Source: https://patents.google.com/patent/US6058348A/en
Timestamp: 2018-04-23 21:52:02
Document Index: 616181225

Matched Legal Cases: ['arts 371', 'arts 371', 'art 406', 'art 406', 'art 406', 'art 406', 'art 406', 'art 433', 'arts 477', 'art 477', 'art 477', 'art 477']

US6058348A - Control apparatus for drive system composed of engine and transmission - Google Patents
US6058348A
US6058348A US08431028 US43102895A US6058348A US 6058348 A US6058348 A US 6058348A US 08431028 US08431028 US 08431028 US 43102895 A US43102895 A US 43102895A US 6058348 A US6058348 A US 6058348A
US08431028
FIG. 43 is a sectional view along line XIII--XIII shown in FIG. 41;
Referring to FIG. 7, the speed ratio and engine torque computing means 100 obtains an engine desired torque and a speed ratio, and delivers the engine desired torque to the speed ratio control device 16 and the valve opening and closing timing control device 17 so as to optimumly control the revolution speed of the compressor 7 and the closing angle of the intake valve 9.
Referring to FIG. 8, denoting that the ratio between wheel and engine speeds as a speed ratio x, if x is large, the wheel torque falls inside of a curve A while if x is small, the wheel torque falls inside of a curve B. If a desired value Fo is given to the wheel torque F with respect to the vehicle speed, x in the curve B is selected if F1 >F0 and V1 <V0. If the continuously variable transmission is used, the point B2 continuously varies to a point A1. At this time, the speed of the engine is highest. Further, the point B1 continuously varies to the point A1. At this time the engine speed is lowest.
In view of the fuel economy, the engine is preferably driven at a speed and a load which are as low and high as possible, respectively. Accordingly, it is driven along a curve A1 -B1. If F=F2, it is driven along a curve A1 -A2. If F=F1, it is driven along a curve B1 -B2. If V>V1 and F<F1, the output power of the engine is controlled at a speed ratio B. If F≦F1 and V=V1, it is driven along a curve V1 -B1 while if V=V2, it is driven along a curve V2 -A1. If V<V2, the speed of the engine is lowered if no slip occurs, and accordingly, the clutch is slipped in order to maintain the speed of the engine. A curve RL shown in this figure indicates a desired torque F during a horizontal road surface, with which the engine cannot be drive above a point P2 with a lean mixture.
In order to enlarge the vehicle speed range and the wheel torque range, the mixture is enriched. Accordingly, the torque F2 is increased up to a value F2 ' with which the vehicle can ascend a slope. Further, the torque F, is increased up to a value F1 so that the driveable point is increased to a value P2 '. Referring to FIG. 9, in a range α, it is driven at a lowest engine speed with a partial load and a lean mixture. In a range β where the speed ration is lowest, it is driven at a highest engine speed with a rich mixture. In a range γ, it is driven at a highest speed ratio x with a rich mixture. In a range .di-elect cons. where the speed ratio x is highest, it is driven with a rich mixture. In a range ε, it is driven with a lean mixture while the throttle valve is fully opened. The speed ratio x varies, depending upon the speed V.
In the range ε2, it can be driven at a low engine speed with a rich mixture. Accordingly, on the left side of a curve E1 -E2, it is driven with a rich mixture while the throttle valve is fully opened in order to restrain the engine speed from increasing. In the range ε3, it is driven with a rich mixture and a fully opened throttle valve. On the left side of the curve E1 -E2, it is driven with a rich mixture and a fully opened throttle valve. That is, though it can be driven with a lean mixture and a fully opened throttle valve in the range ε2, the mixture is set to be rich in a range where the engine speed exceeds, for example, 3,000 rpm so as to lower the engine speed since the fuel consumption is increased due to mechanical friction as the engine speed is increased.
As shown in FIG. 11, in the ranges α, β, the air-fuel ratio is controlled in the range from 25 to 80, that is, the higher the air-fuel ratio, the larger the fuel volume. In a range δ, the air-fuel ratio is controlled in a range from 25 to 12. In the range ε, on the left side of the curve E1 -E2, the engine is driven at an air-fuel ratio of 25 while the wheel torque F is adjusted, depending upon a speed ratio x. In the range γ, the air-fuel ratio is set to 12 along a curve a1 -a2 while it is set to 25 along a curve b1 -b2. In this condition, the engine speed is highest, that is, 6,000 rpm.
F=k(n.sub.e /V)T=k.sub.1 (T/x)                             EX1
T.sub.e =T+I*dn.sub.e /dt                                  EX2
where I is an inertia term of a movable part.
T.sub.e =F.sub.0 *x/k.sub.1 +I*dn.sub.e /dt                EX3
In FIG. 14, in a range of A1 -A1 '-B1 '-B1, if the engine is driven by shifting the air-fuel ratio from 25 to 12, the vehicle can be accelerated without changing the speed ratio x. Further, in the range α, if the minimum value of the engine speed has been increased beforehand, the accelerating performance can be enhanced. However, an increase in fuel consumption is inevitable. At a vehicle speed V1, the speed ratio x can be set to be either maximum or minimum. If the speed ratio x is set to be minimum, the wheel torque can be increased along the curve A1,-A2 ' without increasing the engine speed. On the contrary, if the speed ratio x is set to be maximum, the wheel torque along the curve B1 -B2 can only be obtained if the engine speed is maintained to be constant. In order to obtain a higher torque than that, it is required that the speed ratio x is decreased and the engine speed is increased. At this time, a part of the engine generated torque is consumed for accelerating the engine itself.
If the vehicle speed exceed the value V4, the speed ratio x cannot be set to a minimum value, and accordingly, the wheel torque cannot exceeds a curve A2 '-B2 '. Also at this time, if the engine speed during steady-state operation is set to be low, the torque during acceleration is once lowered. The torque along A1 -A2 -B2 can be attained even with a lean mixture. When the curve is exceeded, the mixture is enriched. At a vehicle speed V5, the speed ration x is maximum, and accordingly, the torque level is limited below the curve B1 -B2. In order to obtain the torque larger than that, it is required that the speed ratio x is decreased while the engine speed is increased, or the speed ratio x is set to be maximum while the mixture is enriched. In view of the acceleration performance, the latter is advantageous, but in view of the fuel economy, the former is advantageous. Accordingly, whether the mixture is enriched so as to increase the torque or the speed ratio x is decreased so as to increase the torque, is depend upon a driver's taste or an environment around the vehicle.
In a Miller cycle engine in which the ratio between working and compression strokes is equal to or less than 1, that is, working/compression stroke ratio ≧1, the closing angle of an intake valve is adjusted so as to control the engine generated torque. Alternatively, the pressure of a supercharger is controlled so as to control the generated torque. If the closing angle is retarded, the compression stroke is decreased so as to increase the above-mentioned stroke ratio, and accordingly, the expansion energy can be effectively used so as to enhance the fuel economy. In order to increase the engine generated torque, it is required to increase the compression stroke. However, the fuel economy is accordingly lowered. Whether the speed ratio x is decreased so as to increase the wheel torque or the compression stroke is increased so as to increase the torque, depends upon the driver's taste or an environment around the vehicle at that time. Referring to FIG. 15, when the vehicle speed is constant, if the speed ratio x is large, the engine torque is large, but if the speed ratio x is small, the engine torque is small. Up to a value Te1, the engine can be driven, maintaining the stroke ratio constant, but it is required that the compression stroke is decreased upon shifting from Te1 to Te2. If the speed ration x is decreased, the engine can be driven, maintaining the stroke ratio constant in a range from F01 to F02 since the engine torque is less than the value Te1. However, if the speed ration x is excessively decreased, the engine speed is increased so as increase the fuel consumption. If the engine speed exceeds 3,000 rpm, the speed ratio is set to be large. If acceleration is desired, the speed ratio x is decreased, and if the fuel economy is essential, the value Te1 is set be as small as possible at an engine speed of 3,000 rpm.
In the Miller cycle engine, in order to increase the engine torque while maintaining the working and compression stroke ratio large, it is required to increase the supercharge pressure. Even though the compression stroke is small, a large volume is charged into the engine cylinder so as to increase the engine torque. However, since the compression work is increased. the speed ratio x is increased so as to decrease the desired engine torque until the engine speed becomes 3,000 rpm.
F=k(n.sub.e /V)*T.sub.e                                    EX 4
In a range A in FIG. 16, if it is operated with Te =Te1, it is required to set the engine speed at a value higher than 3,000 rpm. In this phase, the engine torque Te is increased up to Te2 so as to restrain the engine speed from increasing.
Explanation will be made hereinbelow a first embodiment of an engine which is used in the abovementioned drive system. Referring to FIG. 17, the engine 201 incorporates a piston 202 having a concave combustion chamber, an intake valve 203, an exhaust valve 204, a fuel injection valve 205, a spark plug 206, an intake pipe 207 in which an air cleaner 208 is located, and an exhaust pipe 209 in which a catalytic converter 210 for purifying nitrogen oxide is incorporated. The intake valve 203 is driven by a low load cam 211 and a high load cam 212. The exhaust valve 204 is driven by a cam 213. The cam 211 presses a rocker arm 214 while the cam 212 presses a rocker arm 215. In this arrangement, under a low load, a solenoid 216 is energized so as to connect the rocker arm 215 with the intake valve 203. The spark plug 206, the fuel injection valve 205 and solenoids 216, 217 are operated under the control of a control apparatus 218, a position (degree of depression) of an accelerator pedal 219 detected by a potentiometer 220, a speed of the engine detected by a speed sensor 221 and an air-fuel ratio of exhaust gas detected by an air-fuel ratio sensor 222 are delivered to the control apparatus 218.
The shapes of the high load cam 212 and the low load cam 211 are shown in FIG. 20. The low load cam 211 opens the intake valve 203 until a middle point of a compression stroke. On the contrary, the high load cam 212 which has the same shape as that of a conventional one, closes the intake valve 203 at the initiation of a compression stroke. Accordingly, the solenoids are switched so that the high load cam 212 is coupled to the intake valve 203 when the fuel volume is large but the low load cam 211 is coupled to the intake valve 203 when the fuel volume is less. Accordingly, the characteristic of air volume can be obtained as shown in FIG. 21. The exhaust valve 213 is closed at the end of an exhaust stroke, similar to a conventional one. Thus, the air volume is small as the fuel volume is small, and accordingly, it is possible to prevent the air-fuel ratio from increasing in a range where the fuel volume is small, as shown in FIG. 22 so as to stabilize the combustion even though the fuel volume is small. At this time, if the fuel volume is set to a value (a) as shown in FIG. 22, the air volume is set so as to prevent the air-fuel ratio from being lower than 16 since the nitrogen oxide emission becomes locally maximum around an air-fuel ratio of 16. Accordingly, the closing timing of the intake valve shown in FIG. 21 is set so as to allow the air-fuel ratio to satisfy the above-mentioned condition. Accordingly, as shown in FIG. 22, the increment of hydrocarbon emission in a range where the fuel volume is small, can be restrained, and as well, the increment of nitrogen oxide emission can be restrained, as shown in FIG. 22. The air-fuel ratio shown in FIG. 22 is detected by the air-fuel ratio sensor 222, and accordingly, if air-fuel. ratio approaches to a value 16 at the point a, the fuel volume is decreased or the closing timing of the intake valve is advanced so as to correct to the air-fuel ratio. Accordingly, the nitrogen oxide emission is prevented from increasing.
In the first embodiment shown in FIG. 17, the closing timing of the intake valve 203 is controlled so as to increase the air volume in a range where the fuel volume is large. Alternatively, the supercharge pressure may be increased while the closing timing is unchanged in order to increase the air-volume. In this case, the ratio between expansion and compression becomes 2 even in a range where the fuel volume is large, and accordingly, the specific fuel consumption can be decreased as a whole.
Referring to FIG. 25, if the air volume has a value G2, the injection timing becomes more and more negative as the fuel volume is increased. That is, since the compression dead center is set to zero, the injection timing is advanced up to a crank angle of -180 deg. that is, up to the initiation of a compression stroke.
In a Miller cycle engine, the air volume is lowered to, for example, a value G1 if the supercharge pressure is decreased. At this time, should the fuel injection timing be fixed with respect to the fuel volume, as a conventional one, the induction timing near a stoichiometric air-fuel ratio would be -90 deg. so that the mixing of air and fuel could not be promoted. On the contrary, according to the present invention, the injection timing is advanced to -180 deg. so as to promote the mixing of air and fuel, thereby it is possible to stabilize the combustion.
Referring to FIG. 26 which is a control flow chart for the present embodiment, at step 291, the engine speed is detected, and at step 292, a position of the accelerator pedal is detected. At step 293, a desired fuel volume is computed from both detected values. That is, the desired fuel is read from a table on which data obtained from a chart shown in FIG. 18 is mapped. Referring to FIG. 24, when the air volume is larger than a value F1, the air volume is set to a value G2, but when if it is less than the value F1, the air volume is set to a value G1. At step 293', the air volume may be set continuously with respect to the fuel volume. At step 294, an air-fuel ratio is detected by the air-fuel ratio sensor 222, and at step 295, an actual air volume is estimated from the detected value. As step 296, the air volume is corrected in accordance with the estimated air volume by adjusting the supercharge pressure, by adjusting the intake pipe pressure with the use of a throttle valve incorporated in the intake pipe 7, or by adjusting the closing timing of the intake valve. At step 297, with the use of a table as shown in FIG. 25, which gives an injection timing with respect a fuel volume, an air volume and an engine speed, the fuel timing is determined, and at step 298, an actual fuel injection is carried out. At step 299, similar to the determination of the injection timing at step 298, the ignition timing is determined with the use of a table which gives an ignition timing with respect to a fuel volume, an air volume and an engine speed, and at step 300, an actual ignition is carried out.
Referring to FIG. 30 which shows the NOx emission with respect to the air-fuel ratio, the later the injection timing, the larger the air-fuel ratio which exhibit a peak value of NOx. With the air-fuel ratio higher than a mark  on the curve, the combustion of the engine becomes unstable. Accordingly, the engine is operated on the right side of the mark . However, if the air-fuel ratio becomes small, the emission of NOx increases, and accordingly, the engine is driven adjacent to the mark . That is, the setting of the air-fuel ratio with respect to the injection timing, the setting of the injection timing with respect to the air-fuel ratio, or the setting of air-fuel ratio with respect to the fuel volume (refer to FIG. 19) is determined in accordance with empirical data as shown in FIG. 30. If the injection timing is delayed while the air-fuel ratio is unchanged, the emission of NOx increases. When the air-fuel ratio is decreased with respect to the fuel volume, if the advance control of the injection timing is delayed, the emission of NOx is delayed. However, in this embodiment, since the fuel injection valve 5 is electrically controlled, the injection timing is not delayed, and accordingly, it is possible to prevent the emission of NOx from increasing.
Referring to FIG. 31 which shows an arrangement of an engine and components therearond in the second embodiment, the engine in this embodiment is a gasoline type four cylinder MiIller cycle engine 310 having a cylinder head formed therein with an intake port 313 and an exhaust port 314 which are connected respectively to an intake pipe 320 and an exhaust pipe 330. Further, a fuel injection valve 380 and a spark plug 340 are provided in the cylinder head. Further, an intake valve 315 is provided in the intake port 313, and an exhaust valve 316 is provided in the exhaust port 316. The intake pipe 310 is provided therein with a throttle valve 321 for adjusting the flow rate of air flowing therethrough. Meanwhile, the exhaust pipe 330 is provided therein with a catalytic converter 331 for removing detrimental components from exhaust gas flowing therethrough. A water jacket 318 reserving therein cooling water is provided around the outer periphery of an engine cylinder 317. The water jacket 318 is connected with a radiator (which is not shown) through the intermediary of a pipe so that the cooling water is circulated between the water jacket and the radiator.
The intake pipe 320 incorporates an air-flowmeter 319 for detecting a mass flow rate A of air flowing therethrough. Meanwhile, the exhaust pipe 330 incorporates an exhaust gas thermometer 394 for detecting a temperature Tg of exhaust gas flowing therethrough. Further, the water jacket 318 incorporates a cooling water thermometer 393 for detecting a temperature Tw of cooling water flowing therethrough. The throttle valve 321 is provided thereto with a throttle opening degree meter 392 for detecting a degree thereof. The engine has a crankshaft (which is not shown) is provided thereto with a engine speed meter 395 for detecting a speed of the engine.
The air flowmeter 391, the throttle opening degree meter 392, the cooling water thermometer 393 and the exhaust gas thermometer 391 are connected to the control unit 390 which therefore receives detection signals from these meters.
In this embodiment, referring to FIG. 34, two intake valves 315 and two exhaust valves 316 (only one intake valve 315a and one exhaust valve 316 are depicted in the figure) are provided for each engine cylinder. The intake and exhaust valve drive mechanism 350 is adapted to operate these valves 315, 316 with appropriate timing. The intake and exhaust valve drive mechanism 350 has a cam shaft 351 coupled to the crankshaft (which is not shown) of the engine 310 through the intermediary of a timing chain, a cam 352 adapted to be rotated in association with the rotation of the cam shaft 351, rocker arms 353a, 353b making contact at one end thereof wit the peripheral surface of the cam 352 and at the other end thereof with the stem heads of the valves 315a, 316a, and rocker shafts 354a, 354b for swingably supporting the rocker arms 353a, 353b. The rocker arms 353a, 353b swing at one end thereof along the peripheral surface of the cam, and accordingly, press, at the other end thereof, the stems of the valves 315a 316a which are therefore operated. The lifts and the operation timing of the valves 315a, 316a can be adjusted by changing the profile of the cam 352. The operation timing of the valves will be described hereinbelow. Although a drive mechanism for the exhaust valves 316 is not shown in FIGS. 31 and 34, the basic structure thereof is similar to the drive mechanism for the intake valves 315 which is shown in FIG. 34.
In this embodiment, as shown in FIG. 34, the fuel injection valve 380 is arranged so as to inject fuel direct into the cylinder chamber 312 of the engine 310. In an intake port fuel injection system as is seen in a general gasoline type engine, fuel sticks to the inner surface of the intake pipe 320 and the upper surface of the intake valve 315, and as a result, fuel cannot be fed into the cylinder chamber by a desired volume at a desired time, and accordingly, the combustion in the cylinder chamber possibly becomes unstable. In particular, if the lift of the intake valve 315 is small (that is, less than 1.98 mm), fuel stagnating on the upper surface of the intake valve 315 discretely enters into the cylinder chamber, causing the combustion to be unstable, and accordingly, the tendency of unstable revolution of the engine is high. In view of the foregoing, in this embodiment, the fuel is injected direct into the engine cylinder so as to prevent the fuel from sticking to the inner surface of the intake pipe 320 and the upper surface of the intake valve 315. Further, in this embodiment, during an intake stroke in a low engine speed range, of two intake valves 315a, 315b, the one 315b is temporarily stopped, while the other 315a is opened so as to create a swirl flow in the cylinder chamber 312 in order to promote the combustion. As a result, as shown in FIG. 35, according to this embodiment, the engine speed during idle operation is remarkably stable.
The fuel injection valve 380 is composed of, as shown in FIG. 36, a valve element 386, a position adjuster 387, fuel passages 382, 383, a valve displacement space 385, and a valve casing 381 for housing the above-mentioned components. The fuel passages 382, 383 have one ene part formed therein with a fuel inlet port (which is not shown) and the other end part formed therein a fuel jet port 384. The valve displacement space 385 is formed intermediate of the fuel passages 382, 383, and fuel flows into the valve displacement space 385. That is, a part of the valve displacement space 385 serves a fuel passage. The passage (which will be hereinbelow denoted "valve space outlet side passage") 383 between the valve displacement space 385 and the fuel jet port 384 is formed in a cylindrical shape. The passage (which will be hereinbelow denoted "valve space inlet port side passage") 382 is bifurcated into two passages 382a 382b. One of these passages 382a, 382b, which will be hereinbelow denoted "wide angle atomization passage, is extended, perpendicular to the center axis C of the cylindrical outlet port side passage 383, and the other one of them, (which will be hereinbelow denoted "narrow angle atomization passage") 382b is extended in a direction having an obtuse angle to the center axis C of the outlet port side passage 383. The valve element 386 is located in the valve displacement space 385 so as to be movable among a valve closing position where the valve element 386 blocks the valve displacement space side port of the valve space outlet side passage 383, a wide angle atomization position (as shown in FIG. 36) where it opens the valve displacement space side port of the wide angle atomization passage 382a, but it blocks the valve displacement space side port of the narrow angle atomization passage 382b and a narrow angle atomization position (as shown in FIG. 37) where it opens the valve displacement space side port of the wide angle atomization passage 382a and the valve displacement space side port of the narrow angle atomization passage 382b. The position adjuster 387 has a small-sized stepping motor 387a which receives a control signal from the ECU 390, and a stopper 387b adapted to be driven by the stepping. motor 387a. The position adjuster 387 locates the valve element 386 at a desired position since the stopper 389b makes contact with the valve element 385. Specifically, the position adjuster 387 locates the valve element 386 at one of the above-mentioned valve closing position, wide angle atomization position and narrow angle atomization position in accordance with a signal from the ECU 390.
When the valve element 386 is located at the valve closing position, fuel cannot flow from the valve displacement space 385 to the outlet port side passage 383, and accordingly, no fuel is injected from the fuel injection valve 380. When the valve element 386 is located at the wide angle atomization position, only the wide angle atomization passage 382a which is extended in a direction perpendicular to the outlet port of the passage 383 is opened. Accordingly, when the fuel comes out from the wide angle atomization passage 382, the fuel is turned into a swirl flow in the valve displacement space 385, and is jetted from the fuel injection port 384 in a conical shape through the outlet port side passage 383. Further, when the valve element 386 is located at the narrow angle atomization position, both wide angle atomization passage 382a and narrow angle atomization passage 382b are opened. Since the narrow angle atomization passage 382b is extended in a direction having an obtuse angle to the outlet port side passage 383, the swirling power of the fuel having come out from the wide angle atomization passage 382a is decreased. Accordingly, the divergent atomization angle of the fuel jetted from the fuel injection port 384 is narrower when the valve element 386 is located at the narrow angle atomization position, than when it is located at the wide angle atomization position. Specifically, as shown in FIG. 36, when the valve element 386 is located at the wide angle atomization position, the divergent atomization angle of the fuel is 120 deg. while when it is located at the narrow angle atomization position, the divergent atomization angle is 60 deg.
The distributor casing 361 is formed therein with a plunger moving space 365, a fuel inlet port 362 communicated with the plunger moving space 365, and four fuel outlet ports 363a, . . . 363d (refer to FIG. 43) respectively communicated with the fuel injection valves 380a, . . . 380d. The casing fuel inlet port 262 is connected with a fuel pump which is not shown. The plunger 366 is cylindrical, and a main fuel passage 362 is formed at a position corresponding to the center axis of the plunger 366. One end part of the main fuel passage 367 is formed therein with a fuel inlet port 367 for leading fuel having flown into the plunger moving space 365 from the fuel inlet port 362 of the distributor casing 361, into the main fuel passage 367 of the plunger 366, and the other end part of the main fuel passage 367 is formed therein with a fuel discharge port 367b for returning fuel having flown into the main fuel passage 367 into a fuel tank (which is not shown). In an intermediate part of the main fuel passage 37, plunger fuel outlet ports 367c, 367d communicated with the casing fuel outlet ports 363a, . . . 363d are formed. As these fuel outlet ports 367c, 367d, a first fuel outlet ports 367c and a second fuel outlet ports 367d are present, and both ports 367c, 367d are symmetric with each other about the center axis of the plunger 366, and are slightly shifted from each other in the direction in which the center axis of the plunger 366 extends.
The plunger drive mechanism 370 is composed of a cam disc 371 fixed to one end part of the plunger 366, a roller 372 making contact with the outer surface of the cam disc 371 near to the outer periphery thereof, a roller support plate 373 for rotatably supporting the roller 372, a cam shaft 374 coupled to the crankshaft of the engine 319 through the intermediary of a timing belt of the like, and a connecting rod 375 having one end part which is coupled to cam shaft 374 so as to be movable in the direction of the center axis of the plunger 366, and the other end part fixed to the cam disc 371. The crankshaft of the engine 310 and the cam shaft 374 are connected together so that the cam shaft 374 is rotated by one revolution as the crankshaft of the engine 310 rotates by one revolution. Thus, when the crankshaft of the engine 310 rotates by one revolution, the plunger 366 is rotated by two revolutions about the center axis thereof by means of the cam shaft 374, the connecting rod 375 and the cam disc 371. Four convex parts 371a, 371b . . . are formed on the outer surface of the cam disc 371 near the outer periphery thereof. The roller 372 is arranged to make contact with these convex parts 371a, 371b . . . Accordingly, when the cam shaft 374 is rotated by one revolution, the cam disc 371 and the plunger 366 fixed to thereto are rotated by one revolution which reciprocates them by four times.
The fuel flow rate adjusting mechanism 368 is composed of a flow rate adjusting ring 368a which is annular so as to make contact with the outer periphery of the cylindrical plunger 366 and which is reciprocatable between a position where it blocks the plunger fuel discharge port 367b and a position where it opens the fuel discharge port 367b, a solenoid 368b for reciprocating the ring 368a and a connecting rod 368c connecting between the solenoid 368b and the flow rate adjusting ring 368a.
The fuel injection timing adjusting mechanism 376 is composed of an injection timing adjusting ring 377 which is annular so as to make contact with the outer periphery of the cylindrical plunger 366, and which is reciprocatable between a plunger first injection position where it opens the plunger first fuel outlet port 367c while it blocks the plunger second fuel outlet port 367d, and a plunger second injection position where it blocks the plunger first fuel outlet port 367c while it opens the plunger second fuel outlet port 367d, and a solenoid 378 for reciprocating the ring 377, and a connecting rod 379 connecting between the solenoid 378 and the injection timing adjusting ring 377. The injection timing adjusting ring 377 is formed therein with communication holes 377a . . . 377d which are communicated with the fuel outlet ports 363a . . . 363d at the first injection position, as shown in FIG. 43.
The fuel distributor 360 leads fuel from the casing fuel inlet port 362 into the plunger moving space 362 due to the reciprocation of the plunger 366 caused by the rotation of the cam shaft 374 while discharges fuel having flown into the plunger moving space 365, from the plurality of casing fuel outlet ports 363a . . . 363d by way of the main fuel passage 367 of the plunger 366, and the communication holes 377a . . . 377d in the fuel injection timing ring 377. Which one of these casing fuel outlet ports 363a . . . 363d discharges fuel is determined by a rotating angle of the plunger 366, relative to the casing 361. The fuel distributor 360 distributes fuel from the fuel tank into the #1 cylinder fuel injection valve 380a, the #3 cylinder fuel injection valve 380c, the #4 cylinder fuel injection valve 380d and the #2 cylinder fuel injection valve 380b, successively in the mentioned order.
The volumes of fuel discharged from the casing fuel outlet ports 363a, . . . 363d, are adjusted by the fuel flow rate adjusting mechanism 368. Fuel having flown into the plunger main fuel passage 367 from the plunger fuel inlet port 367a can flow out from the plunger fuel discharge port 367b, in addition to the plunger fuel outlet ports 367c,. . . 367d. Accordingly, the fuel discharged from the plunger fuel discharge port 367b is adjusted by suitably moving the flow rate adjusting ring 368a of the fuel flow rate adjusting mechanism 366, and accordingly, the flow rate of fuel discharged, outside of the casing 361, from the plunger fuel outlet ports 367c, 367d through the casing fuel outlet ports 363a, . . . 363d is indirectly adjusted. It is noted that the fuel discharged from the plunger fuel discharge port 376b, is returned into the fuel tank.
The feed timing of fuel into each of the fuel injection valves 380 from the fuel distributor 360 is adjusted by the fuel injection timing adjusting mechanism 376. For example, as shown in FIGS. 41 to 44, when the plunger first fuel outlet port 367c is aligned with the #1 cylinder casing fuel outlet port 363a, and when the fuel injection timing adjusting ring 377 is aligned with the first injection position, the plunger first injection port 367c is communicated with the #1 cylinder casing fuel outlet port 363a through the communication hole 377a in the fuel injection timing adjusting ring 377. Accordingly, the fuel in the main fuel passage 367 of the plunger 366 is fed into the #1 cylinder fuel injection valve 380a through the plunger first fuel outlet port 367c, the communication hole 377a in the ring 377, and the #1 cylinder casing fuel outlet port 363a. Further, even though the plunger first fuel outlet port 367c is aligned with the #1 cylinder casing fuel outlet port 363a, if the fuel injection timing adjusting ring 377 is located at the second injection position, as shown in FIG. 34, the plunger first fuel outlet port 367c is blocked by the injection timing adjusting ring 377 while the plunger second fuel outlet port 367d is opened so that the plunger second fuel outlet port 367d is communicated with the #4 casing fuel outlet port 363d. Accordingly, the fuel in the main fuel passage 367 of the plunger 366 is fed into the #4 cylinder fuel injection valve 380d by way of the plunger second fuel outlet port 367d and the #4 cylinder casing fuel outlet port 363d. Thus, the fuel is not fed into the #1 cylinder fuel injection valve 380a but into the #4 cylinder fuel injection valve 380d at the timing of feeding fuel into the #1 cylinder fuel injection valve 380a by moving the injection timing adjusting ring 377. In other words, the phase of the fuel injection can be changed by 180 deg., as shown in FIG. 44 by actuating the fuel injection timing adjusting mechanism 376.
In this embodiment, the working stroke is set to be longer than the compression stroke by controlling the opening and closing timing of the intake valve 315. Specifically, at first the intake valve 315 is opened while the piston 311 descends so that air is introduced into the cylinder chamber 312 (refer to FIG. 47a, and then, the piston 311 comes to the bottom dead center (refer to FIG. 47b). Thereafter, the intake valve 315 is closed slightly after piston 311 slightly ascends (refer to FIG. 47c). The compression stroke extends during the period from the time when the intake valve 315 is closed, to the time when the piston comes to the top dead center. The ignition of fuel is carried out just before the time when the piston comes to the top dead center (refer to FIG. 47d). When the piston 311 comes up to the top dead center, it is depressed by explosion of the fuel (refer to FIG. 47e). The working stroke extends during the period in which the piston 311 moves from the top dead center to the bottom dead center (refer to 47f). The exhaust valve 316 is opened just before the time when the piston 311 comes to the bottom dead center. The piston 311 again initiates its ascent so that exhaust gas is discharged into the exhaust pipe 330 from the cylinder chamber 312 (FIG. 47g).
FIG. 39 shows the relationship between the injection timing and the density of hydrocarbon in exhaust gas for each fuel divergent atomization angle, which is exhibited by the fuel injection valve 380 in this embodiment, and which has been explained hereinabove. In this figure, the injection timing at zero deg. on the abscissa corresponds to the dead center in a compression stroke.
As shown in this figure, if the fuel divergent atomization angle is set to 120 deg., the density of hydrocarbon in exhaust gas increases as the injection timing is retarded (that is, the injection timing is changed in the direction approaching the top dead center (0 deg.)), and accordingly, the density of hydrocarbon becomes greatest when the injection timing is set to a crank angle of about -100 deg. from the top dead center (0 deg.). If the injection timing is retarded further, the density of hydrocarbon is contrarily decreased. Further, if the fuel divergent atomization angle is set to 60 deg., the density of hydrocarbon is lower than that obtained by setting the fuel divergent atomization angle to 120 deg., and is not substantially changed even though the injection timing is retarded up to a crankangle of about -40 deg. from the top dead center (0 deg.). As mentioned above, the reason why the density of hydrocarbon is higher at 120 deg. of fuel divergent atomization angle than at 60 deg. of fuel divergent atomization angle, is such that the volume of fuel sticking to the wall surface of the cylinder 17 is greater at 120 deg. of fuel divergent atomization angle than at 60 deg. of fuel divergent atomization angle. Further, after the -injection timing is retarded to about -40 deg. from the dead center (0 deg.) in compression stroke, the reason why the density of hydrocarbon is higher at 60 deg. of fuel divergent atomization angle than at 120 deg. of fuel divergent atomization angle, is such that the piston 311 comes near to the fuel injection valve 380 by retarding the injection timing near to the top dead center in compression stroke, and accordingly the quantity of fuel sticking to the top surface of the piston 311 is greater at 60 deg. of fuel divergent atomization angle than at 120 deg. of fuel divergent atomization angle.
Thus, if the injection timing is advanced, and specifically if the fuel is injected into the cylinder before a crankangle of about -40 deg. from the top dead center (0 deg.) in compression stroke, the fuel divergent atomization angle is set to 60 deg. Meanwhile, if the injection timing is retarded, and specifically, if the fuel is injected into the cylinder after an crankangle of about -40 deg. from the top-dead center (0 deg.) in compression stroke, the fuel divergent atomization angle is set to 120 deg. so as to decrease the density of hydrocarbon in exhaust gas as far as possible.
In summary, as shown FIG. 44, during partial load operation, the injection timing is retarded while the atomization angle is set to 120 deg. Meanwhile during high load operation, the injection timing is advanced while the atomization angle is set to 60 deg. Accordingly, the density of hydrocarbon in exhaust gas can be lowered, and stable combustion can be ensured. In this embodiment, in order to carry out the abovementioned control, the ECU 390 instructs the fuel injection timing adjusting mechanism 376 in the fuel distributor 360 to retard the fuel injection timing and instructs the valve position adjuster 357 in the fuel injection valve 380 to set the atomization angle to 120 deg. when the fuel injection volume which is determined in accordance with an air flow rate detected by the air flowmeter 391 and an opening degree of the throttle valve 321 detected by the throttle opening degree meter 392 and the like, is less than a predetermined value. Further, the ECU 390 instructs the fuel injection timing adjusting mechanism in the fuel distributor 360 to advance the fuel injection timing by a crankangle of about 180 deg. and instructs the valve position adjuster 357 in the fuel injection valve 380 to set the atomization angle to 60 deg. when the fuel injection volume determined by the CPU 390 itself exceeds a predetermined value.
It has been found from experiments made by the applicants, that the atomization angle is satisfactorily set to a value larger 100 deg. during partial load operation, and to a value smaller than 90 deg. during high load operation. Further, in this embodiment, when the valve element is located at the narrow angle atomization position, the valve displacement space side port of the wide angle atomization passage 382a and the valve displacement space side port of the narrow angle atomization passage 382b are opened. However, it is possible to close the valve displacement side port of the wide angle atomization passage 382a when the valve displacement space side port of the narrow angle atomization passage is opened.
Metal ion exchange zeolite catalyst 331a is located on the engine 310 side of the catalytic converter 331 connected to the exhaust pipe 330, and platinum and alumina group catalyst is located on the exhaust port side thereof. The metal ion exchange zeolite catalyst 331a has such characteristics that its low temperature activity is high, but its NOx selective reduction activity is low. Further, the platinum and alumina group catalyst has such that the low temperature activity is low but the NOx selective reduction activity is high. Accordingly, in an operating range where the density of hydrocarbon (HC) is high during high engine speed and high load operation so that, as previously explained with reference to FIG. 40, the density of hydrocarbon (HC) is low during high engine speed, and accordingly, the efficiency of nitrogen oxide-nitrogen conversion of catalyst tends to be decreased, the platinum alumina group catalyst 318 having a high NOx selective reduction activity mainly becomes effective if the catalytic environment temperature is high. On the contrary, in an operating range where the density of HC is high during low engine speed and low load operation so that, as mentioned above, the efficiency of nitrogen oxide-nitrogen conversion of the catalyst is high, the metal ion-exchange zeolite which is active even though the catalytic environment temperature is low is mainly effective. It is noted that the reason why the density of HC is low during high engine speed and high load, is such that the oxidative reaction is promoted since the temperature of exhaust gas is high during a process in which HC is exhausted from the cylinder chamber 312 into the exhaust pipe 330. Further, the reason why the density of HC is high during low engine speed and low load operation, oxidative reaction is not promoted since the temperature of exhaust gas is low, and accordingly, HC is directly exhausted as it is.
HC exhausted upon a start of the engine 310 is mainly adsorbed by the metal ion-exchange zeolite catalyst 331a. When the temperature of the catalytic converter 331 is raised by exhaust gas, HC adsorbed to the metal ion-exchange zeolite catalyst 331a is separated away, and is oxidized by the platinum alumina group catalyst 331b. Usual platinum alumina group catalyst is likely to produce nitrogen dioxide N2 O when HC is not oxidized at a low temperature. In order to evade this problem, the platinum alumina group catalyst in this embodiment is added therein with palladium or the like in order to enhance the catalytic activity at a low temperature. Further, in order to enhance the conversion efficiency of nitrogen dioxide N2 O during a start of the engine, the fuel injection timing is retarded so as to increase the temperature of exhaust gas. Alternatively, the ratio between working and compression strokes may be changed during Miller cycle, so as to raise the temperature of exhaust gas.
Referring to FIG. 36, just after a start of the engine, since the temperature of exhaust gas is low, HC and NOx are adsorbed onto the metal ion-exchange zeolite catalyst 331a. During this period, the fuel injection timing is retarded so as to increase the temperature of exhaust gas in order to restrain generation of NOx as small as possible. When the temperature of exhaust gas is higher and higher, HC and NOx adsorbed to the metal ion-exchange zeolite catalyst 331a are gradually converted into H2 O, CO2, N2. At this time, NOx in the exhaust gas by HC which is also adsorbed. When HC adsorbed to the metal ion-exchange zeolite catalyst 331 runs out, the injection timing is advanced so as to increase the density of HC in exhaust gas. Further, if the temperature becomes high, the platinum and alumina group catalyst is mainly effective.
A fuel injection valve 400 in a variant form shown FIG. 48, comprises a valve element 406, a position adjuster 407 for adjusting the position of the valve element 406, and a valve casing 401 formed therein with fuel passages 402, 403 and a valve displacement space 405, and incorporating the valve element 406 and the adjuster 407. The fuel passages 402, 403 are formed at their one end with fuel inlet ports (which are not shown), and at their the other end with fuel jet ports 404. The valve displacement space 405 is formed intermediate of these fuel passages 402, 403, and the fuel also flows into this valve displacement space 405. A plurality of passages 403 (which will be hereinbelow denoted "space outlet port side passages") are formed between the valve displacement space 405 and the fuel jet ports 404. One group (which will be hereinbelow denoted "narrow angle atomization passages") of these space outlet port side passages) extend in a direction having an angle of 30 deg. to the injection center axis C, and the group of the remaining passages 403a (which will be hereinbelow denoted "wide angle atomization passages") extend in a direction having an angle of 60 deg. to the injection center axis C).
The valve element 406 comprises a valve end part 406b adapted to block the valve displacement space side ports of the wide angle atomization passages 403a and the valve displacement space side ports of the narrow angle atomization passages 403b, and a body 406a having a front end part formed thereto with the valve end part 406b. The valve displacement space 405 is composed of a valve end part displacement space 405b into which only the valve end part 406b of the valve element 406 enters, and a body displacement space 405a into which the body 406a of the valve element 406 is fitted. The valve casing 401 is formed therein with a valve seat 401a at the boundary between the valve end displacement space 405b and the body displacement space 405a.
The valve element 406 is located in the valve displacement space 405 so as to be movable among three positions, that is, a valve closing position where fuel does not flow into the valve end displacement space 405b from the body displacement space 405 (as shown in FIG. 48), a wide angle atomization position where the valve displacement space side ports of the wide angle atomization passages 403a are opened while the valve displacement side ports of the narrow angle atomization passages 403b are closed (as shown in FIG. 49), and a narrow angle atomization position where the valve displacement side ports of the wide angle atomization passages 403a are closed but the valve displacement side ports of the narrow angle atomization passages 403b are opened (as shown in FIG. 50).
The position adjuster 409 comprises a small size stepping motor 407a receiving a signal from the ECU 390, and a stopper 407b operated under the drive of the stepping motor 407a. In the position adjuster 407, the stopper 407b makes contact with valve element 406 so as to locate the valve element 406 at a desired position. Specifically, the position adjuster 407 locates the valve element 406 at one of the abovementioned valve closing position, and wide and narrow angle atomization positions in accordance with a signal from the ECU 490.
Referring to FIG. 48, when the valve element 406 is located at the valve closing position so as to make contact with the valve seat 401a of the valve casing 401, fuel cannot flow into the valve end displacement space 405b from the body displacement space 405a, and accordingly fuel cannot be injected from the valve 400. Further, as shown in FIG. 49, when the valve element 406 is slightly lifted up so as to be located at the wide angle atomization position where the valve end part 406b of the valve element 406 blocks the valve displacement space side ports of the narrow angle atomization passages 403b, fuel flows from the valve displacement space 405 into the wide angle atomization passages 403a and is then jetted from the fuel outlet ports 404a at the ends of the passages 403a. At this time, the fuel divergent atomization angle at this time is set to 120 deg. Further, as shown in FIG. 50, the valve element 406 is further lifted up and is located at the narrow angle atomization position where the valve end part 406b of the valve element 406 blocks only the valve displacement space side ports of the wide angle atomization passages 403a, the fuel flows from the valve displacement space 405 and through the narrow angle atomization passages 403b, and then are jetted from the fuel outlet ports 404b at their ends. At this time, the fuel divergent atomization angle is set to 60 deg.
In this variant form, the fuel injection valve 410 has a spherical valve element 412 is used as shown in FIG. 53. This spherical shape valve element 412, is flexibly connected to a piston 417 through the intermediary a pin 416 by a flexible thin connecting rod 415. Accordingly, even though an error such as an eccentricity caused by machining, is present at a valve seat 411a or the like of a valve casing 411, this error can be absorbed by this flexible connecting rod 415. When a valve element 419a of a solenoid valve 419 is lifted up, the pressure of the pressure chamber 418 is lifted up, and the pressure in a pressure chamber 418 is lowered. Accordingly, a piston 419 is pushed up by the force of a spring 414, causing the spherical valve element 412 to ascend, and the space between the valve seat 411a and the spherical valve element 412 is obtained so that the fuel is injected. The spherical valve element 412, a guide 413, the connecting rod 415 and the piston 417 are movable, and are all small-sized, lightweight and high responsive so that two times of injection per one cycle can be made. When the valve element 419a of the solenoid valve 419 is closed, the pressure of the pressure chamber 418 is increased so that the piston 417 is depressed, and as a result, the spherical body 412 is closed.
In this arrangement, although the solenoid valve 419 is used for adjusting the pressure in the pressure chamber 418, a laminated type piezoelectric element can be used, instead of the solenoid valve 418. As shown in FIG. 54, the laminated type piezoelectric element 420 is provided in a part of the wall surface forming the pressure chamber 418. Since the fuel flows into the pressure chamber 418, the pressure in this chamber 418 is high. Accordingly, the piston 417 is depressed, and accordingly, the spherical valve element 412 is pressed against the seat 411a of the casing 411 while the pressure element 420 is pressed. Thus, when the piezoelectric element 420 is pressed, a capacitor 421 is charged. At an end of a compression stroke, when a switch 422 is closed, the charged capacitor 421 is discharged so that the piezoelectric element 420 contracts, and accordingly, the pressure in the pressure chamber 112 is lowered. As a result, the piston 417 is slightly raised so as to slightly lift up the spherical valve element 112, and accordingly, the fuel is injected.
As shown in FIG. 55, a fuel induction valve 430 in this variant form comprises a valve casing 431 formed therein with a valve displacement space 432 and a solenoid storage part 433, a valve element 131 adapted to move in the valve displacement space 432, a spring 436 urging the valve element 430 in a valve closing direction, an armature 435 fixed to the end part of the valve element 434, a solenoid 437 for moving the valve element 434 together with the armature 435, a solenoid drive circuit 440 for driving the solenoid 437, and a fuel filter 438 for removing foreign matter in fuel flowing into the valve casing 431. The solenoid 437 is energized and deenergized so as to open and close the valve in order to move the valve element 434. Specifically, when the solenoid 437 is energized, the valve element 434 is opened, overcoming the force of the spring. Meanwhile when the solenoid 437 is deenergized, the valve 434 is closed by the force of the spring.
The distributor 550 in this modified form is the one in which the injection timing adjusting ring 357 in the above-mentioned distributor 360 is modified. In this modified form, the injection timing adjusting ring 357 is formed therein with communication holes 357a, 357b . . . so that when a casing fuel outlet port 363a which is one of two casing fuel outlet ports 363a, 363d symmetrical with each other about the center axis of a plunger is communicated with a plunger first fuel outlet port 367c, the other casing fuel outlet port 363d is communicated with a plunger second fuel outlet port 367d. Accordingly, when fuel flows from one of two casing fuel outlet ports which are located at positions symmetric with each other about the center axis of the plunger, fuel flows also from the other thereof. Specifically, if the fuel flows out from a #1 engine cylinder casing fuel outlet port 363a, fuel flows also from a #4 engine cylinder casing fuel outlet port 363d.
By the way, in the case of a four cylinder engine, when a #1 engine cylinder is shifted from an exhaust stroke into an intake stroke, a #4 engine cylinder is shifted from a compression stroke into a working stroke, as shown in FIG. 44. Meanwhile, when the #1 engine cylinder is shifted from a compression stroke into a working stroke, the #4 engine cylinder is shifted into an exhaust stroke into an intake stroke. Accordingly, as shown in FIG. 58, in such a case that the main fuel injection and the ignition fuel injection are carried out, when the main fuel injection is carried out in the #1 engine cylinder, the ignition fuel injection is carried out in the #4 engine cylinder. Meanwhile, when the main fuel injection is carried out in the #4 engine cylinder, the ignition fuel injection is carried out in #1 engine cylinder. Accordingly, in this modified form, fuel flows from the #1 engine cylinder casing fuel outlet port 363a while fuel flows also from the #4 engine cylinder fuel outlet port 363d.
Incidentally, when main fuel (in a large fuel volume) is fed from the casing fuel outlet port 436s as one of both outlet ports, it is required to feed ignition fuel (in a small fuel volume) from the other casing fuel outlet port 363d. Accordingly, the ring 457 is formed therein with the communication holes 457a, 457d, . . . so that even in a condition in which the one casing fuel outlet port 363a and the plunger first fuel outlet port 363a are completely communicated with each other, the other casing fuel outlet port 363d is in half communicated with the plunger second fuel outlet port 367d.
If the fuel flows from the casing fuel outlet port 363d which is one of two casing fuel outlet ports 363a, 363d arranged symmetric with each other about the center axis of the plunger, simultaneously with flowing of fuel from the other casing fuel outlet port 363a, and if the fuel injection valves have different flow characteristics, either the main fuel injection volume or the ignition fuel injection volume cannot be fed in their respective desired values due to their different flow characteristics. That is, for example, if one of the fuel injection valve has a large pressure loss while the other one of them has a small pressure loss, when the fuel is fed, from one and the same fuel supply source, simultaneously into the fuel injection valves, fuel having a volume less than the desired value flow from one of the fuel injection valves meanwhile fuel having a volume larger than the desired value flows out from the other one of them. Accordingly, a second modified form of the distributor which can feed fuel in desired volumes into the fuel injection valves even though these fuel injection valves have different flow characteristics, will be hereinbelow explained with reference to FIG. 60.
In this distributor 460 shown in FIG. 60, a plunger second fuel outlet port 467a is extended in a direction having an angle which is not 180 deg. to the plunger first fuel outlet port 467c but is slightly smaller than 180 deg. around the plunger center axis as a center. It is noted that the distributor in this second modified form is substantially the same as the distributor in the first modified form, excepting the above-mentioned arrangement. With these plunger fuel outlet ports 467c, 467d, when fuel is fed into the #4 engine cylinder, no fuel is fed into the #1 engine cylinder. However, slightly later, the fuel is fed into the #1 engine cylinder. Accordingly, fuel is not fed simultaneously into the fuel injection valves from the distributor 460, and therefore, fuel in a substantially desired volume, can be fed into each of the fuel injection valves even having different flow characteristics.
The fuel pump 470 in this embodiment has a pump casing 471, a piston 473 adapted to reciprocate in the pump casing 471, and a piston drive mechanism 473. This piston drive mechanism 473 comprises a cam shaft 474 which is coupled to a crankshaft through the intermediary of a timing belt or the like, an advance cam 475a and a retardation cam 475b which are fixed to the cam shaft 474 so as to be rotated in association with the rotation of the cam shaft 474, and an advance cam follower rod 476a arranged so as to make contact with the outer peripheral surface of the advance cam 457a, a retardation cam follower rod 476b arranged so as to make contract with the outer surface of the retardation cam 475b, a swingable rod 476c making contact with an end part of the piston 472, a support pin 478 for swingably supporting the advance cam follower rod 476a, the retardation cam follower rod 476b and the swingable rod 476c, a timing change-over pin 479 for swinging the swingable rod 476c in response to one of the two cam follower rods 476a, 476b, and a solenoid (which is not shown) for moving the timing change-over pint 479. The advance cam follower rod 467a, the retardation cam follower rod 476b and the swingable rod 467c are arranged in parallel with one another, and are supported at their one end part by the support pin 478. The other end part of the swingable rod 476c is formed therein with a timing change-over pin through-hole 477c through which the timing change-over pin 479 pierces, the other end part of the advance cam follower rod 476a and the other end part of the retardation cam follower rod 476b are formed respectively therein change-over pin fitting parts 477a, 477b in which opposite end parts of the timing change-over pin 479 are fitted. The timing change-over pin 479 is always inserted in the timing change-over pin through-hole of the swingable rod 476c, but is fitted, at its either one of the end parts, in either one of the fitting part 477a of the advance cam follower rod 467a and the retardation cam follower rod 467b in accordance with its own position.
When the cam shaft 474 is rotated in association with the rotation of the crankshaft, the advance cam 475a and the retardation cam 475b fixed to the cam shaft 474 are rotated. In association with the rotation of these cams 475a, 475b, the cam follower rods 476a, 476b making contact with the outer peripheral surfaces of these cams 475a, 475b are swung around the support pin 478 as a center in accordance with shapes of the cams with which they make contact. If the timing change-over pin 479 is fitted in the fitting part 477a of the advance cam follower rod 476a at this time, the swingable rod 467c is also swung in association with the swinging of the advance cam follower rod 476c. Alternatively, if the timing change-over pin 479 is fitted in the fitting part 477b of the retardation cam follower rod 476b at this time, the swingable rod 467c is swung in association with the swinging of the retardation cam follower rod 476b. Accordingly, the piston 472 is reciprocated in association with the swinging of the swingable rod 476c.
As mentioned above, when the solenoid is energized in response to the ECU 90 so as to move the timing change-over pin 479, the piston can be actuated, following the actuation of one of the cams 475a, 475b, and accordingly, the fuel injection timing can be changed as shown in FIG. 63.
1. A method of controlling an internal combustion engine, comprising the step of controlling a volume of fuel injected from a fuel injection device incorporating an injection port in a combustion chamber of the internal combustion engine, fuel injection timing, and opening and closing timing of an intake valve for controlling a volume of air to be burnt within said combustion chamber so that said intake valve is opened at a time of initiating suction stroke, and is closed at a time of initiating a compression stroke, wherein said intake valve is opened substantially at a middle of the suction stroke and is closed substantially at a middle of the compression stroke when an engine load is lower than a predetermined value.
US08431028 1994-04-28 1995-04-28 Control apparatus for drive system composed of engine and transmission Expired - Fee Related US6058348A (en)
US09450135 Continuation US6298300B1 (en) 1994-04-28 1999-11-26 Control apparatus for drive system composed of engine and transmission
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US6298300B1 (en) 2001-10-02 grant
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNORS:OHYAMA, YOSHISHIGE;FUJIEDA, MAMORU;NOGI, TOSHIHARU;AND OTHERS;REEL/FRAME:007480/0980