BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a schematic diagram showing an automatic transmission with a shift control apparatus in accordance with embodiments of the present invention.
FIG. 2 is a schematic diagram showing a gear arrangement of an automatic transmission in accordance with the embodiments.
FIG. 3 is a schematic diagram showing a state of the automatic transmission of FIG. 2.
FIG. 4 is a schematic diagram showing another state of the automatic transmission of FIG. 2.
FIG. 5 is a schematic diagram showing an analogous automatic transmission having a simplified construction in relation to a gear shift concerned in the embodiments.
FIGS. 6A, 6B and 6C are time charts showing an example of how an input torque, an input shaft speed, and an output torque of the automatic transmission of FIG. 5 change with time during an upshift.
FIGS. 7A, 7B and 7C are time charts showing another example of how the input torque, input shaft speed, and output torque change with time during an upshift.
FIGS. 8A, 8B and 8C are time charts showing a further example of how the input torque, input shaft speed, and output torque change with time during an upshift.
FIG. 9 is a diagram showing a general relationship in a clutch among input speed, output speed, torque capacity and transmitted torque.
FIG. 10 is a diagram showing how clutch engagement serves to change the input shaft speed in three different states of the relationship among the input shaft speed, the output speed of a first clutch, and the output speed of a second clutch in the automatic transmission of FIG. 5.
FIG. 11 is a block diagram showing a section of the shift control apparatus of FIG. 1 for computing a desired torque capacity correction of each of first and second clutches for torque down in accordance with a first embodiment of the present invention.
FIG. 12 is a block diagram showing a section of the shift control apparatus of FIG. 1 for computing a desired engine-based torque reduction and a desired clutch-based torque reduction for inertia torque cancellation in accordance with the first embodiment.
FIG. 13 is a block diagram showing a section of the shift control apparatus of FIG. 1 for computing a corrected torque capacity of each of the first and second clutches in accordance with the first embodiment.
FIGS. 14A, 14B and 14C are time charts showing an example of how the input torque, input shaft speed and output torque change with time under control during an upshift in accordance with the first embodiment.
FIGS. 15A, 15B and 15C are time charts showing another example of how the input torque, input shaft speed and output torque change with time under control during an upshift in accordance with the first embodiment.
FIG. 16 is a block diagram showing a section of the shift control apparatus of FIG. 1 for computing a target differential speed with a desired dynamic characteristic in accordance with a second embodiment of the present invention.
FIG. 17 is a block diagram showing a section of the shift control apparatus of FIG. 1 for computing a target differential speed with a desired dynamic characteristic in accordance with a modified example of the second embodiment.
FIG. 18 is a block diagram showing a section of the shift control apparatus of FIG. 1 for computing a target differential speed with a desired dynamic characteristic in accordance with another modified example of the second embodiment.
FIGS. 19A, 19B and 19C are time charts showing an example of how the input torque, input shaft speed and output torque change with time under control during an upshift in accordance with the second embodiment.
FIGS. 20A, 20B and 20C are time charts showing another example of how the input torque, input shaft speed and output torque change with time under control during an upshift in accordance with the second embodiment.
FIG. 21 is a block diagram showing a modification of the section of the shift control apparatus of FIG. 12.
FIG. 22 is a block diagram showing another modification of the section of the shift control apparatus of FIG. 12.
DETAILED DESCRIPTION OF THE INVENTION
First, the following describes a principle and a basic construction of an automatic transmission with a shift control apparatus in accordance with embodiments of the present invention with reference to FIGS. 1 to 5. The term “speed” henceforth denotes a rotational speed in general. The automatic transmission includes a plurality of frictional engagement elements (clutches), transmitting a rotation from an engine (external drive unit) to a drive wheel set not shown by engagement of some of the frictional engagement elements. The shift control apparatus controls gear shifts of the automatic transmission.
FIG. 2 schematically shows a gear arrangement of a four-speed automatic transmission. As shown in FIG. 2, the automatic transmission includes an input shaft 11 and an output shaft 12 disposed on first and second axial ends, respectively, and first and second planetary gears 21 and 22 disposed in series between input shaft 11 and output shaft 12. First planetary gear 21 includes a sun gear 21S, a ring gear 21R and a carrier 21C for carrying a planet pinion set in meshed contact with sun gear 21S and ring gear 21R. Second planetary gear 22 includes a sun gear 22S, a ring gear 22R and a carrier 22C for carrying a planet pinion set in meshed contact with sun gear 22S and ring gear 22R.
Disposed between sun gear 21S and a casing 13, a clutch (brake) 23 as a frictional engagement element has an engaged state allowing the sun gear 21S to be held stationary against rotation. Disposed between sun gear 21S and input shaft 11, a clutch 24 as a frictional engagement element has an engaged state allowing the sun gear 21S to rotate with input shaft 11 as a unit. Disposed between carrier 21C and input shaft 11, a clutch 25 has an engaged state allowing the carrier 21C to rotate with input shaft 11 as a unit. Disposed between carrier 21C and casing 13, a clutch (brake) 26 has an engaged state allowing the carrier 21C to be held stationary against rotation. Disposed between carrier 21C and ring gear 22R, a clutch 27 has an engaged state allowing the carrier 21C to rotate with ring gear 22R as a unit. Ring gear 21R is fixedly coupled to carrier 22C. Sun gear 22S is fixedly coupled to input shaft 11. Carrier 22C is fixedly coupled to output shaft 12.
FIG. 3 shows the foregoing automatic transmission in first gear where clutches 26 and 27 are engaged, and clutches 23, 24 and 25 are disengaged. FIG. 4 shows the automatic transmission in second gear where clutches 23 and 27 are engaged, and clutches 24, 25 and 26 are disengaged. Accordingly, a shift from first gear to second gear is implemented by maintaining the clutch 27 in the engaged state, maintaining the clutches 24 and 25 in the respective disengaged states, shifting the clutch 26 from the engaged state into the disengaged state, and shifting the clutch 23 from the disengaged state into the engaged state.
FIG. 5 schematically shows an analogous automatic transmission having a simplified construction in relation to the gear shift shown in FIGS. 3 and 4. This automatic transmission includes an input shaft 35, an output shaft 36, and two parallel power paths therebetween. Input shaft 35 serves as an input section for receiving an input torque from a drive unit. Output shaft 36 serves as an output section for outputting an output torque. One of the power paths includes a first clutch 33 and a gear arrangement 31 arranged in series. The other power path includes a second clutch 34 and a gear arrangement 32 arranged in series. Gear arrangement 31 allows output shaft 36 to rotate at a transmission gear ratio r1 with respect to input shaft 35 in first gear. Gear arrangement 32 allows output shaft 36 to rotate at a transmission gear ratio r2 with respect to input shaft 35 in second gear. First clutch 33 is selectively connected on one side to input shaft 35, and on the other side to output shaft 36 through the gear arrangement 31 and a final gear 37. Similarly, second clutch 34 is selectively connected on one side to input shaft 35, and on the other side to output shaft 36 through the gear arrangement 32 and final gear 37. Final gear 37 gives a gear ratio rf which is assumed to be equal to one in the following. In other words, first clutch 33 has an engaged state allowing the output shaft 36 to rotate at transmission gear ratio r1 with respect to input shaft 35, and a disengaged state allowing the output shaft 36 to rotate at transmission gear ratio r2 with respect to input shaft 35, while second clutch 34 has an engaged state allowing the output shaft 36 to rotate at transmission gear ratio r2 with respect to input shaft 35, and a disengaged state allowing the output shaft 36 to rotate at transmission gear ratio r1 with respect to input shaft 35.
The shift from first gear to second gear in the four-speed automatic transmission of FIGS. 2 to 4 is considered to be implemented by a process analogous to a process including shifting the first clutch 33 from the engaged state into the disengaged state, and shifting the second clutch 34 from the disengage state into the engaged state in the two-speed automatic transmission of FIG. 5. This process is referred to as “clutch changeover”. Such a gear shift requires changes in rotational speed of internal rotating members, and thereby causes an inertia torque. In the case of an upshift, the internal rotating members rotate more slowly after the upshift than before the upshift. Accordingly, such an upshift causes a positive inertia torque. It is desired to cancel (absorb, for an upshift) such an inertia torque. In order to cancel such an inertia torque, an inertia control phase (inertia phase) is provided before or after a clutch changeover phase as described below in detail.
The following describes three examples of how an inertia torque resulting from a rotational speed change is absorbed during an upshift in the automatic transmission of FIG. 5 with reference to FIGS. 6A, 6B, 6C, 7A, 7B, 7C, 8A, 8B and 8C. Each of FIGS. 6A, 7A and 8A shows how an input torque Tin changes with time, where a solid line represents input torque Tin, a dashed line represents the torque capacity (or transmitted torque) of the first clutch (to-be-disengaged clutch) TC1, and a long dashed short dashed line represents the torque capacity (or transmitted torque) of the second clutch (to-be-engaged clutch) TC2. Each of FIGS. 6B, 7B and 8B shows how an input shaft speed ωin changes with time, where a solid line represents input shaft speed ωin, a dashed line represents the output speed of the first clutch ωC1, a long dashed short dashed line represents the output speed of the second clutch ωC2. Each of FIGS. 6C, 7C and 8C shows how an output torque To changes with time. FIGS. 6A to 8C show changes from before a time instant t1 at which the inertia control phase starts. In these examples, the inertia control phase follows the clutch changeover phase.
In the example of FIGS. 6A to 6C, as shown in FIG. 6A, input torque Tin is held constant during the upshift with no engine torque reduction control. The inertia control phase starts at time instant t1 and terminates at a time instant t2. Input shaft speed ωin is reduced generally from first clutch output speed ωC1 to second clutch output speed ωC2 during the inertia control phase, as shown in FIG. 6B. The reduction of input shaft speed ωin is implemented by increasing the torque capacity of the to-be-engaged clutch TC2 and maintaining it constant above the input torque Tin as indicated by F61 in FIG. 6A. This excess torque is proportional to the rate of change of input shaft speed ωin (or proportional to the acceleration of input shaft 35). In other words, the change in input shaft speed ωin causes an inertia torque, which is identical in magnitude to the excess torque. The inertia torque is transmitted to output shaft 36 as an additional output torque. As shown in FIG. 6C, the output torque To decreases gradually due to the difference between the before-shift and after-shift transmission gear ratios during the clutch changeover phase as indicated by F62, and then increases due to the inertia torque during the inertia control phase as indicated by F63. Thus, the output torque To contains a toque step, causing a shock. When the automatic transmission is applied to a motor vehicle, such a shock adversely affects the ride quality.
In the example of FIGS. 7A to 7C, as shown in FIG. 7A, the input torque Tin is reduced by engine torque reduction control as indicated by F71. The reduction of input shaft speed ωin in the inertia control phase is implemented by setting the torque capacity of the to-be-engaged clutch TC2 above the input torque Tin as indicated by F72 in FIG. 7A. This excess torque is equal to the reduction in input torque Tin. Thus, as shown in FIG. 7C, the output torque To decreases gradually due to the difference between the before-shift and after-shift transmission gear ratios during the clutch changeover phase as indicated by F73, and then is held constant during the inertia control phase as indicated by F74. In this case, the inertia torque is cancelled by the reduction in input torque Tin, so that no shock is outputted to output shaft 36.
However, there are situations where it is impossible to cancel the inertia torque by engine torque reduction control as in FIGS. 7A to 7C. The situations include a situation where an upshift is carried out under condition that an accelerator pedal is little depressed or completely released so that no further engine torque reduction is possible because no engine torque is outputted before the upshift. FIGS. 8A to 8C show such a situation. As shown in FIG. 8A, the input torque Tin is reduced by engine torque reduction control during the inertia control phase. The reduction in the input torque Tin is limited as indicated by F81, although a further reduction is desired as indicated by F83. In order to implement the reduction in input shaft speed ωin, the torque capacity of the to-be-engaged clutch TC2 is increased and maintained above the input torque Tin as indicated by F82 in FIG. 8A. This excess torque is larger than the decrease in input torque Tin. Thus, as shown in FIG. 8C, the output torque To decreases gradually due to the difference between the before-shift and after-shift transmission gear ratios during the clutch changeover phase as indicated by F84, and then increases during the inertia control phase as indicated by F86, although reduced to some extent by engine torque reduction control as indicated by F85 as compared to the example of FIGS. 6A to 6C. Thus, the output torque To contains a toque step, causing a shock.
The following describes a method of canceling an inertia torque resulting from a gear shift by controlling clutches. FIG. 9 illustrates a general relationship in a clutch among input speed ω1, output speed ω2, torque capacity C and transmitted torque TC. The relationship is expressed by equation (A).
T
C
=C*sign(ω1−ω2) (A)
- where C represents the torque capacity (=(coefficient of friction)·(apply force)), and sign(x) represents a sign operator indicative of the sign of the variable x.
According to equation (A), the direction of the transmitted torque depends on the relationship between input shaft speed ω1 and output shaft speed ω2.
Equation (A) is applied to the two-speed transmission shown in FIG. 5. FIG. 10 illustrates how clutch engagement serves to change input shaft speed ωin in three different states of the relationship among input shaft speed ωin, first clutch output speed ωC1, and second clutch the output speed ωC2 in the automatic transmission of FIG. 5. Input shaft 35 is subject to input torque Tin, clutch torque TC1 and clutch torque TC2.
FIG. 10 shows a first state of the transmission where input shaft speed ωin is higher than both of first clutch output speed ωC1 and second clutch output speed ωC2. In the first state, both of the clutch torques TC1 and TC2 act in the opposite direction to input torque Tin so as to reduce the input shaft speed ωin. Accordingly, the input shaft speed ωin decreases, when both of first and second clutches 33 and 34 are in respective sliding engaged states.
FIG. 10 also shows a second state of the transmission where input shaft speed ωin is higher than second clutch output speed ωC2, and lower than first clutch output speed ωC1. In the second state, the clutch torque TC1 acts in the same direction as input torque Tin so as to increase the input shaft speed ωin, while the clutch torque TC2 acts in the opposite direction to input torque Tin so as to reduce the input shaft speed ωin. Accordingly, the input shaft speed ωin does not always decrease, when the first and second clutches 33 and 34 are in respective sliding engaged states.
FIG. 10 further shows a third state of the transmission where the input shaft speed ωin is lower than both of first clutch output speed ωC1 and second clutch output speed ωC2. In the third state, both of the clutch torques TC1 and TC2 act in the same direction as input torque Tin so as to increase the input shaft speed ωin.
The following concerns the foregoing second state of the transmission, and describes how to determine the clutch torque capacities in balance so as to cancel the inertia torque due to change in the input shaft speed, and to prevent fluctuations in the output torque.
With regard to the two-speed transmission shown in FIG. 5, an equation of motion as to input shaft 35 is expressed by equation (1), while an equation of motion as to output shaft 36 is expressed by equation (2).
T
in
=T
c1
+T
c2
+I
in{dot over (ω)}in (1)
T
0
=r
1(Tc1−IC1{dot over (ω)}C1)+r2(Tc2−IC2ωC2) (2)
where Tin: an input torque to input shaft 35,
TC1: a transmitted torque of first clutch 33,
TC2: a transmitted torque of second clutch 34,
Iin: the moment of inertia of input shaft 35,
ωin: the rotational speed of input shaft 35,
To: an output torque to output shaft 36,
r1: a transmission gear ratio through a power path with first clutch 33,
r2: a transmission gear ratio through a power path with second clutch 34,
IC1: the moment of inertia of the power path with first clutch 33,
IC2: the moment of inertia of the power path with second clutch 34,
{dot over (ω)}C1: the rate of change of the output speed of first clutch 33, and
{dot over (ω)}C2: the rate of change of the output speed of second clutch 34.
Applying the equation (A) to each of first and second clutches 33 and 34 gives equation (3).
T
C1
=C
1×sign(ωin−ωC1)
T
C2
=C
2×sign(ωin−ωC2) (3)
where C1: the torque capacity of first clutch 33,
C2: the torque capacity of second clutch 34,
ωin: the rotational speed of input shaft 35,
ωC1: the output speed of first clutch 33,
ωC2: the output speed of second clutch 34, and
sign: the sign operator.
It is assumed that first clutch 33 is a to-be-disengaged clutch and second clutch 34 is a to-be-engaged clutch in an upshift of the transmission. When the transmission is in the second state, the relationship of ωC1>ωin>ωC2 holds. Substituting this relationship into equation (3) yields equation (4).
T
C1
=−C
1
, T
C2
=C
2 (4)
The following concerns an under-acceleration upshift as an upshift which is carried out while the vehicle speed is increasing. In such an under-acceleration upshift, disengagement of the to-be-disengaged clutch allows input shaft speed ωin to be increased by input torque Tin, while the upshift requires a decrease in input shaft speed ωin. Accordingly, this upshift is implemented by a prior clutch changeover phase, and a subsequent inertia control phase where the torque capacity of the to-be-engaged clutch is dominant. Input torque Tin is usually positive. In view of the foregoing, the clutch torque capacities just prior to the inertia control phase are expressed by equation (5). When a torque capacity correction of the first clutch ΔC1 and a torque capacity correction of the second clutch ΔC2 are provided in the inertia control phase, the clutch torque capacities are expressed by equation (6).
C
1
C=0
C
2
=C
20
=|T
in| (5)
C
1
=C
10
+ΔC
1
=ΔC
1
C
2
=C
20
+ΔC
2
=|T
in
|+ΔC
2 (6)
where C10: the torque capacity of first clutch 33 just prior to the start of the inertia control phase, and
C20: the torque capacity of second clutch 34 just prior to the start of the inertia control phase.
Substituting the equation (6) into the equation (1) yields equation (7). Assuming Tin>0 gives Tin=|Tin|. Substituting this into the equation (7) yields the equation (8).
T
in
=−ΔC
1
+|T
in
|+ΔC
2
+I
in{dot over (ω)}in (7)
I
in{dot over (ω)}in=ΔC1−ΔC2 (8)
Substituting the equation (6) into the equation (2) yields the equation (9). In order to cause no change in output torque To, and output shaft speed ωout by first and second clutch torque capacity corrections ΔC1 and ΔC2, the equation (10) is required. The equation (10) is reduced to the equation (11).
Substituting the equation (8) into the equation (11) yields the equation (12). The equations (11) and (12) are reduced to the equation (13) where ΔC1 and ΔC2 are related to the inertia torque Iinωin.
The following concerns an under-deceleration upshift as an upshift which is carried out while the accelerator pedal is released after the accelerator pedal is depressed to downshift the transmission. In such an under-deceleration upshift, disengagement of the to-be-disengaged clutch allows input shaft speed ωin to be reduced by input torque Tin which is negative, while the upshift requires a decrease in input shaft speed ωin. Accordingly, this upshift is implemented by a prior inertia control phase where the torque capacity of the to-be-disengaged clutch is dominant, and a subsequent clutch changeover phase. Input torque Tin is usually negative. In view of the foregoing, the clutch torque capacities just prior to the inertia control phase are expressed by equation (14). When a torque capacity correction of the first clutch ΔC1 and a torque capacity correction of the second clutch ΔC2 are provided in the inertia control phase, the clutch torque capacities are expressed by equation (15).
C
1
=C
10
=|T
in|
C
2
=C
20=0 (14)
C
1
=C
10
+ΔC
1
|T
in
|+ΔC
1
C
2
=C
20
+ΔC
2
=ΔC
2 (15)
Substituting the equation (15) into the equation (1) yields the equation (16). Assuming Tin<0 gives Tin=−|Tin|. Substituting this into the equation (16) yields the equation (17).
T
in
=−ΔC
1
−|T
in
|+ΔC
2
+I
in{dot over (ω)}in (16)
I
in{dot over (ω)}in=ΔC1−ΔC2 (17)
The equation (17) is the same as the equation (8). Therefore, the equations (8) to (13) holds also in the case of under-deceleration upshift, when the clutch changeover phase follows the inertia control phase as contrasted to the case of under-acceleration upshift. The above equations (especially, equation (13)) serve to compute a desired torque capacity correction of the first clutch for inertia torque cancellation ΔC1 and a desired torque capacity correction of the first clutch for inertia torque cancellation ΔC2 on the basis of the desired clutch-based torque reduction. The first and second clutches 33 and 34 are controlled to achieve these desired corrections so as to cancel an allocated portion of the inertia torque.
FIG. 1 schematically shows an automatic transmission with a shift control apparatus in accordance with embodiments of the present invention. The shift control apparatus comprises a controller (inertia torque distribution section) 1 for controlling gear shifts of the automatic transmission. Controller 1 controls an engine 2 as an external drive unit, and the first and second clutches so as to cancel an inertia torque resulting from a change in a rotational state of the automatic transmission (absorb a positive inertia torque or compensate for a negative inertia torque). The controller is an electrical control unit (ECU).
As shown in the equation (13), the desired torque capacity correction of each clutch is a product of the inertia torque Iin{dot over (ω)}C1 and a gain (coefficient) expressed in terms of the transmission gear ratios r1 and r2 which are established by engaging respective clutches. The coefficient is expressed as a fraction having a denominator of the difference between the transmission gear ratios (r2−r1). Accordingly, as the term of (r2−r1) decreases, the coefficient increases so as to increase the desired torque capacity corrections. In the case of a typical transmission having a stepwise transmission gear ratio set, the term of (r2−r1) for an upshift between lower gears, such as from first gear to second gear or from second gear to third gear, is larger than an upshift between higher gears. Therefore, it is appropriate to control the desired torque capacity correction of each clutch within an allowable range in the case of an upshift between higher gears.
When the desired torque capacity corrections are achieved, the first and second clutches 33 and 34 are controlled toward respective engaged states in parallel so as to cancel (absorb, for upshifts) an inertia torque resulting from an upshift. In other words, the inertia torque is absorbed by controlling the first and second clutches 33 and 34 into a state close to an interlock state so that the kinetic energy of the moment of inertia is converted into a thermal energy through friction. Accordingly, it is desired to suppress the desired torque capacity corrections as small as possible.
Moreover, it is desired to suppress increasingly a load to the engine in canceling the inertia torque as the vehicle speed increases, because the output torque of the engine is small under such conditions. From this viewpoint, the shift control apparatus allows the desired torque capacity corrections to increase with increasing vehicle speed, and suppresses the desired torque capacity corrections when the vehicle speed is low.
The following describes a shift control apparatus for controlling clutches so as to cancel the inertia torque in accordance with a first embodiment of the present invention with reference to FIGS. 11 to 15C. FIG. 11 shows a section of the shift control apparatus for computing a desired torque capacity correction of each of first and second clutches 33 and 34 for inertia torque cancellation.
As shown in FIG. 11, first, a calculator 101 computes or calculates a difference (r2−r1) on the basis of the before-shift transmission gear ratio r1 and the after-shift transmission gear ratio r2. A calculator 102 calculates a gain for second clutch 34 by dividing the before-shift transmission gear ratio r1 by the difference (r2−r1). A calculator 103 calculates a gain for first clutch 33 by dividing the after-shift transmission gear ratio r2 by the difference (r2−r1). A calculator 104 calculates a corrected desired clutch-based torque reduction on the basis of a desired clutch-based torque reduction and a correction factor for suppressing it. Specifically, calculator 104 corrects the desired clutch-based torque reduction by multiplying the correction factor. The correction factor is stored in a map section 100. The correction factor is determined based on the vehicle speed of a motor vehicle and gear selection. Specifically, the correction factor is equal to zero when the vehicle speed is low, starts to increase when the vehicle speed is middle, increases as the vehicle speed increases, and is equal to a maximum value of 1 when the vehicle speed is high. Such characteristics are prepared in the form of a map for each gear. The lines defining the correction factor are shifted in the positive direction of the axis of the vehicle speed in the map, as the selected gear becomes higher. Thus, the desired clutch-based torque reduction is increasingly suppressed with decreasing vehicle speed, and increasingly suppressed as the elected gear becomes high.
A calculator 105 calculates a desired torque capacity correction of the second clutch ΔC2 by multiplying the desired clutch-based torque reduction corrected by the calculator 104 and the gain calculated by the calculator 102. Similarly, a calculator 106 calculates a desired torque capacity correction of the first clutch ΔC1 by multiplying the desired clutch-based torque reduction corrected by the calculator 104 and the gain calculated by the calculator 103. Thus, the desired torque capacity corrections ΔC1 and ΔC2 are increasingly suppressed with decreasing vehicle speed, and increasingly suppressed as the elected gear becomes high.
According to the above calculation of the desired torque capacity corrections in consideration of the correction factor, it is possible to control the desired engine-based torque reduction and desired clutch-based torque reduction in balance, and to prevent the desired torque capacity corrections from increasing excessively, depending on the vehicle speed.
FIG. 12 shows a section of the shift control apparatus for computing a desired engine-based torque reduction and a desired clutch-based torque reduction for inertia torque cancellation, on the basis of an inertia torque to be cancelled. In general, an engine-based torque reduction is restricted within an upper bound so as to allow the engine to continue to rotate. This upper bound is calculated on the basis of parameters such as engine speed, output torque, combustion mode (lean combustion mode, theoretical air fuel ratio combustion mode, etc.), the number of active cylinders, ignition timing, coolant temperature, and the operating state of auxiliary equipment attached to the engine. A comparator 107 compares the to-be-cancelled inertia torque with the maximum possible engine-based torque reduction Tde, and outputs the smaller one of the two as a desired engine-based torque reduction. A calculator 108 calculates a desired clutch-based torque reduction by subtracting the to-be-cancelled inertia torque by the desired engine-based torque reduction. A comparator 109 compares the desired clutch-based torque reduction with zero, and outputs the larger one of the two as a final desired clutch-based torque reduction. This is because the clutches can attain no negative torque capacity.
Thus, the desired torque capacity correction of each clutch is calculated from the to-be-cancelled inertia torque according to the processes shown in FIGS. 11 and 12. FIG. 13 shows a section of the shift control apparatus for computing a corrected torque capacity of each of the first and second clutches for inertia torque cancellation. In FIG. 13, the corrected torque capacity of each clutch is calculated by adding an uncorrected torque capacity and the desired torque capacity correction. The uncorrected torque capacity is a torque capacity required to implement the gear shift without consideration of the inertia torque. The corrected torque capacity of each clutch is achieved by supplying a corresponding clutch engagement pressure to the clutch.
The uncorrected torque capacity of each clutch is given so that the input shaft speed ωin decreases linearly. This is implemented by setting the uncorrected torque capacity so that the differential speed (ωC1−ωin) increases linearly or the differential speed (ωin−ωC2) decreases linearly. The foregoing calculation of the corrected torque capacity of each clutch is carried out when the relationship of ωC1>ωin>ωC2 holds. As shown in FIG. 13, only when a comparator 110 determines that first clutch output speed ωC1 is higher than input shaft speed ωin, a comparator 111 determines that input shaft speed ωin is higher than second clutch output speed ωC2, and a logic section 112 determines that the relationship of ωC1>ωin>ωC2 holds, addition-permitting switches 113 and 114 allow an adder 115 to calculate the corrected torque capacity of first clutch 33, and allow an adder 116 to calculate the corrected torque capacity of second clutch 34.
The thus-constructed shift control apparatus operates to cancel an inertia torque during an upshift of the automatic transmission as follows. Each of FIGS. 14A and 15A shows how the input torque Tin changes with time, where a solid line represents the input torque Tin a dashed line represents the torque capacity (or transmitted torque) of the first clutch (to-be-disengaged clutch) TC1, and a long dashed short dashed line represents the torque capacity (or transmitted torque) of the second clutch (to-be-engaged clutch) TC2. Each of FIGS. 14B and 15B shows how the input shaft speed ωin changes with time, where a solid line represents the input shaft speed ωin, a dashed line represents the output speed of the first clutch ωC1, a long dashed short dashed line represents the output speed of the second clutch ωC2. Each of FIGS. 14C and 15C shows how the output torque To changes with time. FIGS. 14A to 15C show changes from before the time instant t1 at which the inertia control phase starts.
In the example of FIGS. 14A to 14C, an under-acceleration upshift is implemented by a prior clutch changeover phase and a subsequent inertia control phase. As shown in FIG. 14A, input torque Tin is reduced by engine torque reduction control during the inertia control phase. The reduction in the input torque Tin is limited as indicated by F142, although a further reduction is desired as indicated by F144. A time instant t3 is when the input shaft speed ωin decreases below the first clutch output speed ωC1. In order to implement the speed reduction of input shaft speed ωin, the torque capacity of the to-be-engaged clutch TC2 is increased and maintained above the input torque Tin as indicated by F143. The torque capacity of the to-be-disengaged clutch TC1 and to-be-engaged clutch TC2 are further increased as indicated by hatched patterns of F141 and F145. As shown in FIG. 14C, the output torque To decreases gradually due to the difference between the before-shift and after-shift transmission gear ratios during the clutch changeover phase as indicated by F146, and then is held constant except an instantaneous increase over a time interval from time instant t1 to time instant t3. This time interval exists because the clutch-based torque reduction is impossible while the first clutch output speed ωC1 is lower than the input shaft speed ωin as indicated by F147. Thus, the engine-based torque reduction serves to cancel a portion of the output torque To as indicated by F148, while the clutch-based torque reduction serves to cancel another portion of the output torque To, which cannot be cancelled by the engine as indicated by F149, as indicated by F140.
In the example of FIGS. 15A to 15C, an under-deceleration upshift is implemented by a prior inertia control phase and a subsequent clutch changeover phase. During such an under-deceleration upshift, the engine torque is usually minimum. Accordingly, there is no available engine-based torque reduction. If the engine torque is negative, canceling the inertia torque requires an increase in the engine torque, for example, by terminating a fuel cut. This may effect adversely other functions of the engine. Even in such cases, it is possible to cancel all of the inertia torque resulting from change in the rotational state by controlling the torque capacity of the first and second clutches 33 and 34 with no engine-based torque reduction as follows. As shown in FIG. 15A, the absolute value of input torque |Tin| is held constant during the upshift as indicated by F153. A time instant t4 is when the input shaft speed ωin decreases below the second clutch output speed ωC2. In order to implement the speed reduction of input shaft speed ωin, the torque capacity of the to-be-engaged clutch TC2 is increased and maintained above the input torque Tin as indicated by F152. The torque capacity of the to-be-disengaged clutch TC1 and to-be-engaged clutch TC2 are further increased as indicated by hatched patterns of F151 and F154. As shown in FIG. 15C, the output torque To is held constant during the inertia control phase except an instantaneous increase (decrease in magnitude) over a time interval from time instant t4 to time instant t2 as indicated by F156, and then decreases in magnitude due to the difference between the before-shift and after-shift transmission gear ratios during the clutch changeover phase as indicated by F157. This time interval exists because the clutch-based torque reduction is impossible while the input shaft speed ωin is lower than the second clutch output speed ωC2 as indicated by F155. Thus, the clutch-based torque reduction serves to cancel almost all of the output torque To as indicated by F156.
The following describes a shift control apparatus for controlling clutches so as to cancel the inertia torque in accordance with a second embodiment of the present invention with reference to FIGS. 16 to 20C. In the first embodiment, it is possible that the inertia torque is insufficiently cancelled to cause an increase in the output torque To when the relationship of first clutch output speed ωC1>input shaft speed ωin>second clutch output speed ωC2 does not hold. The second embodiment concerns a solution to this problem. Specifically, in the second embodiment, input shaft speed ωin is controlled to change with such a dynamic characteristic that the rate of change of input shaft speed ωin is small at the start and end of the change as shown in FIGS. 19B and 20B, in contrast to the first embodiment where the input shaft speed changes linearly.
More specifically, the shift control apparatus of the second embodiment sets a target input shaft speed ωin* as indicated by a dashed line in FIGS. 19B and 20B which decreases along a curved line; sets the uncorrected desired torque capacity of each clutch so that input shaft speed ωin follows target input shaft speed ωin*; sets the torque capacity correction of each clutch so as to cancel an inertia torque resulting from change in target input shaft speed ωin*; sets the corrected desired torque capacity of each clutch by adding the uncorrected desired torque capacity and torque capacity correction; and controls the first and second clutches 33 and 34 to attain the corrected desired torque capacity.
Although input shaft speed ωin is directly controlled in the foregoing description, the difference in first clutch 33 between the input and output speeds (ωC1−ωin) or the difference in second clutch 34 between the input and output speeds (ωin−ωC2) may be controlled. In such cases, the torque capacity of each clutch is controlled so that such differential speed changes (increases or decreases) along a curved line.
FIGS. 16 to 18 show three examples of a section of the shift control apparatus for computing a target input shaft speed (or a target differential speed between the input and output of a clutch) with a desired dynamic characteristic. In FIG. 16, a calculator 117 calculates an overall speed change of the input shaft speed resulting from the upshift (ωC1−ωC2) on the basis of first clutch output speed ωC1 and second clutch output speed ωC2. An output controller (switch) 118 outputs the overall speed change (ωC1−ωC2) after an inertia control phase start signal is inputted into output controller 118, and outputs a signal indicative of zero until the inertia control phase start signal is inputted. While output controller 118 outputs the signal indicative of zero, input shaft speed ωin is uncontrolled. After the inertia control phase start signal is inputted, a target differential speed dynamic characteristic section 119 receives the overall speed change (ωC1−ωC2), and imparts a dynamic characteristic to the target differential speed as described below in detail. A target differential acceleration dynamic characteristic section 120 receives the overall speed change (ωC1−ωC2), and imparts a dynamic characteristic to a target differential acceleration as described below in detail. A differential acceleration of a clutch is defined as a rate of change of a differential speed between the input and output of the clutch.
The target differential speed dynamic characteristic section 119 determines a desired change in the differential speed of the second clutch so that the desired change varies every control cycle, and that the differential speed in the second clutch (or input shaft speed ωin) changes along a curved line. A calculator 121 receives the desired change, and calculates a target differential speed by subtracting the desired change from the overall speed change (ωC1−ωC2). Specifically, target differential speed dynamic characteristic section 119 controls the desired change in the differential speed in the second clutch per control cycle to be small at the start of the inertia control phase, to increase gradually with time in the first half, to decrease gradually with time in the second half, and to be small at the end of the inertia control phase.
The target differential acceleration dynamic characteristic section 120 stores a desired dynamic characteristic of the differential acceleration of the second clutch which is beforehand obtained by differentiating the dynamic characteristic of the target differential speed, and determines a desired differential acceleration in the second clutch. A calculator 122 multiplies the desired differential acceleration in the second clutch by a coefficient of inertia to obtain the to-be-cancelled inertia torque. Thus, controller 1 determines a desired path of change of input shaft speed ωin through the gear shift; determines a target rate of change of the input speed in accordance with the desired path; and determines the inertia torque in accordance with the target rate of change of the input speed.
In FIG. 17, the required speed change through the inertia control phase is calculated on the basis of the output shaft speed ωout, before-shift transmission gear ratio r1, and after-shift transmission gear ratio r2. A calculator 123 calculates the first clutch output speed ωC1 by multiplying the before-shift transmission gear ratio r1 and the output shaft speed ωout. A calculator 124 calculates the second clutch output speed ωC2by multiplying the after-shift transmission gear ratio r2 and the output shaft speed ωout. The remaining sections in FIG. 17 are the same as in FIG. 16.
In FIG. 18, the required speed change through the inertia control phase is calculated on the basis of the input shaft speed ωin and the second clutch output speed ωC2. A holder 131 specifies and stores a value of the input shaft speed ωin when the inertia control phase start signal is inputted. The calculator 117 calculates the overall speed change by subtracting the second clutch output speed ωC2 from the stored value of the input shaft speed ωin. The remaining sections in FIG. 18 are the same as in FIG. 16. The example of FIG. 18 may be combined with the example of FIG. 17. Specifically, controller 1 may be configured to: store a value of input shaft speed ωin when the inertia control phase start signal is inputted; determine a reference speed in accordance with the after-shift transmission gear ratio r2 and input shaft speed ωout; and determine the overall speed change as a difference between the stored value and the reference speed.
The thus-constructed shift control apparatus operates to cancel an inertia torque during an upshift of the automatic transmission as follows. Each of FIGS. 19A and 20A shows how the input torque Tin changes with time, where a solid line represents the input torque Tin, a dashed line represents the torque capacity (or transmitted torque) of the first clutch (to-be-disengaged clutch) TC1, and a long dashed short dashed line represents the torque capacity (or transmitted torque) of the second clutch (to-be-engaged clutch) TC2. Each of FIGS. 19B and 20B shows how the input shaft speed ωin changes with time, where a solid line represents the input shaft speed ωin, a dashed line represents the output speed of the first clutch ωC1, a long dashed short dashed line represents the output speed of the second clutch ωC2. Each of FIGS. 19C and 20C shows how the output torque To changes with time. FIGS. 19A to 20C show changes from before the time instant t1 at which the inertia control phase starts.
In the example of FIGS. 19A to 19C, an under-acceleration upshift is implemented by a prior clutch changeover phase and a subsequent inertia control phase. As shown in FIG. 19A, input torque Tin is reduced by engine torque reduction control during the inertia control phase. Input torque Tin starts to decrease slowly at the start of the inertia control phase. The reduction in the input torque Tin is limited, although a further reduction is desired as indicated by F193. In order to implement the speed reduction of input shaft speed ωin, the torque capacity of the to-be-engaged clutch TC2 is increased and maintained above the input torque Tin as indicated by F192. The torque capacity of the to-be-disengaged clutch TC1 and to-be-engaged clutch TC2 are further increased as indicated by hatched patterns of F191 and F194. As shown in FIG. 19C, the output torque To decreases gradually due to the difference between the before-shift and after-shift transmission gear ratios during the clutch changeover phase as indicated by F195, and then is held constant after that. Thus, the engine-based torque reduction serves to cancel a portion of the output torque To as indicated by F198, while the clutch-based torque reduction serves to cancel another portion of the output torque To as indicated by F197. In this example, even when input shaft speed ωin is higher than first clutch output speed ωC1 so that it is impossible to cancel the inertia torque by clutch engagement, there is no instantaneous increase in output torque To, as indicated by F196, because the rate of change of input shaft speed ωin is controlled to be close to zero at the start of the inertia control phase.
In the example of FIGS. 20A to 20C, an under-deceleration upshift is implemented by a prior inertia control phase and a subsequent clutch changeover phase. As shown in FIG. 20A, the absolute value of input torque Tin is held constant with no engine-based torque reduction during the inertia control phase as indicated by F203. A time instant t5 is when the input shaft speed ωin decreases below the second clutch output speed ωC2 In order to implement the speed reduction of input shaft speed ωin, the torque capacity of the to-be-engaged clutch TC2 is increased and maintained above the input torque Tin as indicated by F202. The torque capacity of the to-be-disengaged clutch TC1 and to-be-engaged clutch TC2 are further increased as indicated by hatched patterns of F201 and F204. As shown in FIG. 20C, the output torque To is held constant during the inertia control phase except a slight instantaneous increase (decrease in magnitude) over a time interval from time instant t5 to time instant t2 as indicated by F206, and decreases in magnitude due to the difference between the before-shift and after-shift transmission gear ratios during the clutch changeover phase as indicated by F207. Thus, the clutch-based torque reduction serves to cancel a portion of the output torque To as indicated by F205. The instantaneous increase in output torque To during the time interval from time instant t5 to time instant t2 as indicated by F206 is small, because the rate of change of input shaft speed ωin is controlled to be close to zero at the end of the inertia control phase.
The provision of the dynamic characteristics of the input shaft speed serves to control a gear shift by a universal logic independently of the relationship between the input shaft speed ωin, first clutch output speed ωC1 and second clutch output speed ωC2, and to control suitably the timing of clutch engagement or clutch disengagement. In this embodiment, two different kinds of dynamic characteristics are provided. The first dynamic characteristic concerns the target input shaft speed (or target differential speed), while the second dynamic characteristic concerns the rate of change of the target input shaft speed (or target differential acceleration). The first dynamic characteristic is used to determine a path of change in the target input shaft speed. The second dynamic characteristic is prepared beforehand by differentiating the first dynamic characteristic. The second dynamic characteristic is used to calculate the desired rate of change of the input shaft speed on the basis of the required speed change through the gear shift. The provision of the first and second dynamic characteristics serves to calculate the desired rate of change of the input shaft speed with no differentiating calculation, and to carry out the required calculations in a stable manner, preventing noises from adversely affecting the on-time calculation results.
FIG. 21 shows a modification of the section of the shift control apparatus of FIG. 12. In FIG. 21, a map section 200 determines a distribution ratio on the basis of the maximum possible engine-based torque reduction for inertia torque cancellation. A calculator 201 calculates a basic desired engine-based torque reduction by multiplying the distribution ratio and the maximum possible engine-based torque reduction. A comparator 202 compares the basic desired engine-based torque reduction with the to-be-cancelled inertia torque, and outputs the smaller one of the two as a second basic desired engine-based torque reduction. A calculator 203 calculates a basic desired clutch-based torque reduction by subtracting the second basic desired engine-based torque reduction from the to-be-cancelled inertia torque. A calculator 204 determines a final desired engine-based torque reduction by providing a compensation for delay in response of the hydraulic system for the clutch set. A comparator 205 compares zero with the basic desired clutch-based torque reduction, and outputs the larger one of the two. A calculator 206 determines a final desired clutch-based torque reduction by providing a compensation for delay in response of the engine torque. The distribution ratio is set so that the clutch-based torque reduction is dominant when the maximum possible engine-based torque reduction is small and less effective, and the engine-based torque reduction is dominant when the maximum possible engine-based torque reduction is sufficiently large and effective. The two compensations in the calculators 204 and 206 serve to synchronize the timing of operation of the engine and the clutch set with one another.
FIG. 22 shows another modification of the section of the shift control apparatus of FIG. 12. This modification differs from the modification of FIG. 21 as follows. In FIG. 22, a map section 300 determines a distribution ratio on the basis of the maximum possible engine-based torque reduction for inertia torque cancellation. The distribution ratio is set to be constantly smaller than one, in contrast to the case of FIG. 21. A calculator 301 calculates an output value by multiplying the distribution ratio and the to-be-cancelled inertia torque. Comparator 202 compares the basic desired engine-based torque reduction (calculated by multiplying the maximum possible engine-based torque reduction and the distribution ratio in calculator 201) with the output value of calculator 301, and outputs the smaller one of the two as a second basic desired engine-based torque reduction. Naturally, the second basic desired engine-based torque reduction is constantly smaller than the to-be-cancelled inertia torque. Thus, the desired clutch-based torque reduction is set to a non-zero value at least above the product of the to-be-cancelled inertia torque and the value of (1-maximum distribution ratio), even when the to-be-cancelled inertia torque can be totally covered by engine-based torque reduction.
When the maximum possible value of the engine-based torque reduction is equal to a non-zero value, the distribution ratio may be determined to be smaller than one and larger than zero. When the maximum possible value is equal to zero, the distribution ratio may be changed gradually with time toward zero. When the maximum possible value is larger than zero and smaller than a reference value, the engine-based torque reduction may be determined by multiplying the inertia torque by the distribution ratio. When the maximum possible value of the second portion of the inertia torque is larger than a reference value, the distribution ratio may be changed gradually with time toward one. The foregoing causes no instantaneous and rapid shift in the state of the transmission during the inertia torque cancellation, thereby preventing possible resulting fluctuations in speed and torque.
Although the foregoing description concerns the simplified two-speed transmission shown in FIG. 5, a shift control apparatus may be constructed for other different types of transmissions such as the four-speed automatic transmission shown in FIG. 2. This may be implemented by replacing the input shaft speed ωin of the two-speed transmission by the input speed of each clutch of the four-speed automatic transmission, and by calculating a clutch torque capacity correction for canceling an inertia torque with respect to an uncorrected torque capacity of each clutch.
This application is based on a prior Japanese Patent Application No. 2006-196135 filed on Jul. 18, 2006. The entire contents of this Japanese Patent Application No. 2006-196135 are hereby incorporated by reference.
Although the invention has been described above by reference to certain embodiments of the invention, the invention is not limited to the embodiments described above. Modifications and variations of the embodiments described above will occur to those skilled in the art in light of the above teachings. The scope of the invention is defined with reference to the following claims.
Patent Application
20080021622
