US20020111731A1 - Automatic transmission shift control - Google Patents
Automatic transmission shift control Download PDFInfo
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- US20020111731A1 US20020111731A1 US09/783,116 US78311601A US2002111731A1 US 20020111731 A1 US20020111731 A1 US 20020111731A1 US 78311601 A US78311601 A US 78311601A US 2002111731 A1 US2002111731 A1 US 2002111731A1
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- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W10/00—Conjoint control of vehicle sub-units of different type or different function
- B60W10/04—Conjoint control of vehicle sub-units of different type or different function including control of propulsion units
- B60W10/06—Conjoint control of vehicle sub-units of different type or different function including control of propulsion units including control of combustion engines
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W10/00—Conjoint control of vehicle sub-units of different type or different function
- B60W10/04—Conjoint control of vehicle sub-units of different type or different function including control of propulsion units
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W10/00—Conjoint control of vehicle sub-units of different type or different function
- B60W10/10—Conjoint control of vehicle sub-units of different type or different function including control of change-speed gearings
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W10/00—Conjoint control of vehicle sub-units of different type or different function
- B60W10/10—Conjoint control of vehicle sub-units of different type or different function including control of change-speed gearings
- B60W10/11—Stepped gearings
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W30/00—Purposes of road vehicle drive control systems not related to the control of a particular sub-unit, e.g. of systems using conjoint control of vehicle sub-units, or advanced driver assistance systems for ensuring comfort, stability and safety or drive control systems for propelling or retarding the vehicle
- B60W30/18—Propelling the vehicle
- B60W30/1819—Propulsion control with control means using analogue circuits, relays or mechanical links
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W30/00—Purposes of road vehicle drive control systems not related to the control of a particular sub-unit, e.g. of systems using conjoint control of vehicle sub-units, or advanced driver assistance systems for ensuring comfort, stability and safety or drive control systems for propelling or retarding the vehicle
- B60W30/18—Propelling the vehicle
- B60W30/19—Improvement of gear change, e.g. by synchronisation or smoothing gear shift
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W2510/00—Input parameters relating to a particular sub-units
- B60W2510/06—Combustion engines, Gas turbines
- B60W2510/0638—Engine speed
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W2510/00—Input parameters relating to a particular sub-units
- B60W2510/10—Change speed gearings
- B60W2510/1015—Input shaft speed, e.g. turbine speed
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W2540/00—Input parameters relating to occupants
- B60W2540/10—Accelerator pedal position
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W2710/00—Output or target parameters relating to a particular sub-units
- B60W2710/06—Combustion engines, Gas turbines
- B60W2710/0605—Throttle position
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W2710/00—Output or target parameters relating to a particular sub-units
- B60W2710/06—Combustion engines, Gas turbines
- B60W2710/0616—Position of fuel or air injector
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
- B60W2710/00—Output or target parameters relating to a particular sub-units
- B60W2710/06—Combustion engines, Gas turbines
- B60W2710/0666—Engine torque
Definitions
- the invention relates to a system and method for controlling ratio changes for a multiple-ratio transmission in a vehicle driveline having an engine with an electronic throttle control.
- Automatic transmissions used in contemporary vehicle powertrains include multiple-ratio gearing wherein the torque flow paths through the gearing elements and the relative speeds of the gearing elements are controlled by fluid pressure-operated friction elements; i.e., friction clutches and brakes.
- the management of the torque transfer through the gearing from the throttle-controlled engine to the traction wheels is achieved by a control system having an electronic controller that responds to powertrain variables, including engine variables and driver commands.
- the control system develops clutch and brake control pressure, which is determined by solenoid valves under the control of the controller.
- a fluid pressure circuit would include a solenoid valve dedicated to circuit pressure control, and a separate solenoid valve circuit to effect control of the pressure-operated clutches and brakes during upshifts and downshifts.
- the control strategy for the friction elements involves a direct relationship between the transmission output torque control and the ratio change time control. This may require a compromise in shift quality because optimum transmission output torque control is not necessarily consistent with optimum ratio shift timing control.
- the control strategy of known transmission systems for controlling friction element capacity must be changed rapidly in a controlled fashion to match varying engine torque.
- the friction elements may not have a response to control commands that is fast enough to accomplish a smooth ratio change, which results in lower shift quality.
- variations in shift quality on a shift-to-shift basis may occur.
- Another related copending application which also is assigned to the assignee of the present invention, is Ser. No. 09/665,353, filed Sep. 18, 2000. It discloses a multiple-ratio gear system in which independent pressure control of separate friction elements is achieved using separate variable-force solenoids.
- a microprocessor receives input continuously from driveline sensors and stores and executes a control logic for controlling the oncoming friction element during a ratio change. Information that is learned from the input signals is executed on a real-time basis to calculate clutch pressures to achieve optimum shift quality.
- the slip time during a ratio change is controlled by controlling friction element pressure. That is, the controller controls the solenoid valves that determine friction clutch pressure.
- the control system and method of the invention uses a closed-loop controller to achieve the desired ratio rate during a shift from one gear ratio to the next.
- the ratio rate is a function of accelerator pedal position.
- the transmission output torque is controlled using a friction element capacity control strategy.
- the friction element capacity control involves a determination of the desired wheel torque using a torque feed-forward term. This provides a quick response to changes in transmission input torque.
- the accelerator pedal in the system of the invention is decoupled from the engine throttle. This is achieved by using an electronic throttle control, under the control of the vehicle operator, which is not mechanically connected directly to the engine throttle. Control logic is used to interpret a driver demand for torque by sensing the movement of the accelerator pedal. Controlling the input torque to a manageable level results in a smooth torque transition during a ratio change and reduces variation in shift quality from one shift to the next.
- Engine torque is controlled using the system and method of the invention by using a closed-loop controller to achieve a desired ratio change rate, as mentioned above.
- the controller establishes a commanded ratio rate during a ratio change interval and compares it to the actual ratio rate to detect an error.
- a commanded engine torque then is established, which reduces the error.
- Torque capacity control of the friction elements is an open-loop control in which the desired wheel torque is determined by the controller as speed inputs are received by the controller from an engine speed sensor, a turbine speed sensor and an output shaft speed sensor.
- the controller establishes a desired vehicle wheel torque as a function of accelerator pedal position. It calculates friction element pressure to achieve optimum friction element capacity during a ratio change interval.
- a transmission that would embody the control logic of the invention may include a hydrokinetic torque converter in which the impeller is driven by the engine.
- the converter turbine would be coupled directly to the output shaft through multiple-ratio gearing.
- FIG. 1 is a schematic representation of a transmission having a torque converter and multiple-ratio gearing, the gearing elements being controlled by friction clutches and brakes;
- FIG. 2 is a chart that shows the clutch and brake engagement-and-release pattern for the transmission, schematically illustrated in FIG. 1, as the various multiple, forward-driving ratios and the reverse ratio are established;
- FIG. 3 is a schematic block diagram of the control system together with the signal flow paths between the controller and the transmission and between the controller and the engine;
- FIG. 4 is a schematic illustration of table values for desired torque stored in a ROM memory portion of the controller, the table values being determined by vehicle speed and pedal position;
- FIG. 5 is a plot of wheel torque versus time, which shows the variations in wheel torque at each of the multiple stages of a shift interval
- FIG. 6 is a plot of the commanded clutch pressure versus time during each stage of the ratio change
- FIG. 7 is a plot of the commanded engine torque versus time for each stage of the shift
- FIG. 8 is a plot showing variations in turbine speed and ratio versus time for each stage of the ratio change
- FIG. 9 is a plot of wheel torque versus time during a part throttle shift
- FIG. 10 is a plot of the engine torque for part throttle during each stage of the ratio shift.
- FIG. 11 is a plot of wheel torque, speed ratio and commanded clutch pressure during an advanced throttle shift for each of the stages of the ratio change.
- FIG. 1 shows in schematic form a multiple-ratio transmission for an automotive vehicle capable of embodying the improved control method and strategy of the invention.
- the transmission of FIG. 1 is merely one example of a multiple-ratio transmission that can utilize the control logic of the present invention.
- the invention is not restricted to the specific transmission illustrated in FIG. 1.
- Numeral 10 in FIG. 1 designates a hydrokinetic torque converter that has an impeller 12 and a turbine 14 arranged in known fashion in a toroidal fluid flow circuit.
- the impeller 14 is connected to the crankshaft of an internal combustion engine, as shown at 16 .
- Turbine 14 is connected to turbine shaft 18 , which delivers torque to the carrier of an overdrive simple planetary gear unit 20 .
- Sun gear 22 of the gear unit 20 is connected to overdrive brake drum 24 .
- An overdrive disc brake 26 distributes reaction torque to the transmission housing when it is applied, thereby effecting an overdrive condition for the gear unit 20 .
- a coast clutch 26 directly connects the carrier for the gear unit 20 to the sun gear 22 , thereby accommodating a reverse torque transfer from the vehicle wheels to the engine through the converter.
- An overrunning coupling 28 establishes a direct driving connection between the turbine shaft and torque transfer shaft 30 .
- Shaft 30 serves as a torque input shaft for compound planetary gearing 32 , which comprises gear units 34 and 36 with a common sun gear 38 .
- Ring gear 40 of gear unit 34 is connected to shaft 30 through a forward-drive clutch 42 .
- Sun gear 38 is connected to shaft 38 through a high-ratio clutch 44 .
- Sun gear 38 can be braked by an intermediate speed ratio brake 46 when the clutch 44 is disengaged.
- the carrier 48 for the gear unit 34 is connected to torque output shaft 50 .
- Ring gear 52 for the gear unit 36 is connected to shaft 50 .
- the carrier 54 for the gear unit 36 is connected to reverse brake drum 56 .
- Reverse brake band 58 surrounds drum 56 and anchors the carrier 54 during reverse drive. During forward drive in the lowest speed ratio, brake drum 56 is anchored to the transmission housing through overrunning coupling 60 .
- FIG. 2 is a chart that shows the clutch and brake engagement-and-release pattern for each of the forward driving ratios and the reverse ratio.
- the symbol X represents an engaged friction element.
- the symbol C represents a coasting condition for the coast clutch and the reverse brake.
- Symbol O/R represents an overrunning condition of the overrunning coupling.
- FIG. 1 The schematic diagram of FIG. 1 represents the condition of the friction elements when the transmission is conditioned for overdrive fifth ratio operation. At that time, the clutch 42 and the clutch 44 are engaged, and the overdrive brake 26 is applied.
- a microcontroller for the powertrain includes a calculation block 62 , a transmission control block 64 , an engine control block 66 , a transmission control model 68 , and a clutch and solenoid model 70 .
- the engine is represented by control block 72
- the multiple-ratio transmission is represented by control block 74 .
- the transmission includes a turbine speed sensor 76 and an output shaft speed sensor 78 .
- An engine speed sensor is shown at 80 .
- Feedback control signal flow paths extend from the engine speed sensor 80 , the turbine speed sensor 76 , and the output speed sensor 78 through signal flow paths 82 , 84 and 86 , respectively.
- the signal flow paths extend to the calculation block where the signals are distributed to an input signal conditioning portion of the calculation block 62 through sampling switch 88 for the engine speed signal, switch 90 for the turbine speed signal, and switch 92 for the output shaft speed signal.
- the control system includes a driver-controlled accelerator pedal, which is decoupled from the engine throttle; that is, there is not a direct mechanical connection between the accelerator pedal and the engine throttle.
- a pedal position signal from sensor 94 is distributed to the calculation block 62 .
- a transmission control lever under the control of the operator selects the transmission range, an appropriate transmission lever position signal being distributed to the calculation block as shown at 96 .
- a transmission oil temperature signal is distributed to the calculation block 62 , as shown at 98 .
- the calculation block receives the speed signals from the engine speed sensor, the turbine speed sensor and the output shaft speed sensor and calculates an actual ratio rate during a shift interval, as shown at 100 .
- a calculated engine torque is distributed to the calculation block a shown at 102 .
- the calculation block determines a commanded ratio rate, as shown at 104 , as a function of the pedal position at 94 and the computed engine torque at 102 .
- the commanded ratio rate at 104 is compared to the actual ratio rate 100 at the summing point 106 . This determines an error in the ratio rate, as shown at 108 .
- the error is transmitted to the transmission control block 64 , which determines a commanded engine torque signal at 110 .
- the engine control block 66 receives the commanded engine torque signal and develops control variables for the engine 72 . These variables include a commanded throttle position signal, as shown at 112 . It also determines the calculated engine torque signal 102 , which, as mentioned previously, is received by the calculation block 62 .
- the engine control variables determined by the engine control block 66 include a spark timing control signal 114 and a fuel injection rate signal 116 .
- the spark signal alone or the fuel signal alone could be used.
- the signals at 114 and 116 , together with the commanded throttle position signal at 112 control engine torque.
- the calculation block 62 calculates also a desired torque at the wheels, as shown at 118 . This is distributed to the transmission control model 68 , which develops a value for clutch (friction element) torque based upon the desired torque at the wheels, as shown at 120 .
- the torque signal at 120 is used by the clutch and solenoid model to the develop the variable-force solenoid pressure at 70 .
- the output signal developed at block 70 is a variable-force solenoid signal representing the commanded pressure for the clutches (friction element), as indicated at 122 .
- the pressure-operated friction elements of the transmission 74 are engaged by the variable-force solenoid pressure determined at block 70 .
- each of the feedback signals from the transmission and from the engine are speed signals used by the calculation blocks of the respective control systems to develop an actual ratio rate that can be compared to a desired ratio rate.
- the shift control logic of the present invention consists of four distinct modes. These are the shift start mode, the ratio change mode, the torque ramp mode and shift end mode. The same logic applies to both upshifts and downshifts, although there are some differences.
- the calibration involved in the shifts are unique for each shift.
- control logic is executed and the output control signals transmitted to the solenoids that control the friction elements of the transmission are sensed in real time. This is done in response to the engine, transmission and output shaft speed signals previously discussed, as well as to the pedal position, range selector position and oil temperature signals.
- the desired static torque capacity of the friction element is computed based on input torque and inertia torque.
- a dynamic pressure term which is a function of pedal position and vehicle speed, is added to the desired static pressure to produce a commanded pressure (pr_cmd).
- the commanded clutch pressure is illustrated at 125 in FIG. 6.
- Commanded clutch (friction element) pressure (pr_cmd) is computed as follows:
- the symbol PP is the pedal position value
- the symbol V s is the vehicle speed value
- the symbol tq_trans represents transmission torque
- the symbol tq_i ⁇ represents inertia torque
- Symbol MLTQ is the ratio of clutch torque to input torque
- the symbol GAIN is gain of the oncoming clutch.
- the symbol PRST is the stroke pressure.
- FNMLTE is a temperature compensation multiplier developed during calibration.
- the control strategy for the ratio change mode involves controlling the pressure of the oncoming friction element to obtain the desired wheel torque.
- a torque feed-forward term is used to provide a quick response to changes in the transmission input torque.
- the oncoming friction element pressure is computed in equations (4) as follows:
- tw — dl — des filter [ ⁇ ( FNTWDL ( pcsftcmpt ), pps — rel ⁇ +*FNFRRBCL (loop_counter)]
- tw — des filter [ ⁇ tw — strt +( tw — dl — des+tw — dl — ff _)* MLTW *temperature — MULT ⁇ ]
- pr_dl_acm incremental accumulator pressure
- pr_cmd_sol solenoid commanded pressure
- FRLOSS fraction of total losses between the friction element and the trans. output
- MLTQCL percent of turbine torque carried by the friction element (computed from diagram of FIG. 1)
- FNMLPP desensitizing function to minimize small pedal position movement
- FNTWDL desired wheel torque adder
- FRTW desired wheel torque fractional multiplier
- FNFRRBCL multiplier for shaping at the start of ratio change
- MUPRVF slope factor for conversion to solenoid pressure
- INPRVF intercept factor for conversion to solenoid pressure
- tw_des_tbl is a table value that is obtained from a table stored in ROM, as indicated in FIG. 4.
- the desired torque stored in ROM is determined by the pedal position signal pps_rel and the vehicle speed signal V s .
- the table of FIG. 4 is a “lookup” table.
- the ratio rate is controlled using engine throttle in a closed-loop fashion. This achieves a desired ratio rate when changing from a previous gear ratio to the next gear ratio.
- the ratio rate is a function of pedal position, as mentioned earlier.
- the starting value of desired engine torque is calculated as follows:
- tq_net is the calculated engine torque shown at 102 in FIG. 3.
- a commanded torque is obtained as shown at point 124 in FIG. 7.
- Control logic that is common to all shifts, both upshifts and downshifts, is as follows:
- MLLSTE fractional losses between friction element and the transmission input
- FNFRRBTE shaping function at the st art of ratio change
- Kc PID controller overall gain
- Ki PID controller integral gain
- Kd PID controller derivative gain
- te_cmd is the commanded engine torque that is achieved by adjusting spark timing or fuel.
- throttle position and either spark timing or fuel control can be used. In this way, the torque that is asked for is achieved.
- temperature_mult is a multiplier determined by calibration to take into account changes in transmission temperature.
- the desired engine torque is calculated by interpolating between the current desired engine torque and the desired engine torque at the end of the shift. This is illustrated by the following equations:
- tw — des _end tw — strt +filter ⁇ FNTWDS ( V s , pps — rel ) ⁇ tw — strt ⁇ (8)
- te — des _end ⁇ tq _fead+ tq _pump+[ tw — des _end/( rt — gr _(next/prev)* fdr )+ tq _spin ⁇ tq — i ⁇ ] ⁇ tq — i ⁇ ]/[TR *(1 ⁇ tq _prop)] ⁇ * MLLSTE
- te — cmd [( rt _trans ⁇ rt _trans — 1 st )/( rt — fin ⁇ rt _trans — 1 st )*( te — des _end ⁇ te — cmd )]+ te — cmd
- te_cmd_mid desired engine torque at the start of torque ramp mode
- FNTWDS desired wheel torque
- the routine will exit the torque ramp mode and enter the shift end mode if the maximum allowed time in the torque ramp mode has expired, or if the transmission ratio is less than the target value for an upshift or greater than the target value for a downshift.
- the capacity of the controlling friction element is changed rapidly in a controlled fashion so that the clutch capacity matches the varying engine torque. Shift variations on a shift-to-shift basis are reduced. Changes in pedal position during a shift are filtered and the signals are shaped to provide a smooth and manageable change in input torque consistent with the response characteristics of the controlling friction element.
- the input torque during a speed change phase can be increased in accordance with the strategy of the invention, thereby partially compensating for the output torque dip caused by inertia effects.
- This is illustrated for part-throttle operation in FIG. 9 where the inertia effect on wheel torque for a control system that does not include an engine throttle position commanded signal is shown at 126 .
- the dotted line plot of FIG. 9 shows the reduction in the output torque dip caused by inertia effects. This is illustrated at 128 .
- This result is achieved by reason of the fact that the additional variable, i.e., commanded throttle position, spark timing and fuel control, unlike other systems where spark timing and fuel only are available for controlling engine torque. Improved authority over the engine results in a less severe dip in wheel torque compared to the plot 126 shown in FIGS. 5 and 9.
- FIG. 10 shows a plot of engine torque versus time during a shift interval, which results in the reduction in the torque dip seen in FIG. 9.
- Torque for part throttle is shown at 130 and the corresponding curve for a system that does not employ a commanded throttle pressure term is shown at 132 .
- FIG. 8 is a plot of the transmission ratio during a ratio shift interval, as shown at 138 .
- the turbine speed during the shift interval for each of the shift modes is shown at 140 .
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Abstract
Description
- 1. Field of the Invention
- The invention relates to a system and method for controlling ratio changes for a multiple-ratio transmission in a vehicle driveline having an engine with an electronic throttle control.
- 2. Background Art
- Automatic transmissions used in contemporary vehicle powertrains include multiple-ratio gearing wherein the torque flow paths through the gearing elements and the relative speeds of the gearing elements are controlled by fluid pressure-operated friction elements; i.e., friction clutches and brakes. The management of the torque transfer through the gearing from the throttle-controlled engine to the traction wheels is achieved by a control system having an electronic controller that responds to powertrain variables, including engine variables and driver commands.
- The control system develops clutch and brake control pressure, which is determined by solenoid valves under the control of the controller. Typically, a fluid pressure circuit would include a solenoid valve dedicated to circuit pressure control, and a separate solenoid valve circuit to effect control of the pressure-operated clutches and brakes during upshifts and downshifts.
- In the transmission disclosed in copending application Ser. No. 09/636,729, filed Aug. 10, 2000, a separate solenoid valve is provided for controlling each ratio shift, the offgoing friction element being activated independently of the ongoing friction element during a ratio change.
- Copending patent application Ser. No. 09/366,416, filed Aug. 4, 1999, describes engagement control logic wherein the ratio changes occur with an adaptive engagement feel. The controller for this transmission control provides consistent, smooth engagement of a friction element during vehicle startup. It includes a controller with an adaptive strategy that maintains consistent shift quality and takes into account system variations, such as changes in coefficient of friction, friction element wear, etc. Actual slip rate during engagement of the friction element is maintained at a desired value.
- Each of these copending patent applications is assigned to the assignee of the present invention.
- In prior art systems corresponding to those described in the copending patent applications, the control strategy for the friction elements involves a direct relationship between the transmission output torque control and the ratio change time control. This may require a compromise in shift quality because optimum transmission output torque control is not necessarily consistent with optimum ratio shift timing control. The control strategy of known transmission systems for controlling friction element capacity must be changed rapidly in a controlled fashion to match varying engine torque. The friction elements may not have a response to control commands that is fast enough to accomplish a smooth ratio change, which results in lower shift quality. Furthermore, variations in shift quality on a shift-to-shift basis may occur.
- Another related copending application, which also is assigned to the assignee of the present invention, is Ser. No. 09/665,353, filed Sep. 18, 2000. It discloses a multiple-ratio gear system in which independent pressure control of separate friction elements is achieved using separate variable-force solenoids. A microprocessor receives input continuously from driveline sensors and stores and executes a control logic for controlling the oncoming friction element during a ratio change. Information that is learned from the input signals is executed on a real-time basis to calculate clutch pressures to achieve optimum shift quality.
- The disclosures of the copending patent applications are incorporated in the present disclosure by reference to supplement the present disclosure.
- In each of the systems disclosed in the copending patent applications, the slip time during a ratio change is controlled by controlling friction element pressure. That is, the controller controls the solenoid valves that determine friction clutch pressure.
- The control system and method of the invention uses a closed-loop controller to achieve the desired ratio rate during a shift from one gear ratio to the next. The ratio rate is a function of accelerator pedal position. Concurrently, the transmission output torque is controlled using a friction element capacity control strategy. The friction element capacity control involves a determination of the desired wheel torque using a torque feed-forward term. This provides a quick response to changes in transmission input torque.
- The accelerator pedal in the system of the invention is decoupled from the engine throttle. This is achieved by using an electronic throttle control, under the control of the vehicle operator, which is not mechanically connected directly to the engine throttle. Control logic is used to interpret a driver demand for torque by sensing the movement of the accelerator pedal. Controlling the input torque to a manageable level results in a smooth torque transition during a ratio change and reduces variation in shift quality from one shift to the next.
- Engine torque is controlled using the system and method of the invention by using a closed-loop controller to achieve a desired ratio change rate, as mentioned above. The controller establishes a commanded ratio rate during a ratio change interval and compares it to the actual ratio rate to detect an error. A commanded engine torque then is established, which reduces the error.
- Torque capacity control of the friction elements, unlike the closed-loop control of the engine torque, is an open-loop control in which the desired wheel torque is determined by the controller as speed inputs are received by the controller from an engine speed sensor, a turbine speed sensor and an output shaft speed sensor. The controller establishes a desired vehicle wheel torque as a function of accelerator pedal position. It calculates friction element pressure to achieve optimum friction element capacity during a ratio change interval.
- A transmission that would embody the control logic of the invention may include a hydrokinetic torque converter in which the impeller is driven by the engine. The converter turbine would be coupled directly to the output shaft through multiple-ratio gearing.
- FIG. 1 is a schematic representation of a transmission having a torque converter and multiple-ratio gearing, the gearing elements being controlled by friction clutches and brakes;
- FIG. 2 is a chart that shows the clutch and brake engagement-and-release pattern for the transmission, schematically illustrated in FIG. 1, as the various multiple, forward-driving ratios and the reverse ratio are established;
- FIG. 3 is a schematic block diagram of the control system together with the signal flow paths between the controller and the transmission and between the controller and the engine;
- FIG. 4 is a schematic illustration of table values for desired torque stored in a ROM memory portion of the controller, the table values being determined by vehicle speed and pedal position;
- FIG. 5 is a plot of wheel torque versus time, which shows the variations in wheel torque at each of the multiple stages of a shift interval;
- FIG. 6 is a plot of the commanded clutch pressure versus time during each stage of the ratio change;
- FIG. 7 is a plot of the commanded engine torque versus time for each stage of the shift;
- FIG. 8 is a plot showing variations in turbine speed and ratio versus time for each stage of the ratio change;
- FIG. 9 is a plot of wheel torque versus time during a part throttle shift;
- FIG. 10 is a plot of the engine torque for part throttle during each stage of the ratio shift; and
- FIG. 11 is a plot of wheel torque, speed ratio and commanded clutch pressure during an advanced throttle shift for each of the stages of the ratio change.
- FIG. 1 shows in schematic form a multiple-ratio transmission for an automotive vehicle capable of embodying the improved control method and strategy of the invention. The transmission of FIG. 1 is merely one example of a multiple-ratio transmission that can utilize the control logic of the present invention. The invention is not restricted to the specific transmission illustrated in FIG. 1.
-
Numeral 10 in FIG. 1 designates a hydrokinetic torque converter that has animpeller 12 and aturbine 14 arranged in known fashion in a toroidal fluid flow circuit. Theimpeller 14 is connected to the crankshaft of an internal combustion engine, as shown at 16. -
Turbine 14 is connected toturbine shaft 18, which delivers torque to the carrier of an overdrive simpleplanetary gear unit 20.Sun gear 22 of thegear unit 20 is connected tooverdrive brake drum 24. Anoverdrive disc brake 26 distributes reaction torque to the transmission housing when it is applied, thereby effecting an overdrive condition for thegear unit 20. Acoast clutch 26 directly connects the carrier for thegear unit 20 to thesun gear 22, thereby accommodating a reverse torque transfer from the vehicle wheels to the engine through the converter. - An overrunning
coupling 28 establishes a direct driving connection between the turbine shaft andtorque transfer shaft 30. -
Shaft 30 serves as a torque input shaft for compoundplanetary gearing 32, which comprisesgear units common sun gear 38.Ring gear 40 ofgear unit 34 is connected toshaft 30 through a forward-drive clutch 42.Sun gear 38 is connected toshaft 38 through a high-ratio clutch 44.Sun gear 38 can be braked by an intermediatespeed ratio brake 46 when the clutch 44 is disengaged. Thecarrier 48 for thegear unit 34 is connected totorque output shaft 50. -
Ring gear 52 for thegear unit 36 is connected toshaft 50. Thecarrier 54 for thegear unit 36 is connected to reversebrake drum 56.Reverse brake band 58 surroundsdrum 56 and anchors thecarrier 54 during reverse drive. During forward drive in the lowest speed ratio,brake drum 56 is anchored to the transmission housing through overrunningcoupling 60. - FIG. 2 is a chart that shows the clutch and brake engagement-and-release pattern for each of the forward driving ratios and the reverse ratio. The symbol X represents an engaged friction element. The symbol C represents a coasting condition for the coast clutch and the reverse brake. Symbol O/R represents an overrunning condition of the overrunning coupling.
- The schematic diagram of FIG. 1 represents the condition of the friction elements when the transmission is conditioned for overdrive fifth ratio operation. At that time, the clutch42 and the clutch 44 are engaged, and the
overdrive brake 26 is applied. - The overall control system is illustrated in block diagram form in FIG. 3. A microcontroller for the powertrain includes a
calculation block 62, atransmission control block 64, anengine control block 66, atransmission control model 68, and a clutch andsolenoid model 70. The engine is represented bycontrol block 72, and the multiple-ratio transmission is represented bycontrol block 74. The transmission includes aturbine speed sensor 76 and an outputshaft speed sensor 78. An engine speed sensor is shown at 80. - Feedback control signal flow paths extend from the
engine speed sensor 80, theturbine speed sensor 76, and theoutput speed sensor 78 throughsignal flow paths calculation block 62 throughsampling switch 88 for the engine speed signal, switch 90 for the turbine speed signal, and switch 92 for the output shaft speed signal. - The control system includes a driver-controlled accelerator pedal, which is decoupled from the engine throttle; that is, there is not a direct mechanical connection between the accelerator pedal and the engine throttle. A pedal position signal from
sensor 94 is distributed to thecalculation block 62. A transmission control lever under the control of the operator selects the transmission range, an appropriate transmission lever position signal being distributed to the calculation block as shown at 96. A transmission oil temperature signal is distributed to thecalculation block 62, as shown at 98. - The calculation block receives the speed signals from the engine speed sensor, the turbine speed sensor and the output shaft speed sensor and calculates an actual ratio rate during a shift interval, as shown at100. A calculated engine torque is distributed to the calculation block a shown at 102.
- The calculation block determines a commanded ratio rate, as shown at104, as a function of the pedal position at 94 and the computed engine torque at 102. The commanded ratio rate at 104 is compared to the
actual ratio rate 100 at the summingpoint 106. This determines an error in the ratio rate, as shown at 108. The error is transmitted to thetransmission control block 64, which determines a commanded engine torque signal at 110. Theengine control block 66 receives the commanded engine torque signal and develops control variables for theengine 72. These variables include a commanded throttle position signal, as shown at 112. It also determines the calculatedengine torque signal 102, which, as mentioned previously, is received by thecalculation block 62. - The engine control variables determined by the
engine control block 66 include a sparktiming control signal 114 and a fuelinjection rate signal 116. In the alternative, the spark signal alone or the fuel signal alone could be used. In any case, the signals at 114 and 116, together with the commanded throttle position signal at 112, control engine torque. - The
calculation block 62 calculates also a desired torque at the wheels, as shown at 118. This is distributed to thetransmission control model 68, which develops a value for clutch (friction element) torque based upon the desired torque at the wheels, as shown at 120. The torque signal at 120 is used by the clutch and solenoid model to the develop the variable-force solenoid pressure at 70. The output signal developed atblock 70 is a variable-force solenoid signal representing the commanded pressure for the clutches (friction element), as indicated at 122. The pressure-operated friction elements of thetransmission 74 are engaged by the variable-force solenoid pressure determined atblock 70. - The pressure applied to the friction elements of the transmission are determined by the torque desired value produced by the
calculation block 62 in an open-loop fashion. This is in contrast to the closed-loop control of the engine torque. As in the case of the controllers described in the copending patent applications Ser. Nos. 09/665,353, 09/366,416, and 09/636,729, each of the feedback signals from the transmission and from the engine are speed signals used by the calculation blocks of the respective control systems to develop an actual ratio rate that can be compared to a desired ratio rate. - As in the case of the control systems of the copending patent applications previously discussed, the shift control logic of the present invention consists of four distinct modes. These are the shift start mode, the ratio change mode, the torque ramp mode and shift end mode. The same logic applies to both upshifts and downshifts, although there are some differences. The calibration involved in the shifts are unique for each shift.
- During each control loop of the controller, the exit conditions for the current mode are checked prior to entry into the next mode. If the exit conditions for the current mode are satisfied, then the control logic may proceed during the ratio change to the next mode. Otherwise, the control remains with the current mode.
- The control logic is executed and the output control signals transmitted to the solenoids that control the friction elements of the transmission are sensed in real time. This is done in response to the engine, transmission and output shaft speed signals previously discussed, as well as to the pedal position, range selector position and oil temperature signals.
- During the shift start mode, the desired static torque capacity of the friction element is computed based on input torque and inertia torque. A dynamic pressure term, which is a function of pedal position and vehicle speed, is added to the desired static pressure to produce a commanded pressure (pr_cmd). The commanded clutch pressure is illustrated at125 in FIG. 6. Commanded clutch (friction element) pressure (pr_cmd) is computed as follows:
- base clutch press=[(tq_trans+tq — iα)MLTQ/GAIN+PRST−CF*(NT/1000)2]* (1)
- dyn press=FN(PP,V s) (2)
- pr — cmd=(base press+dyn. press)*FNMLT (temp) (3)
- In
equation 2, the symbol PP is the pedal position value, and the symbol Vs is the vehicle speed value. Inequation 1, the symbol tq_trans represents transmission torque, and the symbol tq_iα represents inertia torque. Symbol MLTQ is the ratio of clutch torque to input torque, and the symbol GAIN is gain of the oncoming clutch. The symbol PRST is the stroke pressure. - In
equation 3 above, the term FNMLTE is a temperature compensation multiplier developed during calibration. - The control strategy for the ratio change mode involves controlling the pressure of the oncoming friction element to obtain the desired wheel torque. A torque feed-forward term is used to provide a quick response to changes in the transmission input torque. The oncoming friction element pressure is computed in equations (4) as follows:
- tw — dl — ff=filter {(tw — des — tbl−tw — strt)*FNMLPP(pps — rel)} (4)
- tw — dl — des=filter [{(FNTWDL (pcsftcmpt), pps — rel}+*FNFRRBCL (loop_counter)]
- tw — des=filter [{tw — strt+(tw — dl — des+tw — dl — ff_)*MLTW*temperature— MULT}]
- tq_turb=(net engine torque−tq_pump)*TR
- tq_loss_trans=tq_spin+tq_prop*tq_turbine
- tq — cl={tw — des/overall_ratio+FRLOSS*tq_loss_trans}*MLTQCL−cc*tq_turb
- pr — cmd — cl=tq — cl/gain+stroke_pressure−centrifugal_force_press
- pr — dl — acm=PRDLAC/*incremental accumulator pressure from spring only */
- pr — cmd — sol=pc — cmd — cl*MUPRVF+INPRVF−pr — dl — acm
- where,
- ppx_rel=pedal position
- tw_strt=wheel torque at the start of ratio change mode
- tw_dl_des=additional desired wheel torque during ratio change
- tw_des=total desired wheel torque
- tq_turb=net turbine torque
- tq_loss_trans=transmission torque losses
- tq_pump=pump loss
- tq_spin=transmission spin losses
- tq_prop=transmission torque proportional losses
- tq_cl=friction element torque desired
- pr_cmd_cl=commanded friction element pressure
- pr_dl_acm=incremental accumulator pressure
- pr_cmd_sol=solenoid commanded pressure
- TR=torque converter torque ratio
- FRLOSS=fraction of total losses between the friction element and the trans. output
- MLTQCL=percent of turbine torque carried by the friction element (computed from diagram of FIG. 1)
- FNMLPP=desensitizing function to minimize small pedal position movement affect
- FNTWDL=desired wheel torque adder
- FRTW=desired wheel torque fractional multiplier
- FNFRRBCL=multiplier for shaping at the start of ratio change
- MLTW=fractional multiplier
- MUPRVF=slope factor for conversion to solenoid pressure
- INPRVF=intercept factor for conversion to solenoid pressure
- cc=conversion constant
- In the preceding discussion, the term tw_des_tbl is a table value that is obtained from a table stored in ROM, as indicated in FIG. 4. The desired torque stored in ROM is determined by the pedal position signal pps_rel and the vehicle speed signal Vs. The table of FIG. 4 is a “lookup” table.
- Unlike friction element capacity control, the ratio rate is controlled using engine throttle in a closed-loop fashion. This achieves a desired ratio rate when changing from a previous gear ratio to the next gear ratio.
- The ratio rate is a function of pedal position, as mentioned earlier. For an upshift, the starting value of desired engine torque is calculated as follows:
- te — cmd_base @ 1st loop=tq_net*function(pps — rel, V s) (5)
- In the foregoing equation, the term tq_net is the calculated engine torque shown at102 in FIG. 3. When this value is multiplied by the multiplier in
equation 5, a commanded torque is obtained as shown atpoint 124 in FIG. 7. - For a downshift, the starting value of desired engine torque is calculated as follows:
- te — cmd_base @ 1st loop={tq_fead+tq_pump+[tw — strt/(rt — gr_(prev)*final_drive)+tq_spin−tq — iα]/[TR*(1−tq_prop)]}*MLLSTE (6)
- In the preceding equation (6), the term tq_spin represents spin losses, and the term tq_iα represents inertia torque losses.
- Control logic that is common to all shifts, both upshifts and downshifts, is as follows:
- rtr — cmd=FNFRRC (normalized ratio)/FNTMDSRC (pps — rel) (7)
- err — t0=rtr — cmd−rtr_act (actual ratio rate)
- te — dl — pid={Kp*(err — t0— err — t1)+Kd*(err — t 0 −2*err — t1+err — t2)+Ki*err — t0)}*FNFRRBTE (loop counter)
- te — cmd_base=te — cmd_base+te — dl — pid
- te — dl — ff=tw — dl — ff/[rt — gr_(old/prev)*fdr] (“prev” means gear coming from, “old” is the gear previous)
- te — cmd=(te — cmd_base+te — dl — ff)*temperature— mult
- where:
- tq_fead=front-end accessory drive torque loss
- te_dl_ff=change in desired engine torque due to change-of-mind
- te_cmd=engine torque desired
- rtr_cmd=ratio rate commanded
- fdr=final drive ratio
- MLLSTE=fractional losses between friction element and the transmission input
- FNFRRC=shaping function for ratio change
- FNTMDSRC=ratio change time desired
- FNFRRBTE=shaping function at the st art of ratio change
- Kc=PID controller overall gain
- Kp=PID controller proportional gain
- Ki=PID controller integral gain
- Kd=PID controller derivative gain
- err_t0=current loop error
- err_t1=control error of previous loop
- err_t2=control error of two loops previous
- In the foregoing equations (7), the term te_cmd is the commanded engine torque that is achieved by adjusting spark timing or fuel. Alternatively, throttle position and either spark timing or fuel control can be used. In this way, the torque that is asked for is achieved. The term temperature_mult is a multiplier determined by calibration to take into account changes in transmission temperature.
- During the torque ramp mode shown in FIGS.5-11, the desired engine torque is calculated by interpolating between the current desired engine torque and the desired engine torque at the end of the shift. This is illustrated by the following equations:
- tw — des_end=tw — strt+filter {FNTWDS(V s , pps — rel)−tw — strt} (8)
- te — des_end={tq_fead+tq_pump+[tw — des_end/(rt — gr_(next/prev)*fdr)+tq_spin−tq — iα]−tq — iα]/[TR*(1−tq_prop)]}*MLLSTE
- te — cmd=[(rt_trans−rt_trans—1st)/(rt — fin−rt_trans—1st)*(te — des_end−te — cmd)]+te — cmd
- where:
- te_cmd_mid=desired engine torque at the start of torque ramp mode
- tw_des_end desired wheel torque at the shift end
- te_des_end=desired engine torque at the shift end
- te_cmd=desired engine torque for the current loop
- rt_trans=current loop transmission ratio
- rt_trans—1st=previous loop transmission ratio
- FNTWDS=desired wheel torque
- rt_fin=ratio at end of shift
- The routine will exit the torque ramp mode and enter the shift end mode if the maximum allowed time in the torque ramp mode has expired, or if the transmission ratio is less than the target value for an upshift or greater than the target value for a downshift.
- The routine for the ratio change will enter the shift end mode and the friction element pressure will be ramped up until it reaches the line pressure, as shown in FIG. 6 at125. The effect of this is shown in FIG. 5 where the wheel torque desired is controlled to a value that is a function of pedal position and vehicle speed.
- By employing the foregoing strategy, the capacity of the controlling friction element is changed rapidly in a controlled fashion so that the clutch capacity matches the varying engine torque. Shift variations on a shift-to-shift basis are reduced. Changes in pedal position during a shift are filtered and the signals are shaped to provide a smooth and manageable change in input torque consistent with the response characteristics of the controlling friction element.
- The input torque during a speed change phase can be increased in accordance with the strategy of the invention, thereby partially compensating for the output torque dip caused by inertia effects. This is illustrated for part-throttle operation in FIG. 9 where the inertia effect on wheel torque for a control system that does not include an engine throttle position commanded signal is shown at126. In contrast, the dotted line plot of FIG. 9 shows the reduction in the output torque dip caused by inertia effects. This is illustrated at 128. This result is achieved by reason of the fact that the additional variable, i.e., commanded throttle position, spark timing and fuel control, unlike other systems where spark timing and fuel only are available for controlling engine torque. Improved authority over the engine results in a less severe dip in wheel torque compared to the
plot 126 shown in FIGS. 5 and 9. - FIG. 10 shows a plot of engine torque versus time during a shift interval, which results in the reduction in the torque dip seen in FIG. 9. Torque for part throttle is shown at130 and the corresponding curve for a system that does not employ a commanded throttle pressure term is shown at 132.
- At advanced throttle, the characteristics of wheel torque, the speed ratio and commanded clutch pressure are shown at134, 134′ and 134″. The corresponding plots for a transmission control that does not include a commanded throttle position term are shown at 136, 136′ and 136″.
- FIG. 8 is a plot of the transmission ratio during a ratio shift interval, as shown at138. The turbine speed during the shift interval for each of the shift modes is shown at 140.
- Although an embodiment of the control system and method of the invention has been disclosed, it will be apparent that modifications may be made by persons skilled in the art without departing from the scope of the invention. All such modifications and equivalents thereof are intended to be covered by the following claims.
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