Method for controlling multiple EVT shifts in a multi-mode hybrid transmission

A method of performing shifts includes determining whether a multiple-shift maneuver is needed, whether a single-staged input profile is needed, and creating the single-staged input speed profile. The profile is matched to first or second multiple-shift patterns, neither of which utilizes fixed-gear propulsion. The patterns utilize a quasi-asynchronous transitional shift event and an electric torque converter transitional shift event. The quasi-asynchronous event induces controlled slip to an offgoing clutch while providing reaction torque from the electric machines, and offloads torque from the offgoing clutch proportionally to reaction torque. The oncoming clutch begins slipping-engagement prior to completing offloading of the offgoing clutch. At least one of the offgoing and oncoming clutches has non-zero slip speed throughout the quasi-asynchronous event. The electric torque converter event utilizes oncoming and offgoing clutches, completely offloads the offgoing clutch while controlling output torque with the electric machines, and synchronizes and engages the oncoming clutch.

TECHNICAL FIELD

The present invention relates generally to hybrid powertrains for motorized vehicles, and hydraulic control thereof.

BACKGROUND OF THE INVENTION

Motorized vehicles include a powertrain operable to propel the vehicle and power the onboard vehicle electronics. The powertrain, or drivetrain, generally includes an engine that powers the final drive system through a multi-speed power transmission. Many vehicles are powered by a reciprocating-piston type internal combustion engine (ICE).

Hybrid vehicles utilize alternative power sources to propel the vehicle, minimizing reliance on the engine for power. A hybrid electric vehicle (HEV), for example, incorporates both electric energy and chemical energy, and converts the same into mechanical power to propel the vehicle and power the vehicle systems. The HEV generally employs one or more electric machines that operate individually or in concert with an internal combustion engine to propel the vehicle. Since hybrid vehicles can derive their power from sources other than the engine, engines in hybrid vehicles may be turned off while the vehicle is stopped or is being propelled by the alternative power source(s).

Parallel hybrid architectures are generally characterized by an internal combustion engine and one or more electric motor/generator assemblies, all of which have a direct mechanical coupling to the transmission. Parallel hybrid designs utilize combined electric motor/generators, which provide traction and may replace both the conventional starter motor and alternator. The motor/generators are electrically connected to an energy storage device (ESD). The energy storage device may be a chemical battery. A control unit is employed for regulating the electrical power interchange between the energy storage device and motor/generators, as well as the electrical power interchange between the first and second motor/generators.

Electrically-variable transmissions (EVT) provide for continuously variable speed ratios by combining features from both series and parallel hybrid powertrain architectures, and also elements of traditional, non-hybrid transmissions. EVTs may be designed to operate in both fixed-gear (FG) modes and EVT modes. When operating in a fixed-gear mode, the rotational speed of the transmission output member is a fixed ratio of the rotational speed of the input member from the engine, depending upon the selected arrangement of the differential gearing subsets. EVTs may also be configured for engine operation that is mechanically independent from the final drive.

The EVT can utilize the differential gearing to send a fraction of its transmitted power through the electric motor/generator(s) and the remainder of its power through another, parallel path that is mechanical. One form of differential gearing used is the epicyclic planetary gear arrangement. However, it is possible to design a power split transmission without planetary gears, for example, as by using bevel gears or other differential gearing.

Hydraulically-actuated torque-transmitting mechanisms, such as clutches and brakes, are selectively engageable to selectively activate the gear elements for establishing different forward and reverse speed ratios and modes between the transmission input and output shafts. The term “clutch” is used hereinafter to refer generally to torque transmitting mechanisms, including, without limitation, devices commonly referred to as clutches and brakes. Shifting from one speed ratio or mode to another may be in response to vehicle conditions and operator (driver) demands. The “speed ratio” is generally defined as the transmission input speed divided by the transmission output speed. Thus, a low gear range has a high speed ratio, and a high gear range has a relatively lower speed ratio. The different operating states of an EVT may be referred to as ranges or modes.

SUMMARY OF THE DISCLOSURE

A method of performing shifts in hybrid powertrains is provided. The powertrain is configured for operation in multiple electrically variable transmission modes (EVT) and includes an internal combustion engine, a first electric machine, and a second electric machine. The method includes determining whether a multiple-shift maneuver is needed. The multiple-shift maneuver includes shifting from an initial EVT mode to an intermediate EVT mode, and then from the intermediate EVT mode to a target EVT mode. The method determines whether a single-staged input profile is needed to execute the multiple-shift maneuver and creates the single-staged input speed profile.

The method matches the single-staged input speed profile with either a first or a second multiple-shift pattern. Both of the multiple-shift patterns are characterized by a lack of fixed-gear propulsion during the multiple-shift maneuver, including shifts from the initial EVT mode, to the intermediate EVT mode, and to the target EVT mode. The method performs one of the first and second multiple-shift patterns by utilizing both a quasi-asynchronous transitional shift event and an electric torque converter transitional shift event.

The quasi-asynchronous transitional shift event utilizes a first oncoming clutch and a first offgoing clutch, and includes inducing controlled slip to the first offgoing clutch while providing reaction torque from at least one of the electric machines. The quasi-asynchronous transitional shift event offloads torque from the first offgoing clutch in proportion to the provided reaction torque of the electric machines. The quasi-asynchronous transitional shift event begins slipping-engagement of the first oncoming clutch prior to completing offloading of the first offgoing clutch, and at least one of the first offgoing clutch and the first oncoming clutch is characterized by non-zero slip speed throughout the duration of the shift event.

The electric torque converter transitional shift event utilizes a second oncoming clutch and a second offgoing clutch, and includes completely offloading the second offgoing clutch and controlling output torque of the transmission with at least one of the electric machines. The electric torque converter transitional shift event synchronizes then engages the second oncoming clutch. The method performs one of the first and second multiple-shift patterns without utilizing or entering fixed-gear propulsion.

The above features and advantages, and other features and advantages of the present invention will be readily apparent from the following detailed description of the preferred embodiments and other modes for carrying out the present invention when taken in connection with the accompanying drawings and appended claims.

DESCRIPTION OF PREFERRED EMBODIMENTS

The claimed invention is described herein in the context of a hybrid-type vehicular powertrain having a multi-mode, multi-speed, electrically-variable, hybrid transmission, which is intended solely to offer a representative application by which the present invention may be incorporated and practiced. The claimed invention is not limited to the particular powertrain arrangement shown in the drawings. Furthermore, the hybrid powertrain illustrated herein has been greatly simplified, it being understood that further information regarding the standard operation of a hybrid powertrain, or a hybrid-type vehicle will be recognized by those having ordinary skill in the art.

Referring to the drawings, wherein like reference numbers refer to like components throughout the several views, there is shown inFIG. 1a lever diagram depiction of an exemplary vehicle powertrain system, designated generally as10. The powertrain10includes a restartable engine12that is selectively drivingly connected to, or in power flow communication with, a final drive system16via a multi-mode, electrically-variable hybrid-type power transmission14.

A lever diagram is a schematic representation of the components of a mechanical device such as an automatic transmission. Each individual lever represents a planetary gearset, wherein the three basic mechanical components of the planetary gear are each represented by a node. Therefore, a single lever contains three nodes: one for the sun gear member, one for the planet gear carrier member, and one for the ring gear member. The relative length between the nodes of each lever may be used to represent the ring-to-sun ratio of each respective gearset. These lever ratios, in turn, are used to vary the gear ratios of the transmission in order to achieve appropriate ratios and ratio progression. Mechanical couplings or interconnections between the nodes of the various planetary gear sets and other components of the transmission (such as motor/generators) are illustrated by thin, horizontal lines. Torque transmitting devices such as clutches and brakes are presented as interleaved fingers. If the device is a brake, one set of the fingers is grounded.

The transmission14is designed to receive at least a portion of its driving power from the engine12, through an input member18, for example. The transmission input member18, which is in the nature of a shaft, may be the engine output shaft (also referred to as a “crankshaft”). Alternatively, a transient torque damper (not shown) may be implemented between the engine12and the input member18of the transmission14. The engine12transfers power to the transmission14, which distributes torque through a transmission output member or shaft20to drive the final drive system16, and thereby propel the vehicle (not shown).

In the embodiment depicted inFIG. 1, the engine12may be any of numerous forms of petroleum-fueled prime movers, such as the reciprocating-piston type internal combustion engines, which includes spark-ignited gasoline engines and compression-ignited diesel engines. The engine12is readily adaptable to provide its available power to the transmission14at a range of operating speeds, for example, from idle, at or near 600 revolutions per minute (RPM), to over 6,000 RPM. Irrespective of the means by which the engine12is connected to the transmission14, the input member18is connected to a differential gear set encased within the transmission14, as explained in more detail herein.

Referring still toFIG. 1, the hybrid transmission14utilizes one or more differential gear arrangements, preferably in the nature of three interconnected epicyclic planetary gear sets, designated generally at24,26and28, respectively. Each gear set includes three gear members: a first, second and third member. In referring to the first, second and third gear sets in this description and in the claims, these sets may be counted “first” to “third” in any order in the drawings (e.g., left to right, right to left, etc.). Likewise, in referring to the first, second and third members of each gear set in this description and in the claims, these members may be counted or identified as “first” to “third” in any order in the drawings (e.g., top to bottom, bottom to top, etc.) for each gear set.

The first planetary gear set24has three gear members: a first, second and third member30,32and34; respectively. The first, second and third members correspond to the first, second and third nodes of the lever diagram shown inFIG. 1, as viewed from top to bottom. In a preferred embodiment, the first member30includes of an outer gear member (which may be referred to as a “ring gear”) that circumscribes the third member34, which may include of an inner gear member (which may be referred to as a “sun gear”). In this instance, the second member32acts as a planet carrier member. That is, a plurality of planetary gear members (which may be referred to as “pinion gears”) are rotatably mounted on the second member, planet carrier32. Each planetary gear member is meshingly engaged with both the first member, ring gear30and the third member, sun gear34.

The second planetary gear set26also has three gear members: a first, second and third member40,42and44, respectively. In the preferred embodiment discussed above with respect to the first planetary gear set24, the first member40of the second planetary gear set26is an outer “ring” gear member that circumscribes the third member44, which is an inner “sun” gear member. The ring gear member40is coaxially aligned and rotatable with respect to the sun gear member44. A plurality of planetary gear members are rotatably mounted on the second member42, which acts as a planet carrier member, such that each planetary gear meshingly engages both the ring gear member40and the sun gear member44.

The third planetary gear set28, similar to the first and second gear sets24,26, also has first, second and third members50,52and54, respectively. In this arrangement, however, the second member52, shown on the middle node of the lever representing the third planetary gear set28, is the outer “ring” gear. The ring gear (second member52) is coaxially aligned and rotatable with respect to the sun gear, third member54. The first member50is the planet carrier in this particular gear set, and is shown on the top node. As such, a plurality of planetary or pinion gear members are rotatably mounted on the planet carrier, first member50. Each of the pinion gear members is aligned to meshingly engage either the ring gear (second member52) and an adjacent pinion gear member or the sun gear (third member54) and an adjacent pinion gear member.

In one embodiment, the first and second planetary gear sets24,26each comprise simple planetary gear sets, whereas the third planetary gear set28comprises a compound planetary gear set. However, each of the planet carrier members described above can be either a single-pinion (simple) carrier assembly or a double-pinion (compound) carrier assembly. Embodiments with long pinions are also possible.

The first, second and third planetary gear sets24,26,28are compounded in that the second member32of the first planetary gear set24is conjoined with (i.e., continuously connected to) the second member42of the second planetary gear set26and the third member54of the third planetary gear set28, as by a central shaft36. As such, these three gear members32,42,54are rigidly attached for common rotation.

The engine12is continuously connected to the first planetary gear set24, namely first member30, for example, through an integral hub plate38, for common rotation therewith. The third member34of the first planetary gear set24is continuously connected, for example, by a first sleeve shaft46, to a first motor/generator assembly56, which is also referred to herein as “motor A”. The third member44of the second planetary gear set26is continuously connected, for example, by a second sleeve shaft48, to a second motor/generator assembly58, also referred to herein as “motor B”. The second member52(the ring gear) of the third planetary gear set28is continuously connected to transmission output member20, for example, through an integral hub plate. The first and second sleeve shafts46,48may circumscribe the central shaft36.

A first torque transfer device70—which is herein interchangeably referred to as clutch “C1”—selectively connects the first gear member50with a stationary member, represented inFIG. 1by transmission housing60. The second sleeve shaft48, and thus gear member44and motor/generator58, is selectively connectable to the first member50of the third planetary gear set28through the selective engagement of a second torque transfer device72—which is herein interchangeably referred to as clutch “C2”. A third torque transfer device74—which is herein interchangeably referred to as clutch “C3”—selectively connects the first gear member40of the second planetary gear set26to the transmission housing60. The first sleeve shaft46, and thus third gear member34and first motor/generator56, is also selectively connectable to the first member40of the second planetary gear set26, through the selective engagement of a fourth torque transfer device76—which is herein interchangeably referred to as clutch “C4”.

A fifth torque transfer device78—which is herein interchangeably referred to as clutch “C5”—selectively connects the input member18of engine12and the first gear member30of the first planetary gear set24to the transmission housing60. Clutch C5is an input brake clutch, which selectively locks the input member18when engine12is off. Locking input member18provides more reaction for regenerative braking energy. As shown below, in reference toFIG. 2, C5is not involved in the mode/gear/neutral shifting maneuvers of transmission14.

The first and second torque transfer devices70,72(C1and C2) may be referred to as “output clutches.” The third and fourth torque transfer devices74,76(C3and C4) may be referred to as “holding clutches”.

In the exemplary embodiment depicted inFIG. 1, the various torque transfer devices70,72,74,76,78(C1-C5) are all friction clutches. However, other conventional clutch configurations may be employed, such as dog clutches, rocker clutches, and others recognizable to those having ordinary skill in the art. The clutches C1-C5may be hydraulically actuated, receiving pressurized hydraulic fluid from a pump (not shown). Hydraulic actuation of clutches C1-C5is accomplished, for example, by using a conventional hydraulic fluid control circuit, as will be recognized by one having ordinary skill in the art.

In the exemplary embodiment described herein, wherein the hybrid powertrain10is used as a land vehicle, the transmission output shaft20is operatively connected to the final drive system (or “driveline”). The driveline may include a front or rear differential, or other torque transfer device, which provides torque output to one or more wheels through respective vehicular axles or half-shafts (not shown). The wheels may be either front or rear wheels of the vehicle on which they are employed, or they may be a drive gear of a track vehicle. Those having ordinary skill in the art will recognize that the final drive system may include any known configuration, including front wheel drive (FWD), rear wheel drive (RWD), four-wheel drive (4WD), or all-wheel drive (AWD), without altering the scope of the claimed invention.

All of the planetary gear sets24,26,28, as well as the first and second motor/generators56,58(motor A and motor B,) are preferably coaxially oriented about the intermediate central shaft36or another axis. Motor A or motor B may take on an annular configuration, permitting one or both to generally circumscribe the three planetary gear sets24,26,28. Such a configuration may reduce the overall envelope, i.e., the diametrical and longitudinal dimensions, of the hybrid transmission14are minimized.

The hybrid transmission14receives input motive torque from a plurality of torque-generative devices. “Torque-generative devices” include the engine12and the motors/generators56,58as a result of energy conversion from fuel stored in a fuel tank or electrical potential stored in an electrical energy storage device (neither of which is shown).

The engine12, motor A (56,) and motor B (58) may operate individually or in concert—in conjunction with the planetary gear sets and selectively-engageable torque-transmitting mechanisms—to rotate the transmission output shaft20. Moreover, motor A and motor B are preferably configured to selectively operate as both a motor and a generator. For example, motor A and motor B are capable of converting electrical energy to mechanical energy (e.g., during vehicle propulsion), and further capable of converting mechanical energy to electrical energy (e.g., during regenerative braking or during periods of excess power supply from engine12).

With continuing reference toFIG. 1, an electronic control apparatus (or “controller”) having a distributed controller architecture is shown schematically in an exemplary embodiment as a microprocessor-based electronic control unit (ECU)80. The ECU80includes a storage medium with a suitable amount of programmable memory, collectively represented at82, that is programmed to include, without limitation, algorithm or method100of regulating operation of the multi-mode hybrid transmission, as will be discussed in further detail below with respect toFIG. 4.

The control apparatus is operable, as described hereinafter, to provide coordinated system control of the powertrain10schematically depicted and described herein. The constituent elements of the control apparatus may be a subset of an overall vehicle control system. The control system is operable to synthesize pertinent information and inputs, and execute control methods and algorithms to control various actuators to achieve control targets. The control system monitors target and parameters including, without limitation: fuel economy, emissions, performance, driveability, and protection of drivetrain hardware—such as, but not limited to, the engine12, transmission14, motor A, motor B, and final drive16.

The distributed controller architecture (ECU80) may include a Transmission Control Module (TCM), an Engine Control Module (ECM), a Transmission Power Inverter Module (TPIM), and a Battery Pack Control Module (BPCM). A Hybrid Control Module (HCP) may be integrated to offer overall control and coordination of the aforementioned controllers.

A User Interface (UT) is operatively connected to a plurality of devices (not shown) through which a vehicle operator typically controls or directs operation of the powertrain. Exemplary vehicle operator inputs to the user interface include an accelerator pedal, a brake pedal, transmission gear selector, vehicle speed cruise control, and other inputs recognizable to those having ordinary skill in the art.

Each of the aforementioned controllers communicates with other controllers, sensors, actuators, etc., via a control area network (CAN) bus or communication architecture. The CAN bus allows for structured communication of control parameters and commands between the various controllers. The communication protocol utilized is application-specific. For example, and without limitation, one useable communication protocol is the Society of Automotive Engineers standard J1939. The CAN bus and appropriate protocols provide for robust messaging and multi-controller interfacing between the aforementioned controllers, and other controllers providing functionality such as antilock brakes, traction control, and vehicle stability.

The engine control module is operatively connected to, and in communication with, the engine12. The engine control module is configured to acquire data from a variety of sensors and control a variety of actuators of the engine12over a plurality of discrete lines. The engine control module receives an engine torque command from the hybrid control module, generates a desired axle torque, and an indication of actual engine torque, which is communicated to the hybrid control module. Various other parameters that may be sensed by the engine control module include engine coolant temperature, engine input speed to the transmission, manifold pressure, and ambient air temperature and pressure. Various actuators that may be controlled by the engine control module include, without limitation, fuel injectors, ignition modules, and throttle control modules.

The transmission control module is operatively connected to the transmission14, and functions to acquire data from a variety of sensors and provide command signals to the transmission14. Inputs from the transmission control module to the hybrid control module may include estimated clutch torques for each of the clutches C1-C5, and rotational speed of the transmission output shaft20. Additional actuators and sensors may be used to provide additional information from the transmission control module to the hybrid control module for control purposes.

Each of the aforementioned controllers may be a general-purpose digital computer, generally including a microprocessor or central processing unit, read only memory (ROM), random access memory (RAM), electrically programmable read only memory (EPROM), high speed clock, analog to digital (A/D) and digital to analog (D/A) circuitry, and input/output circuitry and devices (I/O) and appropriate signal conditioning and buffer circuitry. Each controller has a set of control algorithms, including resident program instructions and calibrations stored in ROM and executed to provide the respective functions of each computer. Information transfer between the various computers may be accomplished using the aforementioned CAN.

In response to operator input, as captured by the user interface, the supervisory hybrid control module controller and one or more of the other controllers described above with respect toFIG. 1determine required transmission output torque. Selectively operated components of the hybrid transmission14are appropriately controlled and manipulated to respond to the operator demand. For example, in the embodiment shown inFIG. 1, when the operator has selected a forward drive range and manipulates either the accelerator pedal or the brake pedal, the hybrid control module determines an output torque for the transmission, which affects how and when the vehicle accelerates or decelerates. Final vehicle acceleration is affected by other variables, including such factors as road load, road grade, and vehicle mass. The hybrid control module monitors the parametric states of the torque-generative devices, and determines the output of the transmission required to arrive at the desired torque output. Under the direction of the hybrid control module, the transmission14operates over a range of output speeds from slow to fast in order to meet the operator demand.

The ECU80also receives frequency signals from sensors for processing into input member18speed, Ni, and output member20speed, No, for use in the control of transmission14. The system controller may also receive and process pressure signals from pressure switches (not shown) for monitoring clutch application chamber pressures. Alternatively, pressure transducers for wide range pressure monitoring may be employed. Pulse-width modulation (PWM) and/or binary control signals are transmitted by the controller80to transmission14for controlling fill and drain of clutches C1-C5for application and release thereof.

Additionally, the controller80may receive transmission fluid sump temperature data, such as from thermistor inputs (not shown), to derive a sump temperature. Controller80may provide PWM signals derived from input speed, Ni, and sump temperature for control of line pressure via one or more regulators.

Fill and drain of clutches C1-C5may be effectuated, for example, by solenoid controlled spool valves responsive to PWM and binary control signals. Trim valves may be employed using variable bleed solenoids to provide precise placement of the spool within the valve body and correspondingly precise control of clutch pressure during apply. Similarly, one or more line pressure regulators (not shown) may be utilized for establishing regulated line pressure in accordance with the control signal. Clutch slip speeds across clutches may be derived from, for example: transmission input speed, output speed, motor A speed, and/or motor B speed.

The multi-mode, electrically-variable, hybrid transmission14is configured for several transmission operating modes. The truth table provided inFIG. 2presents an exemplary engagement schedule of the torque-transmitting mechanisms C1-C4to achieve the array of operating states or modes. The various transmission operating modes described in the table indicate which of the specific clutches C1-C4are engaged (actuated), and which are released (deactivated) for each of the operating modes.

In general, ratio changes in transmission14may be performed such that torque disturbances are minimized, and the shifts are smooth and unobjectionable to the vehicle occupants. Additionally, release and application of clutches C1-C4should be performed in a manner which consumes the least amount of energy, and does not negatively impact durability of the clutches. One major factor affecting these considerations is the torque at the clutch being controlled, which may vary significantly in accordance with such performance demands as acceleration and vehicle loading. Improved shifts may be accomplished by a zero, or close to zero, reactive torque condition at the clutches at the time of application or release, which condition follows substantially zero slip across the clutch. Clutches having zero slip across the clutch may be referred to as operating synchronously.

Electrically-variable operating modes may be separated into four general classes: input-split modes, output-split modes, compound-split modes, and series modes. In an input-split mode, one motor/generator (such as either motor A or motor B) is geared such that its speed varies in direct proportion to the transmission output, and another motor/generator (such as the other of motor A or motor B) is geared such that its speed is a linear combination of the input and output member speeds. In an output-split mode, one motor/generator is geared such that its speed varies in direct proportion to the transmission input member, and the other motor/generator is geared such that its speed is a linear combination of the input member and the output member speeds. A compound-split mode, however, has both motor/generators geared such that their speeds are linear combinations of the input and output member speeds, but neither is in direct proportion to either the speed of the input member or the speed of the output member.

Finally, when operating in a series mode, one motor/generator is geared such that its speed varies in direct proportion to the speed of the transmission input member, and another motor/generator is geared such that its speed varies in direct proportion to the speed of the transmission output member. When operating in series mode, there is no direct mechanical power transmission path between the input and output members and therefore all power must be transmitted electrically.

In each of the four general types of electrically-variable operating modes indicated above, the speeds of the motors are linear combinations of the input and output speeds. Thus, these modes have two speed degrees of freedom (which may be abbreviated for simplicity as “DOF”). Mathematically, the torque (T) and speed (N) equations of this class of modes take the form:

[TaTb]=[a1,1a1,2a2,1a2,2]⁡[TiTo]⁢⁢and⁢[NaNb]=[b1,1b1,2b2,1b2,2]⁡[NiNo]
where a and b are coefficients determined by the transmission gearing. The type of EVT mode can be determined from the structure of the matrix of b coefficients. That is, if b2,1=b1,2=0 or b1,1=b2,2=0, the mode is a series mode. If b1,1=0 or b1,2=0, the mode is an input split mode. If b2,1=0 or b2,2=0, the mode is an output split mode. If each of b1,1, b1,2, b2,1, and b2,2are nonzero, for example, the mode is a compound split mode.

An electrically-variable transmission may also contain one or more fixed-gear (FG) modes. In general, fixed-gear modes result from closing (i.e., actuating) one additional clutch than the number required to select an electrically-variable mode. In fixed-gear modes, the speed of the input and each motor are proportional to the speed of the output. Thus, these modes have only one speed degree of freedom. Mathematically, the torque and speed equations of this class of modes take the form:

[Tb]=[a1,1a1,2a1,3]⁡[TaTiTo]⁢⁢and⁢[NaNbNi]=[b1,1b1,2b1,3]⁡[No]
where a and b are again coefficients determined by the transmission gearing. If b1,1is nonzero, motor A can contribute to output torque during operation in the fixed-gear mode. If b1,2is nonzero, motor B can contribute to output torque during operation in the fixed-gear mode. If b1,3is nonzero, the engine can contribute to output torque during operation in the fixed-gear mode. If b1,3is zero, the mode is an electric-only fixed-gear mode.

An electrically-variable transmission may also be configured for one or more modes with three speed degrees of freedom. These modes may or may not include reaction torque sources such that the transmission is capable of producing output torque proportional to engine torque or motor torque. If a mode with three speed degrees of freedom is capable of producing output torque, the torques of the engine and any motor connected as a reaction to the engine torque will generally be proportional to the output torque. If a motor is not connected as a reaction to the engine torque, its torque can be commanded to control its speed independently of the transmission input and output speed.

In a mode with three speed degrees of freedom, it is generally not possible to easily control battery power independently of output torque. This type of mode produces an output torque which is proportional to each of the reacting torque sources in the system. The fraction of total output power provided by each of the three torque sources may be adjusted by varying the speeds of the motors and input. These modes are hereafter referred to as electric torque converter (eTC) modes in recognition of the fact that power flows to or from the energy storage device as a function of the output torque and the speed of the engine, output, and one of the motors. Mathematically, the torque and speed equations of this class of modes take the form:

[TaTbTi]=[a1,1a1,2a1,3]⁡[To]⁢⁢and⁢[Nb]=[b1,1b1,2b1,3]⁡[NaNiNo]
where a and b are coefficients determined by the transmission gearing. If a1,1is nonzero, motor A serves as a reaction member and its torque is proportional to output torque when operating in the eTC mode. If a1, is zero, motor A is disconnected and its torque is not determined by the output torque. If a1,2is nonzero, motor B serves as a reaction member and its torque is proportional to output torque when operating in the eTC mode. If a1,2is zero, motor B is disconnected and its torque is not determined by the output torque. If a1,3is nonzero, the engine can contribute to output torque during operation in the fixed-gear mode. If a1,3is zero, the input is disconnected and its torque is not determined by the output torque. If all of a1,1, a1,2, and a1,3are zero, the mode is a neutral mode that is not capable of producing output torque.

There are four neutral modes presented inFIG. 2. In Neutral1, all clutches are released. Neutral1may be utilized when the entire vehicle is stopped and in an off-state, and thus there is no power distribution, electrical, mechanical, or otherwise, being actively distributed throughout the powertrain10. In Neutral1, a 12-volt starting-lighting-and-ignition (SLI) battery may be used for engine start.

In Neutral2, only clutch C3is engaged, and motor A and motor B may react engine12for start or to charge the energy storage device. Similar to Neutral2, when transmission14is in Neutral3, motor A and motor B may react engine12for start or to charge the energy storage device, and clutch C4as the only engaged torque-transmitting device. In Neutral4, the third and fourth clutches C3, C4are both in an activated state. In this instance, motor A is locked or “grounded”, and motor B is geared with the engine12for engine start.

The first and second planetary gear sets24,26cooperate with the first and second motor/generators56,58, along with the selective engagement of the first and second clutches C1, C2, to constitute an electric torque converter (eTC). For example, when the transmission14is operating in an eTC mode, the electric output of motor A and/or motor B, depending upon the active control schedule, can be adapted to control the transfer of torque from the engine12through the transmission differential gearing to the output member20. When the vehicle is started, ETC1Mode is established by engaging the first clutch C1. In ETC1Mode, motor A reacts engine12with the first and third planetary gear sets24,28, and motor B freewheels. In ETC1Mode, the stationary vehicle can be smoothly started with the engine12held at a suitable speed by gradually increasing the amount of electric power generated by motor A—i.e., the reaction force of motor A.

There are two other alternative eTC modes available utilizing the transmission configuration presented herein. ETC2Mode, also known as “compound eTC”, can be initiated by engaging clutch C2, and disengaging the remaining clutches. In ETC2Mode, motor A reacts engine12with the first and third planetary gear sets24,28, while motor B reacts engine12and motor A to the output member20. The distribution of engine torque is manipulated through the cooperative management of the amount of electric power output generated by motor A and motor B.

The third eTC mode, ETC12Mode, can be initiated by engaging both clutch C1and clutch C2. Similar to ETC1Mode, motor A reacts the engine12with the first and third planetary gear sets24,28. However, in this instance, motor B is grounded to the transmission housing60. In ETC12Mode, the vehicle can be smoothly accelerated with the engine12held at a suitable speed by gradually increasing the reaction force generated by motor A; which may be proportional to the electric power generated by motor A.

When the engine12is in an off-state, the transmission14can utilize the eTC mode clutch control schedule to vary the amount of electric energy generated by motor A so as to gradually increase the drive torque of motor A and/or motor B. For example, if the transmission14is shifted into ETC1Mode when the engine12is in an off-state, the engine12will create a reaction force, by way of input member18. The motive output of the motor A can then be controlled, and a continuous and uninterrupted transmission output torque maintained, without having to turn the engine12on.

The exemplary powertrain10described herein has three fixed-gear (FG), or “direct,” modes of operation. In all fixed-gear modes of this embodiment of transmission14, the vehicle is driven in the forward direction by operation of the engine12. The selective engagement of clutches C1, C3and C4shifts the transmission14into FG1Mode. In FG1, motor A is grounded, and the engine drives the first planetary gear set24to the third planetary gear set28and, thus, the output member20. FG2Mode is achieved by the selective engagement of clutches C1, C2and C4. In FG2, motor B is grounded, and the engine drives the first and second planetary gear sets24,26to the third planetary gear set28and, thus, the output member20. Likewise, FG3Mode is achieved by the selective engagement of clutches C2, C3and C4. In FG3, motor A is locked, and the engine drives the first planetary gear set24to the second and third planetary gear sets26,28and the output member20. When operating in a fixed-gear mode of operation, the output member speed Nois directly proportional to input member speed Niand the selected gear ratio. Ni=No×GR.

With continued reference toFIG. 2, the transmission14may also operate in four electrically-variable transmission (EVT) modes. In EVT1and EVT4, the transmission14is operating in an input-split mode of operation, wherein the output speed Noof the transmission14is proportional to the speed of one motor/generator56,58(motor A or motor B). Specifically, EVT1Mode is achieved through the selective engagement of the first and third clutches C1and C3. When in EVT1, motor A functions to react the engine12with the first planetary gear set24, to the third planetary gear set28, and the output member20; while motor B drives the second and third planetary gear sets26,28. Motor A propels the vehicle in EVT1. Alternatively, the transmission14may be selectively shifted into EVT4Mode by actuating clutch C2and clutch C3. In EVT4, motor A functions to react the engine12with the first planetary gear set24, to the second and third planetary gear sets26,28, and the output member20, while motor B drives the second and third planetary gear sets26,28. Motor B propels the vehicle in EVT4.

In EVT2and EVT3, the transmission14is operating in a compound-split mode, wherein the output speed Noof the transmission14is not proportional to the speed of a single motor/generator, but is rather an algebraic linear combination of the speeds of both motor/generators. More particularly, EVT2is achieved through the selective engagement of the first and fourth clutches C1, C4. In this mode, motor A and motor B operate to react the engine12with the first and second planetary gears sets. Alternatively, the transmission14may be selectively shifted into EVT3Mode by actuating clutch C2and clutch C4. When operating in EVT3Mode, the two motor/generator assemblies56,58react the engine12with all three planetary gear sets24,26,28.

With reference toFIG. 3, a plot of transmission output speed, No, along the horizontal axis versus input speed, Ni, across the vertical axis is illustrated.FIG. 3is only a graphical representation of exemplary regions of operation for each operating mode with respect to input and output speeds of this embodiment of transmission14.

Synchronous operation in FG1—the input speed and output speed relationships where clutches C1, C3and C4are operating with substantially zero slip speed thereacross—is represented by line91. As such, line91represents an input and output speed relationship at which substantially synchronous shifting between EVT modes can occur. FG1is also a range at which direct mechanical coupling from input to output can be effected by simultaneous application of clutches C1, C3and C4—i.e., fixed- or direct-ratio.

Synchronous operation in FG2—the input speed and output speed relationships where clutches C1, C2and C4are operating with substantially zero slip speed thereacross—is represented by line93. Similarly, the relationships between input and output speed during operation in FG3, whereat clutches C2, C3and C4are operating simultaneously with substantially zero slip speed thereacross, is represented by line95.

To the left of the shift ratio line91is an exemplary region of operation for the first EVT mode, EVT1, wherein both C1and C3are applied, and C2and C4are released. To the right of the shift ratio line91and left of shift ratio line93is an exemplary region of operation for the second EVT mode, EVT2, wherein C1and C4are applied, and C2and C3are released.

To the right of shift line93and left of shift ratio line95is an exemplary region of operation for the third EVT mode, EVT3, wherein both C2and C4are applied, and C1and C3are released. To the right of the shift ratio line95is an exemplary region of operation for the fourth EVT mode, EVT4, wherein C2and C3are applied, and C1and C4are released. As used herein with respect to clutches C1-C5, the terms “applied” or “actuated” indicate substantial torque transfer capacity across the respective clutch. Antithetically, the terms “released” or “deactivated” indicate insubstantial or no torque transfer capacity across the respective clutch.

While the regions of operation specified above may be generally favored for operation of the hybrid transmission14, it is not meant to imply that the various EVT regions of operation depicted inFIG. 3cannot or do not overlap. Generally, however, it may be preferred to operate in the specified regions because each particular mode of operation preferably employs gear sets and motor hardware particularly well suited in various aspects (e.g., mass, size, cost, inertial capabilities, etc.) for that region. Similarly, while the individual regions of operation specified above are generally preferred for the particular modes of operation indicated, it is not meant to imply that the regions of operation for the individual EVT modes cannot be switched.

Generally, a shift into Mode1may be considered a downshift and is associated with a higher gear ratio in accordance with the relationship of Ni/No. In contrast, a shift into Mode4is considered an upshift, and is associated with a lower gear ratio in accordance with the relationship of Ni/No. As discussed herein, other mode-to-mode shift sequences are feasible. For example, a shift from EVT1to EVT3is also an upshift, while a shift from EVT4to EVT2is considered a downshift.

During operation of powertrain10, an initiated shift sequence may be detected by the ECU80or hybrid control module. If a shift sequence is not initiated—for example, by operator command or change in vehicle operating conditions—the ECU80monitors the transmission14and will continue in its current state of operation. The requested shift sequence may be a multiple-shift maneuver, which includes an initial EVT mode, an intermediate EVT mode, and a target EVT mode.

Referring now toFIGS. 4A and 4B, the multiple-shift maneuver or event may occur in multiple ways. For example, the exemplary powertrain10shown inFIG. 1is configured to perform the multiple-shift maneuver by transitioning through a fixed-gear mode, as shown inFIG. 4A. Alternatively, as described in more detail below, the multiple-shift maneuver may utilize quasi-asynchronous (QA) and electric torque converter (eTC) modes.

As shown inFIG. 4A, fixed-gear modes create a proportional, direct-drive relationship between the engine12and the final drive16, determined by the gear ratio (GR) of the specific fixed gear. Therefore, transitioning through fixed-gear modes is available only if engine12is running. These proportional relationships are shown as lines102and104ofFIG. 4A, representing respective fixed-gear modes. On lines102and104, the relationship of input speed to output speed is: Ni=No*GR. If a multiple-shift maneuver is requested, the hybrid control module will determine if engine12is in an engine-on or engine-off state. If engine12is on, the hybrid control module will determine which shift sequence provides for more-optimal shifting under the current operating conditions.

FIG. 4Aschematically shows a fixed-gear multiple-shift maneuver110. The input speed profile, Ni, is shown on line112. For illustrative purposes only, the fixed-gear multiple-shift maneuver110will be discussed herein with lines102and104representing FG1and FG2, respectively. The fixed-gear multiple-shift maneuver110, therefore, represents—as the input speed moves from left to right—shifts from EVT3to EVT2, through FG2, and from EVT2to EVT1, through FG1.

During the fixed-gear multiple-shift maneuver110, as input speed passes through the fixed-gear modes, the input speed, line112, dwells on the corresponding fixed ratios. These fixed-gear dwell portions120and122, corresponding to FG2and FG1, respectively, cause inflection points in the line112. More importantly, the fixed-gear dwell portions120and122are likely felt by the vehicle operator as changes in vehicle acceleration, sound, or driving feel.

FIG. 4Bschematically shows a single-staged multiple-shift maneuver160. Unlike the fixed-gear multiple-shift maneuver110shown inFIG. 4A, the single-staged multiple-shift maneuver160does not dwell on the fixed-gear ratios as input speed, shown on line162, passes lines FG2and FG1, lines102and104, respectively. As input speed increases, the transmission14shifts from the initial EVT mode to the intermediate EVT mode and to the target EVT mode. The single-staged multiple-shift maneuver160uses transitional shift events170and172to maintain a smooth, single-staged input speed profile. The input speed line162does not dwell on the fixed-gear ratios and the vehicle operator feels, at most, only a single shift.

In the corresponding example provided for the fixed-gear multiple-shift maneuver110, the single-staged multiple-shift maneuver160may shift from EVT3(initial mode) to EVT2(intermediate mode) and from EVT2to EVT1(target mode). Those having ordinary skill in the art will recognize that the modes (initial mode, intermediate mode, and target mode) may also be referred to as first, second, and third EVT modes. The important thing is that the transmission10shift through three individual EVT modes along the single-staged input speed profile. As described in more detail below, this specific single-staged multiple-shift maneuver160may utilize a QA shift for transitional shift event170and an eTC shift for transitional shift event172. Those having ordinary skill in the art will also recognize that, althoughFIGS. 4A and 4Bdemonstrate a downshift, the control methods described herein may be applied to an upshift (e.g. EVT1-EVT2-EVT3).

Referring now toFIG. 5, and with continued reference toFIGS. 1-4B, there is shown a schematic flow chart of a control method200for executing a multiple-shift maneuver, such as the single-staged multiple-shift maneuver160shown inFIG. 4B. Method200begins at a start or initiation step202, which may coincide with vehicle ignition (which may not include actual ignition of the engine12) or another starting event. The method200monitors operating conditions of the powertrain10and transmission14in step204.

The method or algorithm200is described herein with respect to the structure illustrated inFIG. 1, preferably executed as algorithms in the controllers of the control system described above, to control operation of the system described with reference toFIG. 1. However, those having ordinary skill in the art will recognize that the present invention may also be incorporated into other powertrain arrangements without departing from the intended scope of the claimed invention.

The ECU80or hybrid control module may determine that a multiple-shift maneuver is needed in step206. For purposes of the method200, the multiple-shift maneuver again includes an initial EVT mode, intermediate EVT mode, and target EVT mode. The method200includes determining whether a single-staged input profile is needed in step208, as opposed to executing the fixed-gear multiple-shift maneuver110shown inFIG. 4A. When the method200determines either that the multiple-shift maneuver is not required or that the single-staged input profile is not required, the method200reverts to monitoring operating conditions at step204.

If needed, the method200creates the single-staged input speed profile, such as the input speed profile shown on line162ofFIG. 4B, in step210. At step212, method200selects a shift pattern to match the single-staged multiple-shift maneuver160to the input speed profile created in step210. A first multiple-shift pattern involves first utilizing the eTC transitional shift event in step214followed by the QA transitional shift event in step216. A second multiple-shift pattern involves first utilizing the QA transitional shift event in step218followed by the eTC transitional shift event in step220.

For the exemplary single-staged multiple-shift maneuver160shown inFIG. 4Band discussed above. The multiple-shift maneuver involves shifting from EVT3to EVT2with the QA shift for transitional shift event170and shifting from EVT2to EVT1with the eTC shift for transitional shift event172. Therefore, for this exemplary maneuver, the method200would select the second multiple-shift pattern and steps218and220.

Both the first multiple-shift pattern and the second multiple-shift pattern are characterized by a lack of fixed-gear propulsion during the initial EVT mode, the intermediate EVT mode, and the target EVT mode. Execution of the eTC transitional shift event and QA transitional shift event are described in more detail below with regard toFIGS. 6 and 7.

Referring now toFIG. 6, and with continued reference toFIGS. 1-5, there is shown an eTC method300of executing an eTC transitional shift event. The eTC method300is a more-detailed process showing the control method for either of steps214or220in the method200ofFIG. 5.

For illustrative purposes, the eTC method300will be described in relation to an EVT2-to-EVT1downshift, generally corresponding to step220ofFIG. 5and transitional shift event172ofFIG. 4B. Step220forms part of the second shift pattern selected by step212ofFIG. 5. Therefore, EVT2is the intermediate mode for the single-staged multiple-shift maneuver160and EVT1is the target (or final) mode. Note however, that the eTC method300may be used for other mode-to-mode shifts.

The eTC method300is described herein with respect to the structure illustrated inFIG. 1. However, those having ordinary skill in the art will recognize that the present invention may also be incorporated into other powertrain arrangements without departing from the intended scope of the claimed invention.

The exemplary EVT2-to-EVT1shift begins with clutches C4and C1engaged and finishes with clutches C3and C1engaged (as shown inFIG. 2). Therefore, C4is the offgoing clutch and C3is the oncoming clutch. The eTC method300begins at step302as the eTC transitional shift event is initiated and the hybrid control module (or other portion of the control architecture of powertrain10) determines the offgoing clutch C4release point in step304, which may coincide with the slip speed of oncoming clutch C3reaching zero.

The eTC method300may optionally begin pre-filling (not shown) the oncoming clutch associated with the target EVT mode to a predetermined pre-fill level. The clutch volume for the oncoming clutch mechanism, C3, can be filled to 80-90% without reaching torque capacity or causing an inordinate amount of slip, which may otherwise interfere with the current operating mode. This pre-fill strategy may shorten the shift time of the eTC transitional shift event by reducing sequential fill times.

At step306the eTC method300determines whether or not the target release point of the offgoing clutch C4has been reached. Once the release point of C4is reached, the eTC method300proceeds to step308. Otherwise, eTC method300returns to step304to again determine the release point of C4. Optionally (not shown), the eTC method300may include a break or abort command to shift transmission14to another operating mode based upon changing conditions.

At step308, the eTC method300begins torque control of the offgoing clutch C4and unloads C4by exhausting fluid from the clutch piston. Step310verifies that C4has released (i.e., is not carrying torque). If C4has not yet released, eTC method300continues torque control and offloading C4(in step308) until step310verifies that C4is released.

In step312, the method determines Ncand Nc—dot(the first time derivative of Nc) profiles and start speed phase control of oncoming clutch C3. Speed phase control begins to synchronize C3by moving slip speed, Nc, toward zero. Determining Nc—dotwill allow the eTC method300to calculate the time it will take to control the slip speed Ncfrom its current level to zero, and therefore synchronize C3.

The eTC method300commands and monitors C3clutch fill level in step314. If oncoming clutch C3is filled, the eTC method300determines whether C3is synchronized in step316. If C3is not filled or synchronized, the eTC method300returns to speed phase control in step312until C3is filled and synchronized.

After oncoming clutch C3is synchronized, full pressure is applied to C3in step318. Applying full pressure to the piston of C3acts to lock C3and begin torque transfer thereacross. Once C3is locked, the eTC method300may also ramp up reactive torque limits across C3. Locking C3places the transmission14in EVT1mode, completing the eTC transitional shift event172. The eTC method300is then complete and exits, either to end the multiple-shift event (step222inFIG. 5) or to proceed to the QA transitional shift event (step216inFIG. 5).

Note that the synchronous disengagement of offgoing clutch C4followed by the synchronous engagement of oncoming clutch C3both occur while the speed of motor A, NA, is at or near zero. Near-zero NAallows transmission14to balance battery power against power output, even while increasing or decreasing input speed Ni. Furthermore, the transmission14may continuously produce an output torque when transitioning from EVT2mode to EVT1mode.

With reference now to the flow chart shown inFIG. 7, and with continued reference toFIGS. 1-6, there is shown a QA method400of executing QA transitional shift event. The QA method400is a more-detailed process showing the control method for either of steps216or218in the method200ofFIG. 5. Asynchronous shifting may be characterized in that the relative slip across the at least one of the offgoing and oncoming clutches is not equal to zero throughout the entire shift operation. Furthermore, at least one of the oncoming and offgoing clutches is carrying torque while being controllably slipped during the shift operation.

For illustrative purposes, the QA method400will be described in relation to an EVT3-to-EVT2downshift, generally corresponding to step218ofFIG. 5and transitional shift event170ofFIG. 4B. Step218forms part of the second shift pattern selected by step212ofFIG. 5. Therefore, EVT3is the initial mode for the single-staged multiple-shift maneuver160and EVT2is the intermediate mode. Note however, that the QA method400may be used for other mode-to-mode shifts.

The QA method400is described herein with respect to the structure illustrated inFIG. 1. However, those having ordinary skill in the art will recognize that the present invention may also be incorporated into other powertrain arrangements without departing from the intended scope of the claimed invention.

The QA method400begins at step401as the QA transitional shift event is initiated and the hybrid control module (or other portion of the control architecture of powertrain10) determines input speed and acceleration profiles at step403. The input speed and acceleration profiles may generally be derived from the single-staged input speed profile162(shown inFIG. 4B) created in step210(shown inFIG. 5). Based on the initial and target speeds of the oncoming clutch, as well as certain calibrations that characterize the “deflection points” of the profile.

Prior to, contemporaneous with, or subsequent to steps401and403, the controller will determine, at step405, if the QA transitional shift event is preferably completed using the oncoming clutch C1or the offgoing clutch C2. There are two ways to perform the quasi-asynchronous shift: using the oncoming clutch C1or the offgoing clutch C1. Each option has certain advantages and disadvantages. Optionally, the controller may command fluid be distributed to the oncoming clutch C1, pre-filling C1to a predetermined pre-fill level, which is less than the level necessary to achieve full torque capacity.

If decision step405determines that the QA transitional shift event will be completed using the oncoming clutch C1, then step409will determine if an over/under ratio is needed. Over/under ratio is where the engine speed is increased or decreased to over or under the target gear ratio, respectively. In one example of an under ratio, if the target gear ratio is 1.00 and the output speed is 1000 rpm, the target engine speed would be 1000 rpm. The engine speed is brought from 2000 to 800 rpm on an upshift and therefore moves under the target ratio. In the same example, if the input speed is brought from 800 to 1200 rpm for the upshift, it is considered an over ratio.

The over/under ratio may be needed, for example, if the shift operation is a power-on downshift using the oncoming clutch C1. In contrast, under a power-on case (i.e., positive output torque and no regenerative braking) in which the oncoming clutch C1is used for completing the upshift, no under ratio is needed. The over ratio is needed in the power-on downshift case because oncoming clutch C1will only produce positive output torque during the under-ratio period. If the offgoing clutch C2is used to complete the QA transitional shift event, then the case is just the opposite.

Performing a clutch-to-clutch shift may occur in two distinct phases: a torque phase and a speed phase. The torque phase refers to the time during which the offgoing clutch is unloaded and released. The speed phase refers to the period during which the certain member speed of one of the members is controlled from the old speed target toward the new speed target, with the use of the oncoming or offgoing clutch and other torque generative devices, such as engine12or motors A or B. Speed phase may include control of either or both of the input speed, Ni, and clutch slip speed, Nc, profiles.

Once it is determined whether the over/under ratio is needed, the shift operation enters into the torque phase, and the QA method400will then determine the slip speed and acceleration profiles of the oncoming clutch C1in either step411or step413. If an over/under ratio is needed, the oncoming clutch C1slip speed and acceleration profiles are determined based, in part, on the over/under ratio, as indicated in step411. For example, the oncoming clutch slip speed profile is first targeted to a slip speed equivalent of the over/under ratio speed, and thereafter retargets to zero to engage the oncoming clutch when clutch torque is to be exchanged between the oncoming and offgoing clutch. However, if the over/under ratio is not needed, the QA method400proceeds to step413, and the oncoming clutch slip and acceleration profiles are determined without the over/under ratio.

In step415, the torque-generative devices of powertrain10, which include the engine12and the motors A and B, are used to control the input speed. In steps417and419, the system controller will determine if the slip sign of the oncoming clutch C1is correct. Step419will determine whether C1is filled, and therefore ready for full engagement and full torque transfer. The slip sign of C1is deemed correct if the sign of the output torque command is the same as the oncoming clutch slip sign.

Clutch torque and clutch slip always have the same sign (i.e., either both positive or both negative), and the transfer function between clutch torque and output torque has a fixed relationship. Therefore, if the output torque command is positive, and the transfer function has a positive coefficient, a positive clutch torque will produce positive output torque. If steps417or419return a negative signal—i.e., the oncoming clutch slip sign is incorrect or the apply chamber of C1is not filled, the system will revert back to step415and adjust input speed until steps417and419return positive signals.

If the oncoming clutch C1is filled, and the controller can ascertain that the slip sign is correct, step421of the QA method400will calculate and apply the oncoming clutch (C1) torque based on the output torque command, and then complete the torque phase of the shift sequence by exhausting the offgoing clutch. Because the transfer function between oncoming clutch (C1) torque to output torque is fixed, once output torque command is known, the torque for oncoming clutch C1can be determined.

The QA method400then transitions into the speed phase. In step423, the powertrain torque generative devices are used to control the transmission input speed, Ni, and the speed of oncoming clutch C1by adjusting torque input from the engine12, motor A, motor B, or combinations thereof. In step425, the system will monitor, sense, or otherwise determine if the amount of slip of oncoming clutch C1is less than a predetermined slip threshold. If the oncoming clutch slip is less than the slip threshold, the controller will respond by locking C1at step427, and then exiting the QA method400.

Torque exchange between the oncoming clutch C1and offgoing clutch C2occurs as the torque capacity of oncoming clutch C1is increased by increasing hydraulic pressure thereto. Torque contributions of Motors A and B are reduced as the magnitude of reactive torque of the oncoming clutch C1increases, and the system transitions the target EVT mode.

The QA transitional shift events described may be characterized as a quasi-asynchronous shift because the slip speed across at least one of the oncoming and offgoing clutches is non-zero throughout substantially the entire transitional shift event. Furthermore, at least one of the oncoming and offgoing clutches is carrying torque while being controllably slipped during the shift operation. As such, the input speed on line162ofFIG. 4Bwill not dwell on any of the fixed gear ratios, such as lines102and104. Ideally, the slip speed across the oncoming and offgoing clutches is controlled by adjusting input torque from either the engine12or one of the motors A and B.

If it is determined in step405that the QA transitional shift event is will use the offgoing clutch C2to complete the shift, the QA method400proceeds to step429. When using the offgoing clutch C2, the speed phase occurs first and the oncoming clutch C1comes on during the torque phase. Accordingly, the offgoing clutch C2must be controllably slipped prior to synchronizing the oncoming clutch C1, as indicated at429. Optionally, the oncoming clutch C1may be pre-filled to a predetermined pre-fill level, which is less than that to achieve full torque capacity or clutch slippage. Concurrently therewith, the offgoing clutch C2may be slipped by controlling the torque capacity of C2to below the reactive torque.

Offgoing clutch C2is controllably slipped by reducing torque capacity by reducing hydraulic pressure in the clutch fill-chamber of C2. As torque capacity is reduced, it becomes equal to the magnitude of reactive torque of clutch C2. Reactive torque is commonly defined as a magnitude of torque transmitted through a torque-transfer device. Torque capacity is commonly defined as a maximum amount of torque transmissible across a clutch, and is generally based upon the magnitude of clutch pressure and clutch friction. When the magnitude of clutch torque exceeds the torque capacity, clutch slip occurs. The reactive torque is always less than or equal to the torque capacity. Clutch pressure is created by controlling the magnitude of hydraulic pressure applied to the clutch by the hydraulic circuit of the transmission.

In order to reduce reactive torque to the offgoing clutch C2without adversely affecting torque output of the powertrain10, the controller transfers sufficient energy to motors A and B such that their outputs torque are equal to the decrease in reactive torque across the offgoing clutch C2, and thus able to maintain the output torque at shaft20of the transmission14. The torque output of motors A and B may then be concurrently increased with continued decrease in torque capacity and reactive torque of offgoing clutch C2. When C2starts to slip, the reactive torque is the same as clutch capacity torque, and the output torque has a fixed ratio to the reactive torque, or capacity torque of the off going clutch C2.

Once the offgoing clutch C2is slipped, the QA method400will enter the speed phase. In step431, similar to step413described above, the slip speed and acceleration profiles of the oncoming clutch C1are determined. Ideally, if the shift operation is completed using the offgoing clutch, the transmission input speed, Ni, and slip speed of oncoming clutch C1are modified prior to exhausting the offgoing clutch C2, as indicated in step433. This is in contrast to the order presented in steps409-427, using the oncoming clutch C1to complete the shift, in which the transmission input speed and oncoming clutch speed are modified after exhausting the offgoing clutch C2.

In step435, the QA method400calculates and applies the offgoing clutch torque based on the output torque command, in a manner similar to that described above with respect to step421. In contrast to step421, however, step435completes the speed phase, and therefore does not include exhausting the offgoing clutch C2. Subsequently, the shift sequence enters into the torque phase. In steps437and439, the QA method400will determine if the slip sign oncoming clutch C1is correct and whether the oncoming clutch C1is filled. If so, the QA method400will lock the oncoming clutch C1and exhaust the offgoing clutch C2in step441. and thereafter exit the shift sequence.

Referring again toFIGS. 4B and 5, completion of both the QA transitional shift event170and the eTC transitional shift event172completes the single-staged multiple-shift maneuver160. The transmission14has been shifted from the initial EVT mode (EVT3) to the intermediate EVT mode (EVT2) with the QA transitional shift event170. The transmission14was then shifted from the intermediate EVT mode (EVT2) to the target EVT mode (EVT1) with the eTC transitional shift event172. Furthermore, these two shifts have occurred along a single-staged input profile162and the vehicle driver and occupants may not have sensed more than a single shift.

While the best modes and other modes for carrying out the present invention have been described in detail, those familiar with the art to which this invention pertains will recognize various alternative designs and embodiments for practicing the invention within the scope of the appended claims.