pose barriers to space efficient integration of both systems into a tightly packaged transmission. Further, higher amounts of computational resources may be devoted to actively controlling the flowrate valve downstream of the relief valve for relief flow adjustment, when compared to system which may use coordinated hydraulic and lubrication system control. Because of the added structural and computational complexity, previous hydraulic systems have been costly and exhibited packaging inefficiencies. Further, Ohya&#39;s oil supply strategy for the clutch may, under some circumstances, fall short of meeting the clutch&#39;s lubrication demands, particularly during shifting transients. 
     To resolve at least a portion of the aforementioned issues, the inventors have developed a hydraulic system in a transmission including a lubrication valve positioned in a lubrication line. The hydraulic system further includes a clutch line in fluidic communication with a multi-disc wet clutch and a passive adjustment device of the lubrication valve. The hydraulic system also includes a clutch control valve designed to adjust the hydraulic pressure in the clutch line. In the system, the passive adjustment device transitions the lubrication valve between a limited flow state and an open flow state based on a pressure of the hydraulic fluid in the clutch line. In this way, the hydraulic system may effectively adjust the lubricant flow to the clutch using a system that may be more space efficient and less costly and complex than systems with lubrication valves that are directly electronically actuated. 
     As one example, the hydraulic system may further include a controller that is designed to determine a desired lubrication valve state based on a transmission gear ratio. The controller is further designed to adjust the pressure in the clutch line by actuation of the clutch control valve to place the lubrication in the limited flow state or the open flow state based on a transmission shift event. For example, the controller may maintain the clutch line pressure above a threshold value when the clutch is engaged and the transmission remains in a desired gear. However, after the transmission is shifted out of the gear, the clutch remains disengaged, and the transmission ratio deviates from the shift point ratio by a predetermined value, the clutch line pressure may be decreased below the threshold value. In this manner, the system delivers a higher lubricant flow to the clutch while the clutch is engaged, for instance, to enhance clutch performance and decrease clutch degradation and deliver a lower lubricant flow to the clutch during disengagement, to decrease drag torque losses and consequently increase transmission efficiency. 
     It should be understood that the summary above is provided to introduce in simplified form a selection of concepts that are further described in the detailed description. It is not meant to identify key or essential features of the claimed subject matter, the scope of which is defined uniquely by the claims that follow the detailed description. Furthermore, the claimed subject matter is not limited to implementations that solve any disadvantages noted above or in any part of this disclosure. 
    
    
     
       BRIEF DESCRIPTION OF THE FIGURES 
         FIG.  1    is a schematic representation of a vehicle transmission with a hydraulic system. 
         FIG.  2    depicts a first example of a hydraulic system with a lubrication valve and a hydraulic clutch. 
         FIG.  3    shows a second example of a hydraulic system with a lubrication valve and a hydraulic clutch. 
         FIG.  4    shows the lubrication valve, depicted in  FIG.  3   , in an open flow state. 
         FIG.  5    shows the lubrication valve, depicted in  FIG.  3   , in a limited flow state. 
         FIG.  6    shows a method for operating a hydraulic system to adjust a state of a lubrication valve. 
         FIG.  7 A  shows an example control strategy for adjusting a lubrication valve. 
         FIG.  7 B  shows a timing diagram depicting a lubrication valve control strategy. 
         FIG.  8    shows a schematic representation of a vehicle with a hydromechanical transmission. 
     
    
    
     DETAILED DESCRIPTION 
     The following description relates to a hydraulic system in a transmission and a control technique for adjusting a lubrication valve state using hydraulic clutch pressure, as determined by a clutch control valve. The hydraulic system includes a pressure controlled lubrication valve that alters a flow of lubricant supplied to the clutch. In the hydraulic system, a clutch line delivers a hydraulic fluid to a multi-disc wet clutch. A branch of the clutch line is routed to the pressure controlled lubrication valve to permit the clutch line pressure to induce opening and closing of the lubrication valve. The lubrication valve may therefore be switched between a limited flow state and an open flow state using clutch line pressure control that is adjusted through operation of the clutch control valve. The clutch line pressure may, for example, be kept above a threshold pressure that maintains the lubrication valve in a high flow state while the clutch is engaged. Conversely, the clutch line pressure may be decreased below the threshold pressure to switch the lubrication valve to a limited flow state (e.g., choked flow state) when the clutch is disengaged. In one specific example, when the transmission&#39;s operating gear ratio is far enough from the gear shift point (corresponding to clutch disengagement), the lubricant valve may be switched to the limited flow state using pressure control. In this way, the lubrication valve is passively and efficiently operated to provide a higher flow of the hydraulic fluid when clutch lubrication demands are higher and then decrease lubricant flow to the clutch when higher lubricant flow is not demanded. Drag torque from the clutch is decreased, as a result, and the transmission efficiency may be correspondingly increased. Further, in some cases, to increase system compactness and efficiently route lubricant to the clutch, the lubrication valve, the lubrication line, and the clutch line may be included within a rotating shaft of the transmission. 
       FIG.  1    schematically illustrates a hydraulic system in a vehicle transmission.  FIG.  2    shows an example hydraulic system having a clutch control valve for actuating a first hydraulic clutch and adjusting a position of a lubrication valve, where the lubrication valve controls a supply of hydraulic fluid to the first hydraulic clutch and a second hydraulic clutch.  FIG.  3    depicts an example of a rotating shaft in a transmission having the lubrication valve positioned therein.  FIGS.  4  and  5    show the lubrication valve of  FIG.  3    in a limited flow state and an open flow state, respectively, as determined by a hydraulic pressure in a clutch line.  FIG.  6    shows an example control routine for operating a hydraulic system based on a gear ratio to adjust a lubrication valve state.  FIGS.  7 A- 7 B  show control and timing diagrams for a hydraulic system.  FIG.  8    schematically depicts a hydromechanical transmission which may include a hydraulic system according to the examples described herein. 
       FIG.  1    schematically illustrates a vehicle system  100  including a vehicle  102  having a powertrain  104 . The vehicle may take a variety of forms in different embodiments such as a light, medium, or heavy duty vehicle. To generate power, the powertrain  104  may include a motive power source  106 . The power source may include an electric motor (e.g., electric motor-generator), an internal combustion engine (ICE) (e.g., a spark and/or compression ignition engine), combinations thereof, or other suitable devices designed to generate rotational energy. The internal combustion engine may include conventional components such as cylinder(s), piston(s), valves, a fuel delivery system, an intake system, an exhaust system, etc. Further, the electric motor may include conventional components such as a rotor, a stator, a housing, and the like for generating mechanical power as well as electrical power during a regeneration mode, in some cases. As such, the powertrain may be utilized in a hybrid or battery electric vehicle. In other examples, however, the vehicle may solely use an internal combustion engine for motive power generation. 
     The motive power source  106  may provide mechanical power to a differential  108  via a transmission  110 . The power path may continue through the differential  108  to drive wheels  112 ,  114  by way of axle shafts  116 ,  118 , respectively. As such, the differential  108  distributes rotational driving force, received from transmission  110 , to the drive wheels  112 ,  114  of axle shafts  116 ,  118 , respectively during certain operating conditions. In some examples, the vehicle  102  may further include a second axle  120  that is coupled to an ICE for providing power to wheels  122 ,  124 . 
     The transmission  110  may be a hydromechanical variable transmission (HVT). Further, the transmission  110  may function as an infinitely variable transmission (IVT) where the transmission&#39;s gear ratio is controlled continuously from a negative maximum speed to a positive maximum speed with an infinite number of ratio points. In this way, the transmission can achieve a comparatively high level of adaptability and efficiency in relation to transmissions which operate in discrete ratios. Alternatively, the transmission  110  may be another type of continuously variable transmission (CVT) capable of seamlessly shifting through a continuous range of gear ratios, such as, for instance, a hydrostatic CVT using a variable displacement pump and a hydraulic motor to convert hydraulic pressure to rotation of an input shaft. In still other examples, the transmission may be a dual-clutch automatic transmission (DCT), employing two input clutches which connect a pair of input shafts to a motive power source. The DCT may be able to efficiently switch between gears by timing the operation of one clutch to engage as the other is disengaging so that there is little or substantially no interruption of torque supply to the wheels during shifting, where the capability for smooth transitions between gear ratios may enhance vehicle drivability and shift quality. Further, the differential  108  may be a locking differential, an electronically controlled limited slip differential, or a torque vectoring differential, in some examples. 
     The transmission  110  may further include a hydraulic system  126  designed to provide hydraulic fluid (e.g., natural and/or synthetic oil) to various components and assemblies of the transmission. Specifically, in one example, the hydraulic system  126  may be designed to deliver oil to a multi-disc wet clutch for actuation and/or lubrication thereof via a control valve, which will be discussed in further detail with regard to  FIGS.  2 - 7 B . In some examples, the hydraulic system may be configured to supply hydraulic fluid to multiple clutches, each in fluidic communication with a respective control valve. 
     The hydraulic system  126  may include an actuation circuit  127  for hydraulically actuating clutches within the transmission  110  and a lubrication circuit  137  for delivering lubricant to transmission components such as the clutches, bearings, and the like. The actuation circuit  127  may include a pump  128 , with a filter  130 , for driving hydraulic fluid flow through a fluid delivery line  132  to an inlet  134 . The lubrication circuit  137  may include a pump  138  for driving lubricant flow through the circuit with a filter  139  arranged upstream thereof. In some examples, the pumps  128 ,  138  may be provided at an exterior of the transmission. However, in other examples, the pumps may be included within the transmission. In some examples, the pump  128  and/or the pump  138  may receive hydraulic fluid from a common reservoir, such as, for instance, receiving a common hydraulic fluid from a sump within the transmission  110 . However, in other examples, the pumps  128 ,  138  may be in fluidic communication with different reservoirs or sources of hydraulic fluid (e.g., oil). Further, in some cases, the hydraulic system  126  may be designed to deliver hydraulic fluid, via the pump  128  and/or  138 , to other hydraulic devices and other components of the transmission  110 . 
     The vehicle  102  may further include a control system  140  with a controller  142 . The controller may include a processor  143  and a memory  144  with instructions stored therein that, when executed by the processor, cause the controller to perform various methods, control techniques, etc. described herein. The processor may include a microprocessor unit and/or other types of circuits. The memory may include known data storage mediums such as random access memory, read only memory, keep alive memory, combinations thereof, etc. 
     The controller  142  may receive various signals from sensors  145  positioned in the vehicle  102  and the transmission  110 . The sensors may include a transmission input and output speed sensor  160 ,  162 , an engine or motor speed sensor  164 , wheel speed sensors  165 , an oil temperature sensor  166 , oil pressure sensors, clutch position sensors, etc. Conversely, the controller may send control signals to various actuators  146  at different locations in the vehicle and transmission based on received signals and instructions stored in the memory  144  of the controller  142 . For instance, the controller may send a command signal to a clutch control valve in the transmission  110  and responsive to receiving the command signal, an actuator in the control valve may adjust the valve to induce clutch engagement or disengagement. Further, a controller similar to the controller  142 , illustrated in  FIG.  1   , may be used to adjust the controllable components shown in  FIGS.  2 - 3   , such as the clutch control valves. 
     An axis system  150  is provided in  FIG.  1   , as well as  FIGS.  2 - 5  and  8   , for reference. The z-axis may be a vertical axis (e.g., parallel to a gravitational axis), the x-axis may be a lateral axis (e.g., horizontal axis), and/or the y-axis may be a longitudinal axis, in one example. However, the axes may have other orientations, in other examples. 
       FIG.  2    depicts an example hydraulic system  200  including hydraulic circuits  201  and  202 . The hydraulic system  200  may be included in a transmission of a vehicle, such as transmission  110  of vehicle  102  depicted in  FIG.  1   , and may thus share similarities with hydraulic system  126 . The hydraulic circuit  201  may be designed for routing a hydraulic fluid (e.g., oil) to control various hydraulic components in a transmission, such as, for instance, hydraulically-actuated clutches, hydrostatic pumps, hydraulic motors, etc. The hydraulic circuit  202  may also be designed to route a hydraulic fluid (e.g., oil), which may the same hydraulic fluid in hydraulic circuit  201 , for lubricating various components in the transmission, such as clutches, bearings, etc., as will be described herein. Alternatively, distinct hydraulic fluids may be used in the circuits  201  and  202 , in other examples. 
     In the illustrated example, the hydraulic circuit  201  may include a control valve  208  operable to control a flow of hydraulic fluid delivered to a clutch  204 , and the hydraulic circuit  202  may include a lubrication valve  222  for allowing or restricting a flow of hydraulic fluid to the clutch  204  as well as a clutch  206 . Further, while the control valve  208  may be electronically controlled to supply a hydraulic pressure in the hydraulic circuit  201 , the lubrication valve  222  may, in some examples, not be directly electronically controlled. Rather, the lubrication valve  222  may be controlled to an open or limited flow (e.g., choked flow) state, to allow or restrict a flow of lubricant in the hydraulic circuit  202 , respectively, due to a pressure of the hydraulic fluid within hydraulic circuit  201 , as determined by the control valve  208 . To elaborate, the lubrication valve  222  may include a passive adjustment device  225  which reacts to a hydraulic pressure provided through clutch control valve  208  within hydraulic circuit  201 . In some examples, the passive adjustment device  225  may include a spool valve that resides within a chamber in fluidic communication with each of the hydraulic circuits  201 ,  202 , so as to selectively allow flow through the hydraulic circuit  202  based on a position of the device  225  determined by a pressure within hydraulic circuit  201 . In this way, the passively controlled lubrication valve may effectively and efficiently lubricate hydraulic components during certain operating conditions, as dictated by the control valve, in a simple manner which reduces system complexity. To expound, the system may avoid the use of an additional control valve for adjusting the lubrication valve, if wanted, which enables the system to achieve a passively adjusted and cost-effective lubrication valve arrangement. 
     The hydraulic clutches  204 ,  206  are specifically illustrated as multi-disc wet clutches. The clutches  204 ,  206  may be similar in structure and function, and may each be in fluidic communication with a separate control valve so as to receive hydraulic fluid therefrom. As previously introduced, the hydraulic circuit  201  may include the first clutch control valve  208  in fluidic communication with the clutch  204 , while a second clutch control valve in fluidic communication with the clutch  206  may be included in a separate hydraulic circuit. Further, in one example, the hydraulic clutch  204  may be associated with a first forward drive gear, and the hydraulic clutch  206  may be associated with a second forward drive gear or, in some cases, a first reverse drive gear. Thus, the control valve  208  may be operated to shift the first clutch  204  into an engaged state which places the transmission in a selected gear (e.g., a first forward drive gear, a second forward drive gear, or a reverse gear). 
     The multi-disc wet clutch  204  may include an actuator with a hydraulic chamber  210  and a piston  212  that is axially slidable in the chamber. The clutch may further include a clutch pack having interleaved separator and friction discs  214  that are designed to selectively engage one another, and the piston may be biased (e.g., by a spring  216 ) into a disengagement position (away from the clutch pack) where the plates are disengaged. Pressurized hydraulic fluid is supplied to the hydraulic chamber  210  via a clutch line  218  (fluidly connected to a port of the actuation cylinder), such that hydraulic pressure in the chamber causes the piston  212  to move into an engagement position (towards the clutch pack) to urge the separator and friction plates to engage to allow torque transmission through the clutch. It will be understood that clutch  206 , being substantially identical to clutch  204 , may include similar components and operate in a similar manner. In other examples, however, other types of hydraulically operated clutches may be used, such as hydraulically operated dog clutches. 
     The control valve  208  may be an electronically actuated valve designed to control the pressure and/or flow of hydraulic fluid in the clutch line  218  that extends between the multi-disc wet clutch  204  and the valve. The hydraulic system  200  further includes a pump  220  that supplies hydraulic fluid to the control valve  208 . The pump  220  may be a fixed displacement pump, such as a positive displacement hydraulic pump having a housing and a pumping element (e.g., a reciprocating piston or rotary element) designed to deliver a pressurized hydraulic fluid during each cycle, in some examples. The control valve  208  adjusts the pressure of the hydraulic fluid in the line  218  to engage and disengage the multi-disc wet clutch  204 . For instance, the control valve may increase the pressure in the line  218  to induce clutch engagement during a gear shift event. Conversely, the control valve may decrease the pressure in the line to disengage the clutch when the transmission is shifted into another gear. In some cases, the pump  220  may receive the hydraulic fluid from a reservoir  221  (e.g., via a fluid pick-up line), and the control valve  208  may return hydraulic fluid to the reservoir  221  via a fluid return line  219 . 
     In one example, the actuation of the clutch  204  may include a prefilling phase where the control valve  208  delivers a prefill pressure to the hydraulic piston  212 . In the prefill phase of the clutch, the piston is readied for clutch engagement by providing the piston cylinder with a hydraulic pressure that is lower than a spring preload pressure, such that the piston is pressurized but remains at a rest position. The prefilling step enables (e.g., ensures that) the clutch line to be filled with oil (e.g., without entrainment of air bubbles therein) to facilitate reliable and repeatable behavior with regard to both the clutch line and the control valve in preparation for clutch engagement. Subsequent to the prefill phase, when clutch engagement is anticipated, the clutch may enter a filling phase in which the hydraulic pressure supplied to the piston may be further increased to induce clutch engagement. The increased hydraulic pressure moves the piston from the prefill position (e.g., rest position) to or near a kiss point position (i.e., the point at which the clutch begins to transmit torque). The filling phase may be performed at discrete times in anticipation of a shift event, and may decrease shift delays. After the clutch reaches the kiss point, when clutch engagement is demanded, the clutch enters an engagement phase where additional hydraulic pressure applied to the piston induces engagement of the clutch plates. In some cases, controller logic may trigger the transition from the prefill phase to the filling phase and from the filling phase to the engagement phase and vice versa. 
     The hydraulic circuit  201  may be in hydraulic communication with the lubrication valve  222  via a pilot line  224 . The pilot line  224  may branch off from the clutch delivery line  218 , at junction  209 . In this way, the pressurized hydraulic fluid provided by clutch control valve  208 , in addition to being routed to the clutch  204 , is provided to lubrication valve  222  to adjust a position thereof. More specifically, in some examples, the lubrication valve  222  may be biased (e.g., by a spring) into a choked position, realizing a limited flow state where the lubrication valve is substantially or nearly closed. Further, upon application of hydraulic pressure through pilot line  224  (e.g., when the control valve permits a pressurized flow of hydraulic fluid through clutch delivery line  218  to actuate the clutch  204 ), the lubrication valve may be shifted into an open position, allowing a flow of hydraulic fluid through the lubrication valve into a lubrication line  226 . Thus, the lubrication valve  222  may not be directly electronically controlled, but instead be passively controlled using the pressure of the hydraulic fluid provided to the clutch  204  via the control valve  208 . In this way, by using a combined control scheme to route oil both for actuation and lubrication of the clutches, several complexities and costs associated with the design and control of the hydraulic system  200  may be reduced. 
     In some examples, a pressure gauge  223  may be provided in pilot line  224  to measure a hydraulic pressure being applied to the lubrication valve  222  to control a position thereof. In some cases, the lubrication valve may be designed as an on/off type valve. However, in other cases, the lubrication valve may be of the continuously adjusting type, transitioning between multiple opened and closed states. In any case, it will be understood that a clutch actuation pressure within the clutch line  218  affects a state of the lubrication valve  222 , selectively permitting or restricting the flow of lubricant to the clutches. Consequently, it will be understood that the clutch actuation pressure (provided to the lubrication valve  222 ) may be adjusted in accordance with a desired state of the lubrication valve  222  (e.g., the open flow state or the limited flow state) for adjusting the flow of hydraulic fluid through the lubrication valve. 
     In order to provide hydraulic fluid to the lubrication valve  222  for lubricating the clutches  204 ,  206 , the hydraulic circuit  202  may further include a pump  232  in fluidic communication with the lubrication valve via a lubrication conduit  234 . In some examples, the pump  232  may receive hydraulic fluid (e.g. oil) from a reservoir, which may be the reservoir  221 . In this way, by providing a common reservoir and hydraulic fluid to be sent to the control valve  208  and the lubrication valve  222  via the pumps  220 ,  232 , respectively, the hydraulic system  200  may be realized in a compact package with reduced size and complexity in a transmission. However, in other examples, the pump  232  may receive hydraulic fluid from a separate reservoir. Further, in some examples, the hydraulic fluid delivered for lubricating clutches  204 ,  206  may be returned to a reservoir for subsequent use in the hydraulic system  200 . 
     When the lubrication valve  222  is in an open state (e.g., in anticipation of and/or during a gear shift event), hydraulic fluid may be routed from the pump  232  to the lubrication valve  222  and into the lubrication line  226 . The lubrication line  226  may then direct the hydraulic fluid to the clutches  204 ,  206  via lubrication conduits  228 ,  230  for lubrication thereof, respectively. Hydraulic fluid may flow freely in this manner when the lubrication valve  222  is in the open flow state, under certain operating conditions. However, in other operating conditions, when the lubrication valve  222  is in the limited (e.g., choked) flow state, flow through the lubrication valve and thus the lubrication line  226  is restricted. Specific control strategies for transitioning the lubrication valve  222  between the open and limited flow states under certain conditions will be expanded upon herein, particularly with regard to  FIGS.  6 - 7 B . 
     Another example of a hydraulic system  300  including a lubrication valve  310  is depicted in  FIG.  3   , incorporated in a portion of a transmission. The hydraulic system  300  may be included in or similar to the hydraulic systems  126 ,  200  depicted in  FIGS.  1  and  2   , respectively. Specifically, the lubrication valve  310  may be similar to the lubrication valve  222  depicted in  FIG.  2   , and may thus be adjusted between an open and limited flow (e.g., choked flow) state using clutch actuation pressure as dictated by a clutch control valve. 
     The hydraulic system  300  may be disposed within in a rotating shaft  302  (e.g., driving shaft) having a rotational axis  303 . Therefore, the shaft  302  may drive rotation of sets of discs in the clutch units  306 ,  308 . Specifically, in one example, the rotating shaft  302  may be selectively rotationally coupled to a driven shaft  304  via clutch units  306 ,  308 . The driven shaft may be rotationally coupled to gears that deliver power to the transmission&#39;s output. Each of the clutch units  306 ,  308  may be multi-disc wet clutches having clutch packs  307 ,  309 , respectively, and other components configured to operate in the manner previously described with regard to  FIG.  2   . 
     In some cases, roller bearings  314 ,  316  (e.g., spherical and/or cylindrical roller bearings) may be included on the rotating shaft  302  for supporting the shaft  302  in the transmission in relation to other transmission components. Additionally, multiple bearings  318  (e.g., spherical ball bearings) may be included between the rotating shaft  302  and the driven shaft  304 . These bearings permit the driving shaft  302  and the driven shaft  304  to rotate independently when the clutch unit  306  is disengaged. 
     The rotating shaft  302  may include a lubrication line  320 , for delivering lubricant to the clutch packs  307 ,  309  in the clutch units  306 ,  308 , respectively and bearings  318  via lubricant delivery lines  322 . The lubrication line  320  may receive hydraulic fluid from a lubricant conduit  323  via the lubrication valve  310 . Further, in some examples, the lubricant conduit  323  may receive hydraulic fluid from an inlet  321 . In some cases, the inlet  321  may be located at an axial end of the rotating shaft, although other configurations have been envisioned, such as a radially extending inlet channel opening to an outer diameter of the rotating shaft. The inlet  321  may be in fluidic communication with a pump and reservoir, in any manner previously described herein. 
     The lubrication line  320  and the lubricant conduit  323  may share a common longitudinal axis  325 , which may be offset from and parallel to the central rotational axis  303  of the rotating shaft  302 . Further, in some examples, the delivery lines  322  may be oriented normal to the common axis  325 , radially extending from the lubrication line  320  to an outer surface of the rotating shaft  302 . The delivery lines  322  may be radially aligned and are shown opening into sections of the transmission between two of the bearings  318 , in one example. In this way, the lubricant may be efficiently routed through the transmission. However, other line contours have been contemplated. For instance, the delivery lines  322  may extend from the lubrication line  320  to the outer surface of the rotating shaft at a different angle (e.g., 45 degrees, 30 degrees, 15 degrees, etc.). 
     The rotating shaft may further include a clutch actuation line  330  for delivering hydraulic fluid to an actuation chamber of the clutch unit  308 . The actuation line may receive hydraulic fluid at an inlet  331 . Further, the inlet  331  may be in fluidic communication with a control valve which receives pressurized hydraulic fluid from a reservoir via a pump, in any manner previously described herein. In some cases, the inlet  331  may be located on an outer diameter of the rotating shaft  302  proximate one end. However, other configurations have been contemplated, such as an inlet opening at an axial end of the shaft (e.g., similar to the inlet  321 ). Further, in the illustrated embodiment, the inlet  321  of the lubricant conduit  323  and the inlet  331  of the clutch actuation line  330  are positioned at or near opposing ends of the shaft  302 . However, in other examples, the inlets may have different relative arrangements, such as, for instance, where both of the inlets  321 ,  331  may be positioned at or near the same end of the rotating shaft. Further, the clutch actuation line  330  may have a longitudinal axis  335 , which, similar to the axis  325  of the lubrication lines, may be parallel to and offset from the central rotational axis  303  of the rotating shaft  302 . 
     The lubrication valve  310  includes a passive adjustment device  329  that is designed to switch the state of the valve using the pressure of the fluid in the clutch actuation line  330 . To carry out this device functionality, the passive adjustment device  329  may include a control conduit  312  defining a chamber of the lubrication valve  310  and a spring  315  that may be coupled to a piston in the valve. The control conduit  312  may extend (e.g., radially extend) from the clutch actuation line  330 . As such, a first end (e.g., upper end) of the control conduit  312  is in hydraulic communication with the clutch actuation line  330 , and a second end (e.g., lower end) of the control conduit  312  is in hydraulic communication with the passive adjustment device  329 . 
     Detailed views of the lubrication valve  310  in a limited flow state (e.g., choked flow state) and an open flow state are shown in  FIGS.  4  and  5   , respectively. It should be noted that the lubrication valve is also shown in the limited flow state in  FIG.  3   . The lubrication valve  310  may be a spool-type valve, such as a piston spool valve, in one example. The piston of the valve  310  may be slidably movable within the control conduit  312  along an axis thereof to control a hydraulic flow from the lubrication line  321  to the lubrication line  320 . In some cases, the piston of the valve  310  may be biased by a spring  315 , so as to be positioned in the limited flow state  400  in  FIGS.  3 - 4   , whereby a flow of oil into the lubrication line  320  is restricted. Further the piston of the valve  310  may include a seal proximate a central portion thereof, so as to seal off the conduit  312  to maintain a separate hydraulic flow in the clutch actuation line  330  independent from a hydraulic flow between the lubricant conduit  323  and line  320 . 
     In the hydraulic system configuration depicted in  FIG.  4   , the hydraulic pressure in the clutch line  330  does not induce actuation of the passive adjustment device  329 . As such, the spring  315  in the passive adjustment device  329  is decompressed and the device is restricting flow through the valve  310  and between the lubricant conduit  323  and to the lubrication line  320 . However, alternate arrangements of the passive adjustment device have been contemplated. 
       FIG.  5    depicts the lubrication valve  310  in the open flow state  500 . To elaborate, in the open flow state, the hydraulic pressure in the clutch actuation line is greater than the threshold pressure that moves the clutch into the open flow state. In order to provide a hydraulic pressure to the clutch actuation line  330  and therefor the control conduit  312 , the clutch control valve is operated (e.g., electronically operated) to supply a pressurized flow of hydraulic fluid through the inlet  331 , shown in  FIG.  3   , into the clutch actuation line  330 . The clutch actuation line  330  also routes the pressurized flow to the actuation chamber of the clutch unit  308  via an actuation delivery line  332 , shown in  FIG.  3   . The line  332  may be routed through the driving shaft  302  to provide fluid to the clutch&#39;s control piston. 
     Returning to  FIG.  5   , the hydraulic pressure in the clutch actuation line is communicated to the control conduit  312  to exert a force on the piston of the lubrication valve  310 . When the hydraulic pressure surpasses a threshold pressure (e.g., overcoming the biasing force of the spring  315  in the passive adjustment device  329 ), the piston of the lubrication valve  310  will be moved (e.g., downwardly) into an open position where the flow  502  of hydraulic fluid from the lubricant conduit  323  into the lubrication line  320 . Thus, the hydraulic fluid in lubrication line  320  may be delivered to clutch components via lubrication delivery lines  322 , particularly in anticipation of or during actuation of the clutch the clutch  308  is being actuated and lubrication demands of the system may be higher. 
     Oil may be provided at a high (e.g., maximum) flow rate to lubricate the clutches  306 ,  308  when one of the clutch is engaged or shift event is anticipated. However, supplying a high flow rate of oil to a disengaged clutch decreases the efficiency of the transmission. Thus, it may be desired to limit the flow of hydraulic fluid in the lubrication line  320  when the clutch  308  is disengaged and a shifting event is not anticipated. The spring-biased lubrication valve  310  may therefore be placed in the limited flow state  400 , as depicted in  FIG.  4   , to restrict the flow of hydraulic fluid in the lubrication line  320  under certain operating conditions (e.g., when a gear shift event is not anticipated), thereby reducing drag torque experienced at the clutch and increase transmission efficiency. 
     During transmission operation, the rotating shaft  302  may be rotated while the hydraulic pressure supplied to the clutch actuation line  330  (and the resulting adjustment of the lubrication valve  310 ) is adjusted. In some system configurations, centrifugal forces result from the rotation, due to misalignment of a center of gravity of the lubrication valve  310  and the central rotational axis  303  of the shaft  302 , which may have some effect on the adjustability of the lubrication valve. However, the spring  315  may be constructed with a spring constant that allows the spring to exert a biasing force on the piston in the valve that is greater than the expected centrifugal forces which may urge the valve&#39;s piston into the open state from the limited flow state. In this way, the lubrication valve is inhibited from inadvertently opening due to the centrifugal forces generated via shaft rotation. In this way, the simple design of the lubrication valve may provide a reliable, inexpensive solution for passively controlling the flow of lubrication to clutch components during certain operating conditions which takes into account drive shaft rotation. 
     In some examples, with reference to  FIGS.  3 - 5   , the clutch  308  may be a first forward drive clutch, such that actuating the clutch  308  to the engaged position places the transmission in a first forward drive gear. Further, the clutch  306  may be a reverse drive clutch or a second forward drive clutch, such that engagement of the clutch  306  may place the transmission in a reverse drive gear or a second forward drive gear, respectively. When the transmission is operated to shift into the first forward drive gear, the hydraulic pressure in the clutch actuation line may be increased (e.g., via actuation of a control valve) so as to actuate the clutch unit  308  and correspondingly adjust the lubrication valve  310  from the choked position  400  to the open position  500  to allow a higher flow of oil to lubricate each of the clutches  306 ,  308  and associated bearings. Thus, control techniques for the hydraulic system may be dependent on an operating gear ratio of the transmission, as well as other operating conditions, as will be described herein with reference to  FIGS.  6 - 7 B . 
     A control strategy or method  600  for determining a desired lubrication valve state and operating a clutch control valve to achieve the desired state is shown in  FIG.  6   . The method  600  may be carried out by any of the transmission and hydraulic systems system in  FIGS.  1 - 5  and/or  8    or combinations of the transmissions and hydraulic systems depicted therein. However, in other examples, the method  600  may be implemented in other suitable transmissions and hydraulic systems. Instructions for carrying out the method  600  and other control strategies or routines described herein may be stored on a memory of a controller and executed by a processor of the controller in conjunction with signals received from sensors at the controller. The controller may employ actuators in different hydraulic system components to implement the method steps described below. 
     At  602 , the method determines a transmission ratio (e.g., the mechanical transmission ratio representing the input to output speed ratio of the transmission). In one example, determining the transmission ratio may include measuring shaft speeds (e.g., the speed of the transmission&#39;s input and output shaft, using shaft speed sensors, for instance) and subsequently calculating the transmission&#39;s ratio from the sensor inputs. However, in other examples, the transmission ratio may be modelled. In one example, the determined gear ratio may indicate that the transmission is in a first forward drive gear (e.g., a first positive, non-zero ratio), a second forward drive gear (e.g., a second positive, non-zero ratio, greater than the first ratio), or a first reverse drive gear (e.g., a negative, non-zero ratio). Further, in some cases, in each of the three gear ratios the clutch associated with the operating gear ratio is engaged while the other clutches are disengaged. 
     At  604 , the method judges if the commanded state of the lubrication valve should be an open flow state or a limited (e.g., choked) flow state based on the transmission ratio and/or other operating conditions. For instance, when the transmission ratio approaches a gear ratio shift point (e.g., transitioning into the first forward drive gear from the second forward drive gear or the first reverse drive gear), it may be desired to provide a greater flow of hydraulic fluid through the lubrication valve to lubricate clutch components to enable a smooth transition into the relevant gear. Under these conditions it may be determined that the lubrication valve should be transitioned into the open flow state. Conversely, when the transmission ratio is far enough away from a gear shift point (e.g., transitioning into the first forward drive gear is not anticipated), a desire for greater efficiency may be desired, since the clutch is not engaged (or near actuation to cause engagement thereof). Under these conditions, it may be determined that the lubrication valve should be transitioned into the limited flow state. In some cases, determining the desired valve state may further take into account inputs from other sensors in the hydraulic system and/or vehicle, such as an oil temperature measured by an oil temperature sensor. 
     Therefore, when it is determined that the transmission ratio is approaching (or within) a range associated with the first forward drive gear, the method transitions the lubrication valve into the open flow state, so as to provide a high pressure of oil for lubrication of clutch components, in order to enhance shift quality during the shift event. Conversely, when the gear ratio is determined to be far enough away from the first forward drive gear range, lubrication demands in the hydraulic system may lower, the method transitions the lubrication valve into the limited flow state. As such, in the limited flow state, the hydraulic flow to the clutch discs is restricted, thereby reducing drag torque and increasing transmission efficiency. 
     Next, at  606 , the method determines a speed of the rotating shaft in which the lubricant valve resides and a desired lubricant pressure. In other examples, however, the step  606  may determine other operating conditions, as measured by various sensors. 
     At  608 , the method calculates a desired clutch pressure, which may be a hydraulic pressure to be supplied via a clutch actuation line to engage or disengage a clutch. In one example, the desired clutch pressure may be zero or near-zero, when the clutch is to be disengaged (e.g., when the transmission ratio is outside of the first forward drive gear range), or a high pressure value, when the clutch is to be fully engaged (e.g., in the first forward drive gear). In some examples, the desired clutch pressure may be a value between the near-zero value and the high pressure value, such as during a prefilling step of a clutch. 
     At  610 , the method adjusts the clutch control valve to achieve the desired clutch actuation pressure to transition or maintain the lubrication valve in the desired state. As previously discussed, the desired lubrication valve state may be realized by adjusting the hydraulic pressure in the clutch actuation line via control valve adjustment, and detailed discussion will be omitted for brevity. 
     By accounting for the clutch lubricant demands using the transmission ratio, the method provides a simple and effective way to supply clutch lubrication when the clutch is engaged and decreases the clutch lubrication when the clutch is disengaged to decrease drag torque and increase transmission efficiency. 
       FIG.  7 A  depicts a graphical representation of a prophetic use-case control diagram  700  with transmission gear ratios, and  FIG.  7 B  depicts a prophetic use-case timing diagram  750 . In some cases, control diagram  700  and/or timing diagram  750  may be implemented in accordance with the method  600  depicted in  FIG.  6   , and may thus be implemented in the hydraulic systems described above, with reference to  FIGS.  1 - 5   , or another suitable hydraulic system. 
     Turning to  FIG.  7 A , in each chart and plot of the diagram  700 , the transmission gear ratio is indicated on the abscissa and increases in the direction of the arrow. Chart  702  graphically represents different transmission operating modes corresponding to the gear ratio, which, in the illustrated example, include a first reverse drive gear  704  (e.g., a negative, non-zero ratio), a first forward drive gear  706  (e.g., a first positive, non-zero ratio), and a second forward drive gear  708  (e.g., a second positive, non-zero ratio, greater than the first ratio). The ordinate of plot  710  indicates a lubrication valve state (e.g., open or choked), and the ordinate for plot  712  indicates a hydraulic pressure applied to the first forward drive clutch&#39;s actuator (e.g., control piston). 
     When the transmission is operating in the reverse drive gear  704  or the second forward drive gear  708  far enough away from a shift point, the clutch pressure supplied to the clutch associated with the first forward drive gear  706  is held at a lower value (e.g., at or near zero), such that the lubrication valve remains in a choked state. 
     When the gear ratio approaches a gear shift point between the reverse gear and the first gear, the clutch pressure is increased above a threshold pressure  713  at GR 1 . The threshold pressure  713  is the clutch line pressure at which the lubrication valve transitions from the choked flow state to the open flow state or vice versa. However, the clutch pressure may be below a pressure that triggers clutch engagement (e.g., full clutch engagement). The clutch pressure may be increase above the threshold pressure when a variance between the current gear ratio and a shift point ratio (GR 0 ), representing the shift point between the reverse gear and the first forward drive gear, is less than a threshold ratio variance: GR 0 −GR 1 , in the illustrated use-case. As depicted, the shift point ratio (GR 0 ) is zero. 
     Next the clutch pressure is further increased to cause full engagement of the first forward drive clutch and allow the transmission to shift from the reverse drive gear  704  into the first forward drive gear  706 . Therefore, at GR 0  the lubrication valve transitions to the open state. 
     As the gear ratio increases, the transmission shifts from the first forward drive gear  706  to the second forward drive gear  708 , at GR 2 . However, at GR 2 , the clutch pressure is reduced but remains above the threshold pressure  713 , so that the lubrication valve remains in the open state. In this way, the chance of clutch degradation caused by insufficient lubrication is reduced. However, when the difference between the operating gear ratio and the shift point ratio (GR 2 ), representing the shift point between the first forward gear and the second forward gear, surpasses the threshold ratio variance (GR 3 −GR 2  in the illustrated use-case), the clutch pressure is further reduced to a pressure below the threshold pressure  713 , at GR 3 . As a result, the lubrication valve transitions from the open state to the choked state while the transmission remains in the second forward drive gear. Further, it will be understood that the same sequence of clutch pressure adjustment may occur when the transmission shifts from the second forward drive gear  708  to the first forward drive gear  706 , and so on. In this way, controlling the clutch pressure to effect an open state of the lubrication valve may allow the hydraulic system to efficiently lubricate clutch components when there is a high demand for lubricant (e.g., before and/or after a gear shift event, and during clutch actuation), while avoiding the use of an electronic actuator directly controlling the lubrication valve. 
     Turning to  FIG.  7 B , in each plot of the timing diagram  750 , time is indicated on the abscissa and increases in the direction of the arrow. The ordinate of plot  752  indicates a clutch state (engaged or disengaged), of the first forward drive gear clutch. The ordinate of plot  754  indicates a hydraulic pressure (e.g., supplied to the clutch via a control valve in a hydraulic system of the transmission), and the ordinate of plot  756  indicates a lubrication valve state (open or choked). 
     From t 0  to t 2 , the first gear clutch is disengaged. However, at t 1 , when the transmission approaches a gear shift event (into the first gear from the reverse gear), the hydraulic pressure is increased above a threshold pressure  755 , and the lubrication valve is transitioned from the choked flow state to the open flow state. Consequently, hydraulic fluid may flow through the lubrication valve to lubricate the first gear clutch at a time before the clutch is fully engaged, allowing for a smooth transition into the first gear. In this way, the system may provide a greater flow of lubricant when a demand for lubrication at the clutch is higher (e.g., before and during clutch actuation). While efficiency losses may arise when the lubrication valve is in the open flow state and the clutch is disengaged, these losses may be decreased by employing the open flow state in this manner. As such, the duration of the open flow state may be constrained to specific times when clutch lubrication is desired for shifting operations and during clutch engagement. 
     At t 2 , the hydraulic pressure is further increased to a higher pressure that causes full engagement of the first gear clutch, and the lubrication valve remains in the open flow state. Further, from t 2  to t 3 , the hydraulic pressure remains at the higher value, the first gear clutch remains engaged, and the lubrication valve remains in the open state. 
     At t 3 , the hydraulic pressure is decreased and the first gear clutch is disengaged. From t 3  to t 4 , the hydraulic pressure is sustained above the threshold pressure  755  while the transmission remains in a second gear, after the transmission shift event at t 3 , and the lubrication valve remains in the open flow state. 
     At t 4 , the hydraulic pressure is further reduced to a value lower than the threshold pressure  755 . Consequently, the lubrication valve is transitioned from the open flow state to the choked flow state, and the first gear clutch remains disengaged. In this way, by maintaining the lubrication valve in the open flow state for some time after a shift event (from t 3  to t 4 ), the system can avoid an abrupt drop in lubrication immediately upon shifting out of the first gear, which may decrease the likelihood of clutch component degradation and increase system longevity. 
       FIG.  8    shows a schematic depiction of a transmission  800  (e.g., an HVT) in a vehicle  802 . The vehicle  802  and the transmission  800  are examples of the vehicle  102  and transmission  110  depicted in  FIG.  1   . Therefore, the transmission  800  may include structural and/or functional features of the transmission  110 , or vice versa. 
     The transmission  800  may function as an IVT where the transmission&#39;s gear ratio is controlled continuously from a negative maximum speed to a positive maximum speed with an infinite number of ratio points. In this way, the transmission can achieve a comparatively high level of adaptability and efficiency when compared to transmission which operate in discrete ratios. 
     The transmission  800  may have asymmetric maximum output speeds for forward and reverse direction. This forward-reverse speed asymmetry may enable the transmission to achieve a desired breadth of speed ranges. However, other suitable output speed variations have been contemplated, such as symmetric output speeds in the forward and reverse directions, which may however, demand the use of an additional clutch which may increase system complexity. 
     The transmission  800  may include or receive power from a motive power source  804 . The power source  804  may include an internal combustion engine, electric motor (e.g., electric motor-generator), combinations thereof, and the like. 
     A torsional damper coupling  806  may be further provided in the transmission. Gears  808 ,  810 , such as bevel gears, may be used to rotationally couple the power source  804  to an input shaft  812 . As described herein, a gear may be a mechanical component which rotates and includes teeth that are profiled to mesh with teeth in one or more corresponding gears to form a mechanical connection that allows rotational energy transfer therethrough. 
     A mechanical PTO  814  may be coupled to the input shaft  812 . The mechanical PTO  814  may drive an auxiliary system such as a pump (e.g., a hydraulic pump, a pneumatic pump, and the like), a winch, a boom, a bed raising assembly, etc. To accomplish the power transfer to auxiliary components, the PTO may include an interface, shaft(s), housing, and the like. However, in other examples, the PTO and/or the disconnect clutch may be omitted from the transmission. A gear  816  may be coupled to the input shaft  812 . A mechanical assembly  818  is further included in the transmission  800 . The mechanical assembly  818  may include the shaft  812  and/or the gear  816  as well as shaft  867 , described in greater detail herein. Further, the transmission may include a shaft  820  and a gear  822  rotationally coupled to the gear  816  on the input shaft  812 . Dashed line  824  and the other dashes lines depicted in  FIG.  8    indicate a mechanical connection between components which facilitates rotational energy transfer therebetween. 
     A gear  826  meshing with gear  822  may be rotationally attached to a charging pump  828 . The charging pump  828  may be designed to deliver pressurized fluid to hydraulic components in the transmission such as a hydraulic motor  834  (e.g., hydrostatic motor), a hydraulic pump  836  (e.g., hydrostatic pump), and the like. The fluid pressurized by the charging pump may additionally be used for clutch actuation and/or transmission lubrication. The charging pump may include a piston, a rotor, a housing, chamber(s), and the like to allow the pump to move fluid. The mechanical assembly  818  is rotationally coupled in parallel to a hydrostatic assembly  830 . Further, the hydrostatic assembly  830  may have a U-shape design where the mechanical interface shafts  831 ,  833  for a hydraulic pump  836  (e.g., variable displacement pump) and a hydraulic motor  834  (e.g., fixed bent axis motor), respectively, are parallel to one another and arranged on one side of the assembly. This U-shaped layout permits the hydrostatic assembly&#39;s size to be reduced and enables the use of high pressure hoses to be forgone to reduce manufacturing costs as well the chance of hydrostatic unit degradation, if desired. Still further, the hydrostatic assembly  830  may be arranged on an opposite side of the transmission as the charging pump  828  and/or axially offset from clutches  870 ,  872 . Arranging the hydrostatic assembly in this manner permits the width and length of the transmission to be reduced and allows the installation of the transmission in the vehicle to be simplified. 
     The coupling of the hydrostatic assembly to the mechanical assembly enables the transmission to achieve power split functionality in which power may synchronously flow through either path to additively combine or recirculate power through the system. This power split arrangement enables the transmission&#39;s power flow to be highly adaptable to increase efficiency over a wide range of operating conditions. Thus, the transmission may be a full power split transmission, in one example. 
     The mechanical assembly  818  may include multiple mechanical paths that are coupled in parallel to the hydrostatic assembly. To elaborate, the shaft  867  may serve as a junction for a first mechanical path (e.g., branch)  819  and a second mechanical path (e.g., branch)  821 . The first mechanical path  819  may provide rotational energy transfer capabilities from an interface of the hydrostatic assembly  830  to a ring gear  858  of a first planetary gear set  848 , during certain operating conditions. Additionally, the second mechanical path  821  may provide rotational energy transfer capabilities from the interface of the hydrostatic assembly  830  to a carrier  860  of a second planetary gear set  850 . 
     The hydrostatic assembly  830  includes the hydraulic motor  834  and the hydraulic pump  836 . Further, the hydraulic pump  836  may include a first mechanical interface  838  and a second mechanical interface  840 . The first mechanical interface  838  may be rotationally coupled to a mechanical bushing  832  and the second mechanical interface  840  may be rotationally coupled to another mechanical PTO  842 . Again, the mechanical PTO may be used to drive an auxiliary vehicle system such as an air compressor, a mechanical arm or boom, an auger, etc. In this way, the transmission may be adapted for a variety of end-use operating environments. Providing multiple PTOs, in the arrangement depicted in  FIG.  8   , enables the transmission system to meet end-use design goals in a variety of different types of vehicles, if wanted. As such, the system&#39;s applicability is expanded and the customer appeal of the transmission is increased. However, in other examples, the PTOs  814  and/or  842  may be omitted from the transmission. 
     The hydraulic pump  836  may be a variable displacement bi-directional pump, in one example. Further, the pump may be an axial piston pump, in one instance. To elaborate, the axial piston pump may include a swash plate that interacts with pistons and cylinders to alter the pump&#39;s displacement via a change in swivel angle, in one specific example. However, other suitable types of variable displacement bi-directional pumps have been contemplated. 
     The hydraulic motor  834  may be a fixed displacement bi-directional motor (e.g., fixed bent axis motor). The fixed bent axis motor is relatively compact when compared to variable displacement motors. The system can therefore achieve greater space efficiency and pose less space constraints on other systems in the vehicle, if desired. However, alternate types of pumps and/or motors may be used, if motor adjustability is favored at the expense of compactness, for instance. 
     Hydraulic lines  844 ,  846  are attached to hydraulic interfaces in each of the motor and pump to enable the hydrostatic assembly to provide additive and power circulation functionality with regard to the mechanical branches arranged in parallel with the hydrostatic assembly  830 . For example, in an additive power mode, power from both the hydrostatic and mechanical assemblies is combined at one of the planetary gear sets and delivered to the transmission output. In a recirculating power mode, power is recirculated through the hydrostatic assembly. Therefore, the hydraulic pump  836  and the motor  834  may be operated to flow power to the sun gears of either planetary assembly from the hydraulic motor. Conversely, the pump and the motor may be operated to flow power back to the gear set and the mechanical branches. 
     The transmission  800  further includes the first planetary gear set  848  and the second planetary gear set  850 . The first planetary gear set  848  may include a carrier  852  on which planet gears  854  rotate. The planet gears  854  may mesh with a sun gear  856  and the ring gear  858 . Likewise, the second planetary gear set  850  may include the carrier  860 , planet gears  862 , a sun gear  864 , and a ring gear  866 . Therefore, the second planetary gear set  850  may again be a simple planetary gear set. Further, bearings arranged between the planet gears and the carrier in each planetary arrangement may facilitate rotation thereof. The sun gears and/or shafts to which they are attached may further have bearings coupled thereto. The bearings may be roller bearings (e.g., needle roller bearings), ball bearings, or other suitable types of bearings that enable component rotation while constraining other relative motion. 
     The carrier  860  of the second planetary gear set  850  may be rotationally coupled to the ring gear  858  of the first planetary gear set  848 . Further, the carrier  860  of the second planetary gear set  850  may be rotationally coupled to a shaft  867 . The shaft  867  may extend through a central opening in an extension  886 , described in greater detail herein. This rotational attachment scheme may be conceptually described as a formation of mechanical branches attached in parallel to the hydrostatic assembly  830 . 
     As described herein a parallel attachment between components, assemblies, etc., denotes that the input and output of the two components or grouping of components are rotationally coupled to one another. This parallel arrangement allows power to recirculate through the hydrostatic assembly, during some conditions, or be additively combined from the mechanical and hydrostatic branches, during other conditions. As a result, the transmission&#39;s adaptability is increased, which allows gains in operating efficiency to be realized, when compared to purely hydrostatic transmissions. 
     The sun gears  856 ,  864  of the first and second planetary gear sets  848 ,  850  may be rotationally coupled (e.g., directly attached) to one another. Attaching the sun gears in this manner may enable the transmission to achieve a desired gear ratio, compactness, and efficiency. 
     The hydraulic motor  834  may be rotationally coupled to the sun gear  856  via a mechanical bushing  868 , for instance. The transmission  800  further includes a reverse clutch  870 , a first forward drive clutch  872 , and a second forward drive clutch  874 . The clutches  870 ,  872 ,  874  may be positioned near to an output shaft  871  and downstream of the planetary assembly. Arranging the clutches in this location allows a targeted compromise between clutch size and clutch speed. For instance, relatively high clutch speeds may generate higher power losses. Further, the reverse clutch  870  and the first forward drive clutch  872  may be arranged adjacent and coaxial to one another. In one particular example, the clutches may have a similar design to reduce manufacturing complexity. This twin clutch arrangement therefore permits manufacturing costs to be reduced and increases the transmission&#39;s compactness. 
     The clutches  870 ,  872 ,  874  may be friction clutches that each includes two sets of plates. The clutch plates may rotate about a common axis and are designed to engage and disengage one another to facilitate selective power transfer to downstream components. In this way, the clutches may be closed and opened to place them in engaged and disengaged states. In the disengaged state, power does not pass through the clutch. Conversely in the engaged state, power travels through the clutch during transmission operation. The carrier  852  may include an extension  875  with a gear  876  that meshes with a gear  877 . The gear  877 , in the illustrated example, is rotationally coupled to the reverse clutch  870  and the first forward clutch  872 . The reverse clutch  870  and the first forward clutch  872  are shown arranged adjacent to one another and may share a common rotational axis. Because of this proximal clutch arrangement, the system may exhibit greater compactness which poses less space constraints on adjoining vehicle systems. Alternatively, the reverse clutch may be spaced away from the first forward clutch which may, however, decrease system compactness. 
     A gear  879  may reside on an output shaft  880  of the reverse clutch  870 . Likewise, a gear  881  may reside on an output shaft  882  of the first forward clutch  872 . Both gears  879 ,  881  may be rotationally attached to the system output shaft  871  via gears  883 ,  884  respectively. In this way, both the reverse clutch and the first forward clutch deliver power to the transmission&#39;s output, during different operating conditions. 
     The system output shaft  871  may include one or more interfaces  885  (e.g., yokes, gears, chains, combinations thereof, etc.). The output shaft is specifically illustrated with two outputs. However, the transmission may include an alternate numbers of outputs. The gear  879  is rotationally coupled to the output shaft via meshing with gear  883 . Arrows  891  depict the flow of power from the transmission system to drive axles  892  and/or other suitable downstream vehicle components or vice versa. A driveline with a shaft, joints, etc. may be used to carry out the power transfer between the transmission and the axles. It will be understood that the drive axles may include drive wheels.  FIG.  8    further depicts rotational speed sensors  899  that measure the speed of a gear and shaft on which the gear resides. 
     The ring gear  866  of the second planetary gear set  850  may include the extension  886  with a gear  887  position thereon. The gear  887  may be rotationally attached to a gear  888  in the second forward clutch  874 , as indicated via a dashed line. The gear  888  may be coupled to a first set of plates in the clutch  874 . A second set of plates in the clutch may be attached to an output shaft  889  and a gear  890 . The gear  890  may be rotationally coupled to the gear  883 , as indicated by a dashed line. Due to the aforementioned arrangement of the clutches and the planetary gear sets, the transmission  800  achieves a higher efficiency and enhanced drivability, comfort, and productivity than previous hydromechanical transmissions. 
     The technical effect of the control methods for the transmission hydraulic systems described herein is to efficiently and passively control a lubrication valve, based on a gear ratio in a transmission and an efficiency standpoint, and operating a control valve accordingly in order to provide lubricant to a wet friction clutch when demanded to increase clutch longevity but decrease the lubricant flow to the clutch during clutch disengagement to decrease drag torque and correspondingly increase transmission efficiency. 
       FIGS.  2 - 5    are drawn approximately to scale. However, other relative component dimensions may be used, in other embodiments. 
       FIGS.  1 - 5  and  8    show example configurations with relative positioning of the various components. If shown directly contacting each other, or directly coupled, then such elements may be referred to as directly contacting or directly coupled, respectively, at least in one example. Similarly, elements shown contiguous or adjacent to one another may be contiguous or adjacent to each other, respectively, at least in one example. As an example, components laying in face-sharing contact with each other may be referred to as in face-sharing contact. As another example, elements positioned apart from each other with only a space there-between and no other components may be referred to as such, in at least one example. As yet another example, elements shown above/below one another, at opposite sides to one another, or to the left/right of one another may be referred to as such, relative to one another. Further, as shown in the figures, a topmost element or point of element may be referred to as a “top” of the component and a bottommost element or point of the element may be referred to as a “bottom” of the component, in at least one example. As used herein, top/bottom, upper/lower, above/below, may be relative to a vertical axis of the figures and used to describe positioning of elements of the figures relative to one another. As such, elements shown above other elements are positioned vertically above the other elements, in one example. As yet another example, shapes of the elements depicted within the figures may be referred to as having those shapes (e.g., such as being circular, straight, planar, curved, rounded, chamfered, angled, or the like). Additionally, elements co-axial with one another may be referred to as such, in one example. Further, elements shown intersecting one another may be referred to as intersecting elements or intersecting one another, in at least one example. Further still, an element shown within another element or shown outside of another element may be referred as such, in one example. In other examples, elements offset from one another may be referred to as such. 
     The invention will be further described in the following paragraphs. In one aspect, a hydraulic system in a transmission is provided that comprises a lubrication valve positioned in a lubricant line in fluidic communication with a multi-disc wet clutch; a clutch line in fluidic communication with a clutch actuator of the multi-disc wet clutch and a passive adjustment device of the lubrication valve; and a clutch control valve coupled to the clutch line; wherein the passive adjustment device transitions the lubrication valve between a limited flow state and an open flow state based on a pressure of a hydraulic fluid in the clutch line. 
     In another aspect, a transmission is provided that comprises a hydraulic system including a lubrication valve positioned in a lubrication line in fluidic communication with a plurality of discs in a multi-disc wet clutch; a clutch line fluidic communication with an actuation piston of the multi-disc wet clutch; a clutch control valve designed to adjust a pressure of a hydraulic fluid in the clutch line; wherein a passive adjustment device of the lubrication valve configured to transition the lubrication valve between a limited flow state and an open flow state based on a pressure of the hydraulic fluid in the clutch line; and a controller including instructions that, when executed after the transmission is shifted into a second gear from a first gear, cause the controller to adjust the pressure in the clutch line to transition the lubrication valve from an open flow state to a choked flow state via operation of the clutch control valve to decrease a pressure of the oil in the clutch line below a threshold pressure. 
     In yet another aspect, a method for operating a hydraulic system in a transmission is provided that comprises adjusting a pressure of a hydraulic fluid in a clutch line to switch a lubrication valve between an open flow state and a limited flow state; wherein the lubrication valve is positioned in a valve conduit in fluidic communication with a plurality of discs in a multi-disc hydraulic clutch; and wherein the clutch line is in fluidic communication with a passive adjustment device of the lubrication valve and a piston actuator of the multi-disc hydraulic clutch. In one example, adjusting the pressure in the clutch line may include maintaining the pressure of the hydraulic fluid in the clutch line above a threshold pressure that adjusts the passive adjustment device to transition between the open flow state and the limited flow state in the lubrication valve based on an operating gear ratio of the transmission. In another example, adjusting the pressure in the clutch line may include transitioning the lubrication valve from the open flow state to the limited flow state via decreasing the pressure of the hydraulic fluid in the clutch line below the threshold pressure, in response to a difference between a current transmission ratio and a shift event transmission ratio exceeding a threshold value, wherein the shift event includes switching the transmission from a first gear to a second gear. In yet another example, the method may include, while the pressure of the hydraulic fluid in the clutch line is sustained above the threshold pressure, holding the transmission in the first gear via maintaining engagement of the multi-disc hydraulic clutch. In another example, transitioning the lubrication valve from the open flow state to the limited flow state may include discretely transitioning the lubrication valve from the open flow state to the limited flow state. In another example, the method may include rotating a shaft in which the lubrication valve is positioned while the pressure of the hydraulic fluid in the clutch line is adjusted. 
     In any of the aspects or combinations of the aspects, the hydraulic system may further comprise a controller including instructions that when executed, during operation of the transmission, cause the controller to: determine a desired lubrication valve state based on a transmission ratio; and adjust the pressure in the clutch line via clutch control valve actuation to reconfigure the passive adjustment device to place the lubrication valve in the limited flow state or the open flow state based on a transmission shift event. 
     In any of the aspects or combinations of the aspects, adjusting the pressure in the clutch line via clutch control valve actuation may include: decreasing the pressure in the clutch line below a threshold pressure that transitions the lubrication valve from the open flow state to the limited flow state, in response to a difference between a current transmission ratio and a shift point transmission ratio exceeding a threshold value. 
     In any of the aspects or combinations of the aspects, adjusting the pressure in the clutch line via clutch control valve actuation may include: sustaining the pressure in the clutch line above a threshold pressure that transitions the lubrication valve from the open flow state to the limited flow state, while the difference between the current transmission ratio and the shift point transmission ratio remains below a threshold value. 
     In any of the aspects or combinations of the aspects, the lubrication valve may continuously transition between the limited flow state and the open flow state. 
     In any of the aspects or combinations of the aspects, the lubrication valve may discretely transition between the limited flow state and the open flow state. 
     In any of the aspects or combinations of the aspects, the lubrication valve, the lubrication line, and the clutch line may be included in a rotating shaft. 
     In any of the aspects or combinations of the aspects, the passive adjustment device may not be directly electronically actuated. 
     In any of the aspects or combinations of the aspects, the limited flow state may be a choked flow state. 
     In any of the aspects or combinations of the aspects, the transmission may further comprise instructions that when executed, while the transmission is maintained in the first gear, cause the controller to: sustain the lubrication valve in the open flow state via operation of the clutch control valve to maintain the pressure of the oil in the clutch line above the threshold pressure. 
     In any of the aspects or combinations of the aspects, the lubrication valve may be transitioned to the choked flow state from the open flow state in response to a duration of the transmission remaining in the second gear after the transmission shift event exceeding a threshold duration. 
     In any of the aspects or combinations of the aspects, the first gear may be a first forward drive gear and the second gear may be a first reverse drive gear or a second forward drive gear. 
     In any of the aspects or combinations of the aspects, the transmission may be a hydromechanical variable transmission (HVT). 
     In another representation, a lubrication system in a hydromechanical transmission is provided that comprises a passive pressure controlled lubricant valve that includes an pressure adjustable actuator that is in fluidic communication with a clutch conduit that extends between a clutch control valve and a clutch control piston; and a controller configured to adjust the clutch control valve to selectively switch the lubricant valve between an open flow and a restricted flow configuration based on a transmission shift event. 
     Note that the example control and estimation routines included herein can be used with various powertrain and/or vehicle system configurations. The control methods and routines disclosed herein may be stored as executable instructions in non-transitory memory and may be carried out by the control system including the controller in combination with the various sensors, actuators, and other transmission and/or vehicle hardware. Further, portions of the methods may be physical actions taken in the real world to change a state of a device. The specific routines described herein may represent one or more of any number of processing strategies such as event-driven, interrupt-driven, multi-tasking, multi-threading, and the like. As such, various actions, operations, and/or functions illustrated may be performed in the sequence illustrated, in parallel, or in some cases omitted. Likewise, the order of processing is not necessarily required to achieve the features and advantages of the example examples described herein, but is provided for ease of illustration and description. One or more of the illustrated actions, operations and/or functions may be repeatedly performed depending on the particular strategy being used. Further, the described actions, operations and/or functions may graphically represent code to be programmed into non-transitory memory of the computer readable storage medium in the vehicle and/or transmission control system, where the described actions are carried out by executing the instructions in a system including the various hardware components in combination with the electronic controller. One or more of the method steps described herein may be omitted if desired. 
     While various embodiments have been described above, it should be understood that they have been presented by way of example, and not limitation. It will be apparent to persons skilled in the relevant arts that the disclosed subject matter may be embodied in other specific forms without departing from the spirit of the subject matter. The embodiments described above are therefore to be considered in all respects as illustrative, not restrictive. Thus, the configurations and routines disclosed herein are exemplary in nature, and that these specific examples are not to be considered in a limiting sense, because numerous variations are possible. For example, the above technology can be applied to powertrains that include different types of propulsion sources including different types of electric machines, internal combustion engines, and/or transmissions. The subject matter of the present disclosure includes all novel and non-obvious combinations and sub-combinations of the various systems and configurations, and other features, functions, and/or properties disclosed herein. 
     As used herein, the terms “substantially” and “approximately” are construed to mean plus or minus five percent of the range, unless otherwise specified. 
     The following claims particularly point out certain combinations and sub-combinations regarded as novel and non-obvious. These claims may refer to “an” element or “a first” element or the equivalent thereof. Such claims should be understood to include incorporation of one or more such elements, neither requiring nor excluding two or more such elements. Other combinations and sub-combinations of the disclosed features, functions, elements, and/or properties may be claimed through amendment of the present claims or through presentation of new claims in this or a related application. Such claims, whether broader, narrower, equal, or different in scope to the original claims, also are regarded as included within the subject matter of the present disclosure.