 |
Description  |
|
|
TECHNICAL FIELD
The present invention relates to an engine control system for an engine
equipped with a variable valve actuating mechanism, such as a variable
valve timing system, and a variable capacity supercharger, such as a
turbocharger having moveable vanes to vary the cross sectional area of the
exhaust gas passage leading to a turbine wheel, and in particular to such
an engine control system which permits the advantages of both a variable
valve actuating mechanism and a variable capacity supercharger to be fully
utilized by harmonious combination of the two variable elements.
BACKGROUND OF THE INVENTION
A valve operation switching unit for improving a volume efficiency of the
combustion chambers over a wide operation range by changing at least
either the angular interval of opening the intake valves and/or exhaust
valves for each cylinder or the lift of the valves is proposed for
instance in Japanese patent laid open publication No. 63-16111.
A variable capacity supercharger offering an optimum supercharge pressure
over a wide operating range with a high responsiveness by varying the A/R
ratio of an exhaust passage leading to a turbine wheel by means of a flap
or a plurality of vanes is proposed in Japanese patent laid open
publication No. 62-282128.
According to such a variable capacity supercharger, since a supercharge
pressure which is suitable for each operating condition can be arbitrarily
and accurately obtained, an even further improvement can be achieved
particularly by combining a valve operating condition switching unit and a
variable capacity supercharger.
In low speed range it is possible to increase the speed of intake flow
directed to the combustion chambers by reducing the angular interval of
opening the valves and/or the valve lift, but this tends to limit the
intake flow rate as the rotational speed of the engine increases.
Conversely, by increasing the angular interval of opening the valves
and/or the valve lift in high speed range, the volume efficiency of the
engine intake improves as the rotational speed of the engine increases.
Therefore, if a variable capacity supercharger used in conjunction with a
valve operating condition switching unit is controlled in the same way as
if it were used for an engine without any such valve operating condition
switching unit, it would not be possible to obtain an optimum performance
of the engine in all of its operating range.
In particular, since the change in the movement of the valves during each
cycle of engine operation will affect conditions of the intake passages
(such as the resonance frequency of the intake passage, the volume
efficiency of engine intake, etc.), it is advantageous to adapt the mode
of controlling the supercharger to such changes. For instance, in an
engine using a valve timing adjusting system which switches over valve
timing in step-wise fashion according to the change in the rotational
speed of the engine, as the rotational speed of the engine is increased,
the torque output reaches a peak value and then gradually diminishes
before the valve timing is switched over to from a low speed mode to a
high speed mode. This decline in the torque output between the point of
the torque peak and the point of valve timing switch over may be felt by
the operator as a torque dip, and it is desired to remove such a torque
dip.
As an additional consideration, such a complex control action should not
involve any undue delay as such a delay will seriously impair the
commercial value of the vehicle on which the engine is mounted. However, a
high responsiveness of an engine must be accompanied by a sufficient
control stability.
Further, in view of the complexity of the overall control system, it is
desired to have a fail safe feature to be incorporated into the system.
BRIEF SUMMARY OF THE INVENTION
Based upon such and other recognitions, a primary object of the present
invention is to provide an engine control system which can achieve a
maximum improvement in the performance of an engine which incorporates
both a variable capacity supercharger and a variable valve actuating unit.
A second object of the present invention is to provide such an engine
control system which combines a fast response and a stable control action.
A third object of the present invention is to provide such an engine
control system which can eliminate the occurrence of a torque dip which
may occur if the valve actuating mechanism is varied in step-wise fashion
and the rotational speed is increased close to a point of a step-wise
varying action.
A fourth object of the present invention is to provide such an engine
control system which is protected from operating in any undesirable
fashion even in case of a system failure.
These and other objects of the present invention can be accomplished by
providing an engine control system, comprising: valve actuating condition
varying means for varying a state of a valve actuating mechanism of an
engine; capacity varying means for varying a supercharge capacity of a
variable capacity supercharger; and control means for controlling a valve
actuating condition varying operation of the valve actuating condition
varying means and a capacity varying operation of the capacity varying
means according to an operating condition including at least a rotational
speed of the engine; the control means carrying out the capacity varying
operation in dependence upon an operating condition of the valve actuating
mechanism.
Thus, an optimum control of supercharge pressure can be accomplished in
response to an operating condition of valves. In particular, if the
control means increases the supercharge pressure in case of high speed
operation of the engine, the high speed performance of the engine can be
improved. Conversely, if supercharge pressure is decreased according to
the increase in the torque output of the engine owing to the switching of
the valve operating condition, burden on the engine can be reduced in its
high speed operating condition.
According to a preferred embodiment of the present invention, the control
means carries out the capacity varying operation of the supercharger by an
open loop control process at least when the valve actuating means is
adapted for a low speed operating condition of the engine, and by a closed
loop control process at least when the valve actuating means is adapted
for a high speed operating condition of the engine. Typically, the open
loop control process consists of a map control which determines the
supercharge capacity according to a rotational speed of the engine and an
opening angle of a throttle valve or an intake negative pressure.
According to the preferred embodiment of the present invention, in order
that an undesirable dip in the output property of the engine due to the
changes in the intake conditions of the engine due to the activation of
the valve actuating condition varying means may be avoided, the control
means may change the supercharge pressure of the supercharger at least in
a low speed range from a normally controlled level to a boosted level
according to a change rate of a level of supercharge pressure, and/or a
change rate of a rotational speed of the engine. Normally, the supercharge
pressure output of the supercharger should be increased when a decline in
the supercharge pressure is detected as the rotational speed of the engine
is increased because such decline in the supercharge pressure means a
decline in the volume efficiency of the engine intake, and, in order to
remove a torque dip resulting therefrom, the supercharge pressure output
of the supercharger should be increased to compensate for the reduction in
the volume efficiency.
BRIEF DESCRIPTION OF THE DRAWINGS
Now the present invention is described in the following in terms of a
specific embodiment with reference to the appended drawings, in which:
FIG. 1 is an overall structural diagram of the control system for an engine
according to the present invention;
FIG. 2 is a structural view of a part surrounding a valve actuating
mechanism;
FIG. 3 is a view illustrating the mechanism of the variable capacity
turbocharger;
FIGS. 4a through 4d are flow charts of the control program which is related
to the switch over of valve timing;
FIGS. 5a and 5d are flow charts of the control program which is related to
the adjustment of supercharge pressure; and
FIGS. 6 through 11 are flow charts of the subroutines which are related to
the above mentioned programs.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
FIG. 1 shows an overall structure of the intake and exhaust system of an
engine to which the present invention is applied. In this engine main body
1, for instance consisting of an in-line four-cylinder engine, an intake
manifold 3 leading to the intake port 2 of each cylinder is connected to
an intake tube 4, a throttle body 5, an intercooler 6, a compressor unit 8
of a variable capacity supercharger 7, and an air cleaner 9, in that
order. An exhaust manifold 11 leading to the exhaust port 10 of each
cylinder is connected to a turbine unit 12 of the variable capacity
supercharger 7, and a catalytic converter 13.
A valve mechanism 14 provided for controlling the intake of mixture and the
exhaust of combustion gas into and out of the combustion chamber of each
cylinder can change valve timing in a stepwise fashion by controlling
hydraulic pressure produced from an oil pump 15 actuated by the engine
main body 1, by way of a solenoid valve 16 and a switching control valve
17.
The variable capacity supercharger 7 can continually vary a cross sectional
area of a passage for exhaust gas leading to its turbine unit 12 by way of
an actuator 18 which is actuated by supercharge pressure P.sub.2
immediately downstream of the compressor unit 8 or intake negative
pressure immediately downstream of the throttle valve 5 to vary the
supercharging capacity of its compressor unit 8. This turbocharger 7,
along with the intercooler 6, is cooled by cooling water which is
circulated by a water pump 19 actuated by the engine main body 1 through a
cooling water system including a radiator 20, which is separate from the
cooling water system for the engine main body 1.
The engine 1 is equipped with an electronic control circuit 21 for
controlling the amount of fuel injection, valve timing and supercharge
pressure for the engine 1.
The electronic control circuit 21 receives an oil pressure signal O.sub.P
from an oil pressure switch 22 of a normally closed type provided in the
switching control valve 17, an O.sub.2 signal from an oxygen concentration
sensor 23 provided in the exhaust manifold 11, a rotational speed signal
N.sub.E from an engine rotational speed sensor 24, a water temperature
signal T.sub.W from a cooling water temperature sensor 25 provided in the
water jacket of the engine main body 1, a parking/neutral signal PN
indicating the shift position of the automatic transmission system 26 to
be either in a parking or neutral range, an intake temperature signal
T.sub.A and an intake pressure signal P.sub.B from an intake temperature
sensor 27 and an intake pressure sensor 28, respectively, provided in a
part of the intake passage 4a downstream of the throttle body 5, a
throttle valve opening angle signal .theta..sub.TH from a throttle opening
angle sensor 29, a supercharge pressure signal P.sub.2 from a supercharge
pressure sensor 30 provided in a part of the intake passage 4b downstream
of the compressor unit 8, an atmospheric pressure signal P.sub.A from an
atmospheric pressure sensor 31 provided in a part of the intake passage 4c
extending between the air cleaner 9 and the compressor unit 8 of the
turbocharger 7, and a vehicle speed signal V from a vehicle speed sensor
32. Output signals from this electronic control circuit 21 control the
operations of solenoid valves for switching over valve timing, fuel
injection valves 33 for injecting fuel into the intake ports 2, and
solenoid valves 34 and 35 for controlling the supercharge pressure P.sub.2
and the intake negative pressure P.sub.B for actuating the actuator 18 for
varying the supercharger capacity.
The valve actuating mechanism 14 is now described in the following with
reference to FIG. 2.
The engine to which the present invention is applied consists of a
so-called DOHC engine in which intake valves and exhaust valves are
actuated by separate camshafts, and each cylinder is equipped with two
intake valves and two exhaust valves. Since the intake valves and the
exhaust valves have a substantially same structure, only the part of the
valve actutating mechanism 14 related to the intake valve mechanism is
described in the following.
For each cylinder, three rocker arms 41, 42 and 43 are pivotally supported
on a rocker arm shaft 40, which is secured to a cylinder head, adjacent to
each other and so as to be rotatable in mutually independent manner. A
camshaft 45 is rotatably supported by cam journals 44 formed in the
cylinder head above the rocker arms 41, 42 and 43.
The camshaft 45 is provided, for each cylinder, a pair of low speed cams
46a and 46b having a relatively small angular interval of opening the
valves and a relatively small valve lift, and a single high speed cam 47
having a relatively large open valve angular interval and a relatively
large valve lift. A pair of oil supply tubes 48 and 49 are disposed above
the camshaft 45 to lubricate the camshaft 45 and the sliding surfaces
between the cams and the rocker arms. The free ends of the first and
second rocker arms 41 and 42 which engage with the low speed cams 46a and
46b abut the upper ends of the valve stems of a pair of intake valves 50a
and 50b which are elastically urged towards their closed positions. The
third rocker arm 43 disposed between the first and second rocker arms 41
and 42 and engaged with the high speed cam 47 is engaged, at its lower
end, with a lost motion spring not shown in the drawing so as to be
normally urged upwards.
The first through third rocker arms 41 through 43 which are disposed one
next to the other are internally provided with a coupling switch over unit
51 which consists of a guide bore passed through the rocker arms and
switching pins slidably received therein.
The first rocker arm 41 is provided with a first guide bore 52 having an
open end facing the third rocker arm 43 and a closed bottom end and
extending in parallel with the rocker arm shaft 40. The first guide bore
52 slidably receives a first switching pin 53 therein. The bottom end of
the first guide bore 52 defines an oil pressure chamber 54 which is
communicated with an oil supply passage 57 defined in the hollow rocker
arm shaft 40 via an oil passage 55 defined in the first rocker arm 41 and
an oil supply opening 56 opening on an outer periphery of the rocker arm
shaft 40.
The third rocker arm 43 is provided with a second guide bore 58 which
extends in parallel with the rocker arm shaft 40 and aligns with the first
guide bore 52 in a conformal and coaxial manner at its rest position in
which its cam slipper engages with a base circle of the high speed cam 47,
and a second switching pin 59 is received therein with its one end
abutting the first switching pin 53.
The second rocker arm 42 is likewise provided with a third guide bore 60
having a closed bottom end, and receives a stopper pin 61 therein with its
one end abutting the other end of the second switching pin 59.
The stopper pin 61 has a stem portion 63 which is received in a guide
sleeve 62 fitted into the bottom end of the third guide bore 60, and is
normally elastically urged towards the third rocker arm 43 by a return
spring 64.
By displacing the first and second switching pins 53 and 59 in lateral
direction as seen in FIG. 2 by means of the oil pressure applied to the
oil pressure chamber 54 and the biasing force of the return spring 64, one
can selectively obtain either a state in which the rocker arms 41 through
43 can move independently as shown in FIG. 2 and another state in which
the rocker arms 41 through 43 are integrally coupled with one another by
the switching pins 53 and 59 straddling between the adjacent rocker arms
so as to simultaneously and jointly actuate the two intake valves 50a and
50b.
The downstream end of the oil supply passage 57 internally provided in the
rocker shaft 40 is connected to one of the aforementioned oil passages or,
more specifically, the high speed oil supply tube 49. This high speed oil
supply tube 49 is provided with a nozzle 65 for spraying lubricating oil
onto a position corresponding to the high speed cam 47.
The other oil supply tube or the low speed oil supply tube 48 is connected
to a lubrication oil passage 66 branched off from an oil gallery. This low
speed oil supply tube 48 is provided with nozzles 67 for spraying
lubricating oil onto positions corresponding to the cams 46a, 46b and 47,
and, additionally, supplies lubricating oil to the cam journals 44 via oil
passages 68.
The switching control valve 17, mounted on the cylinder head, is actuated
by oil pressure which is supplied from the solenoid valve 16 controlled by
the aforementioned control signal and is internally provided with a spool
valve 70 which is normally urged towards its closed position by a return
spring 69.
When this spool valve 70 is at its upper closed position (as shown in FIG.
2), an inflow port 72 leading to the lubrication oil passage 66 is
communicated, via an oil filter 71, with an outflow port 73 leading to the
oil passage 57 in the rocker arm shaft 40 solely through an orifice 74. At
the same time, the outflow port 73 is communicated with a drain port 75
opening into an upper space of the cylinder head, and the oil pressure in
the oil supply passage 57 thereby drops. Therefore, no oil pressure is
supplied to the oil supply passage 57, and the pins 53 and 59 are urged
all the way towards the oil pressure chamber 54 under the spring force of
the return spring 64 so that the rocker arms are independently actuated by
the associated cams and undergo independent angular displacements. In this
case, the oil supplied by the oil pump 15 from an oil pan 76 to the oil
gallery is supplied to the low speed lubricating oil supply tube 48 via
the lubricating oil passage 66, and lubricates the sliding surfaces
between the cams and the rocker arm as well as the cam journals 44.
When the spool valve 70 is switched over to its lower open position, the
inflow port 72 communicates with the outflow port 73 via an annular groove
77 of the spool valve 70 and the outflow port 73 is disconnected from the
drain port 75 so that oil under pressure is supplied from the lubrication
oil passage 66 to the oil supply passage 57. As pressurized oil is thereby
supplied to the oil pressure chamber 54 of the first rocker arm 41, the
first and second switching pins 53 and 59 are forced into the second guide
bore 58 and the third guide bore 60, respectively, against the biasing
force of the return spring 64, and the rocker arms 41 through 43 are
integrally coupled with one another. The pressurized oil which is supplied
to the oil supply passage 57 to actuate the coupling switch over unit 51
is then supplied to the high speed lubrication oil supply tube 49 via the
downstream end of the oil supply passage 57 to lubricate the sliding
surface between the high speed cam 47 and the third rocker arm 43.
The spool valve 70 is switched over to its open position, against the
biasing force of the return spring 69, by a pilot pressure which is
applied to the upper end of the spool valve 70 from a pilot oil passage 78
branched off from the inflow port 72. The normally closed solenoid valve
16 is interposed in this pilot oil passage 78 and the solenoid of this
solenoid valve 16 is controlled by an output signal from the electronic
control circuit 21 in such a manner that the spool valve 70 is brought to
its open position by opening of the solenoid valve 16 to thereby achieve a
high speed valve timing and the spool valve 70 is brought to its closed
position by closing of the solenoid valve 16 to thereby achieve a low
speed valve timing.
The switching action of the spool valve 70 is monitored by the oil pressure
switch 22 provided in the housing of the switching control valve 17 to be
turned on and off depending on the detection of a low pressure and a high
pressure, respectively, in the oil pressure of the outflow port 73.
The variable capacity turbocharger 7 is now described in the following with
reference to FIG. 3. As this turbocharger 7 is conventional as far as its
compressor unit 8 is concerned, only its turbine unit 12 is described in
the following.
A turbine casing 80 of the turbocharger 7 is provided with a scroll passage
81 whose cross sectional area gradually diminishes towards its downstream
end, and an exhaust outlet 82 opens out from this scroll passage 81 along
its tangential direction. In a central part of the scroll passage 81 is
located a turbine wheel 83 which is integrally attached to an end of a
turbine shaft coaxial to the compressor shaft.
Inside the scroll passage 81 are provided four arcuate fixed vanes 84 which
are integral with the turbine casing 80 and arranged on a common circle at
an equal interval and an equal width. Thus, the scroll passage 81 is
separated into an outer passage 85 and an inner passage 86 by these fixed
vanes 84.
Four moveable vanes 87, each having a substantially same curvature as the
fixed vanes 84, are disposed between the fixed vanes 84 on the same common
circle as the fixed vanes 84. Each of the moveable vanes 87 is pivotally
supported at one of its circumferential ends so as to be pivotable only
into the interior of the aforementioned common circle and define a
continual aerofoil in cooperation with the adjacent fixed vane 84. The
inclination angle of each of the moveable vanes 87 is continually
controlled by a moveable vane actuation control unit which will be
described hereinafter.
The moveable vane actuation control unit comprises lever members 89 each
projecting integrally from a pivot shaft 88 of each of the moveable vanes
87, a pair of see-saw members 91 each pivotally supported and provided
with slots 90 on either end thereof for engagement with two of the lever
members 89, a pair of link arms 94 each of which is coupled with the pivot
shaft 92 of one of the see-saw members 91 at its one end and to a link rod
93 at its other end, and the actuator 18 serving as a drive source for the
moveable vanes 87. This actuator 18 is provided with a drive shaft 95
which is adapted to move axially back and forth by fluid pressure and
coupled with the link rod 93 via a coupling rod 96.
In the above described link mechanism, the drive shaft 95 and the coupling
rod 96 are coupled with each other by a ball joint 97, and the coupling
rod 96 and the link rod 93 are coupled with each other by a crevice joint
98 in such a manner that the actuating force from the actuator 18 may be
smoothly transmitted to the link arm 94. Also, in order to define fully
open positions of the moveable vanes 87 by restricting the stroke of the
drive shaft 95, the coupling rod 96 is integrally provided with a stopper
101 which abuts an adjusting bolt 100 threaded with a bracket 99
integrally mounted on the turbine casing 80.
The actuator 18 comprises a cup-shaped casing 102, and a diaphragm 103
secured to the open end of the casing 102 by crimping a cover 103 thereon,
and this diaphragm 103 defines a negative pressure chamber 105 and a
positive pressure chamber 106 in the actuator 18.
A base end of the drive shaft 95 is attached to a central part of the
diaphragm 104 via retainers 107 and 108. A compression coil spring 109 is
interposed between the retainer 107 facing the negative pressure chamber
105 and the bottom wall of the casing 102 to normally urge the diaphragm
104 along with the drive shaft 95 towards the cover 103 or rightward as
seen in FIG. 3.
The drive shaft 95 is slidably supported by a central part of the bottom
wall of the casing 102. The part of the drive shaft 95 projecting out of
the bottom wall of the casing 102 is enclosed, in an air-tight fashion, by
a soft and frictionless bellows 110 which is made by cutting a cylindrical
fluoride resin member in an annular fashion from both inside and outside
in an alternating fashion at small interval. The interiors of the negative
pressure chamber 105 and the bellows 101 are communicated with each other
by a through hole 111.
The casing 102 is provided with a negative pressure introduction inlet 112
for communicating the negative pressure chamber 105 with the outside and
the cover 103 is provided with a positive pressure introduction inlet 113
for communicating the positive pressure chamber 106 with the outside.
In this actuator 18, when a positive pressure is introduced into the
positive pressure chamber 106 from the positive pressure introduction
inlet 113, the diaphragm 104 is pushed leftward as seen in FIG. 3 against
the biasing force of the compression coil spring 109, and the drive shaft
95 is driven leftward. When a negative pressure is introduced into the
negative pressure chamber 105 from the negative pressure introduction
inlet 112, the drive shaft 95 is likewise driven leftward by the diaphragm
104. In other words, in a low opening angle range of the throttle valve
where the intake negative pressure P.sub.B is high, the actuator 18 is
actuated so as to push out the drive shaft 95. The link rod 93 is thereby
moved leftward as seen in FIG. 3, and this in turn causes the link arms 94
to turn in clockwise direction around the pivot shaft 92 with the result
that the moveable vanes 87 are turned inwards around the pivot shafts 88
by being actuated by the lever members 89 which are each engaged with the
slots 90 on either end of each of the see-saw members 91. By thus opening
the moveable vanes 87, one can obtain a maximum capacity condition in
which the nozzle gaps G.sub.N defined between the leading edges of the
fixed vanes 84 and the trailing edges of the moveable vanes 87 are
maximized (as shown by the imaginary lines in FIG. 3).
When the supply of a negative pressure P.sub.B to the negative pressure
chamber 105 is discontinued by controlling the aforementioned solenoid 35
for controlling negative pressure, the negative pressure in the negative
pressure chamber 105 is reduced and the drive shaft 95 is pulled in by the
spring force of the coil spring 109. As a result, the link rod 93 is moved
rightward as seen in FIG. 3, and the see-saw members 91 are turned by the
link arms 94 in counter-clockwise direction so that the moveable vanes 87
are moved outward around the pivot shafts 88 by way of the lever members
89 each engaged with the slots 90 on either end of each of the see-saw
member 91 (as shown by the dotted lines in FIG. 3). By thus closing the
moveable vanes 87, one can obtained a minimum capacity condition in which
the nozzle gaps G.sub.N defined between the leading edges of the fixed
vanes 84 and the trailing edges of the moveable vanes 87 are minimized.
Therefore, the exhaust gas flow is narrowed and accelerated to the maximum
extent and drives the turbine wheel 83 as a circular flow flowing through
the inner circumferential passage 86 so as to maximize the effect of
supercharging in low speed range of the engine.
As the rotational speed of the engine is increased and a sufficient
supercharge effect is obtained, the solenoid valve 34 for positive
pressure control is controlled and a supercharge pressure P.sub.2 is
introduced into the positive pressure chamber 106. The actuator 18 is
thereby actuated in the direction to push out the drive shaft 95, and the
see-saw members 91 are turned in clockwise direction by turning the link
arms 94 in opposite direction so that the moveable vanes 81 may be turned
inward by way of the lever members 89. By thus expanding the nozzle gaps
G.sub.N, no acceleration is effected on exhaust gas and the exhaust gas
encounters little flow resistance whereby the engine is subjected to less
back pressure.
The opening amount of the moveable vanes 87 was controlled primarily by the
solenoid valve 34 for positive pressure control in the present embodiment,
but it is also possible to use the solenoid valve 35 for negative pressure
control in combination.
A control program incorporated in the electronic circuit 21 to control the
solenoid valve 16 for valve timing switch-over is now described in the
following with reference to FIGS. 4a and 4b.
In the first step 201 it is determined whether an initial mode has been
started or, in other words, the engine is being cranked or not. If the
engine is being cranked, an elapsed time T.sub.DST (for instance 5
seconds) after starting of the engine is set up and the measurement of
time after starting of the engine is set ready in the second step 202.
Then in the third step 203, a valve close command is issued to the
solenoid valve 16, and the engine is operated at low speed valve timing.
In the fourth step 204, an elapsed time T.sub.DHVT (for instance 0.1
second) after switching over to a high speed valve timing is set and the
measurement of delay time after a switch-over to a high speed valve timing
is set ready. In the fifth step 205, maps T.sub.IL and .theta.IGL
corresponding to a low speed valve timing operation is selected as a basic
fuel injection amount determining map and ignition timing map which are to
be used in a fuel injection control routine, and in the sixth step 206 a
revolution limit value N.sub. HFC for cutting off fuel supply is set to a
value N.sub.HFCL which corresponds to low speed valve timing operation.
Now, the amount of fuel injection T.sub.OUT is given by the following
formula:
T.sub.OUT =K1T.sub.1 +K2
where T.sub.1 is a basic amount of fuel injection, K1 is a correction
factor, and K2 is a constant term.
K1 accounts for an intake temperature correction factor K.sub.TA for
increasing fuel supply when the intake temperature T.sub.A is low, a
cooling water temperature correction factor K.sub.TW for increasing fuel
supply when the cooling water temperature T.sub.W is low, a high load fuel
boosting factor K.sub.WOT which increases fuel supply in a high speed
range determined by the engine rotational sped N.sub.E, the intake
negative pressure P.sub.B and the throttle opening angle .theta..sub.TH,
and a feedback correction factor for correcting the deviation of the
air/fuel ratio from a theoretical ratio in an O.sub.2 feedback region of a
relatively low speed range (for instance 4,000 rpm), while K2 accounts for
an acceleration fuel boosting factor which increases fuel supply during
acceleration of the engine.
The basic amount of fuel injection T.sub.I is experimentally determined so
that intake mixture achieves a target air/fuel ratio which is close to an
ideal air/fuel ratio according to the amount of air introduced into the
cylinder in each particular operating condition of the engine as
determined by the rotational speed of the engine N.sub.E and the intake
negative pressure P.sub.B, and the electronic control circuit 21 stores a
T.sub.IL map for low speed valve timing operation and a T.sub.IH map for
high speed valve timing operation, as a T.sub.I map.
The shorter the angular interval of opening the valves, the greater the
valve acceleration becomes during the opening phase of the valves. At the
same time, as the valve acceleration increases, the rotational speed
N.sub.E of the engine at which the valves start jumping becomes lower.
Therefore, the permissible maximum rotational speed of the engine differs
depending on whether a high speed valve timing condition or a low speed
valve timing condition is being selected as they have different intervals
of opening the valves. According to the present embodiment, the revolution
limiter value N.sub.HFCL is set to a relatively low value (for instance
7,500 rpm) during low speed valve timing operation and to a relatively
high value (for instance 8,100 rpm) during high speed valve timing
operation.
If it is determined in the first step 201 that the engine is not being
cranked or, in other words, the engine is already running, it is then
determined in the seventh step 207 whether signals from various sensors
are being normally supplied to the electronic control circuit 21 or not,
or, in other words, a determination is made whether a fail-safe condition
exists or not.
If it is judged that no fail-safe situation exists or, in other words, a
normal condition exists, the time remaining from the time interval
T.sub.DST after starting the engine, which was set up in the second step
202, is evaluated in the eighth step 208. If the remaining time is not
zero, the system flow advances to the third step 203. If there is no
remaining time, the system flow advances to the ninth step 209 where it is
determined whether the cooling water temperature T.sub.W is less than a
predetermined temperature T.sub.W1 (for instance 60 degrees C.) or not,
or, in other words, whether the engine has been warmed up or not. If the
cooling water temperature T.sub.W is found to be less than the
predetermined temperature T.sub.W1, the system flow advances to the third
step 203, and if the cooling water temperature T.sub.W is found to be
equal to or higher than the predetermined temperature T.sub.W1 it is
determined in the tenth step 210 whether the vehicle speed V is lower than
a certain extremely low speed level V.sub.1 (which may contain hysteresis
and range from 5 to 8 km/h) or not. If the vehicle speed V is lower than
the extremely low speed level V.sub.1 the system flow advances to the
third step 203, and if the vehicle speed V is equal to or higher than the
extremely low speed level V.sub.1 it is determined if the vehicle is
equipped with a manual transmission system MT or not in the eleventh step
211.
Thus, a low speed valve timing condition is produced and, at the same time,
a corresponding mode of fuel injection control is selected before
starting, during cranking, immediately after starting, before engine
warm-up, while stopping or while running slowly. This measure is taken in
order to prevent occurrence of faulty operation of the coupling
switch-over unit 51 due to the viscosity of lubricating oil and occurrence
of abnormal combustion.
If it is determined in the eleventh step 211 that the vehicle is not
equipped with a manual transmission system MT or it is equipped with an
automatic transmission AT, it is determined in the twelfth step 212 if
parking range P or a neutral range N is selected for its shift position or
not. If either P or N range is selected, it is determined in the
thirteenth step 313 whether a T.sub.TH map for high speed valve timing was
selected in the previous cycle or not. If not, the system flow returns to
the third step 203. If the vehicle is equipped with a manual transmission
system, a comparison is made, in the fourteenth step 214, between a
rotational speed lower limit N.sub.E1 (which may range between 4,800 and
4,600 rpm including hysteresis), below which the engine output in low
speed valve timing condition is always higher that that in high speed
valve timing condition, and the current rotational speed N.sub.E of the
engine. If N.sub.E is less than N.sub.E1, it is determined in the
fifteenth step 215 if the T.sub.TH map for high speed valve timing was
used in the previous cycle or not in the same way as in the thirteenth
step 213. If not, the system flow advances to the third step 203.
It can be seen in the preceding steps that low speed valve timing is
selected either when the vehicle may be stationary even though the
rotational speed N.sub.E of the engine is high or when, even though the
vehicle may be running, its speed is slow or the rotational speed of the
engine is low, and a high speed running condition has not been experienced
yet.
On the other hand, if it is found in the fourteenth step 214 that N.sub.E
is equal to or higher than N.sub.E1, a T.sub.IL map and the T.sub.IH map
are searched in the sixteenth step 216 according to the subroutine shown
in FIG. 4c to find the values of T.sub.IL and T.sub.IH which correspond to
the rotational speed N.sub.E and the intake negative pressure P.sub.B of
the engine at the current stage. Then, in the seventeenth step 217, a
T.sub.VT value corresponding to the current value of N.sub.E is obtained,
according to the subroutine given in FIG. 4d, from a high load
determination value table T.sub.VT which is experimentally determined as
such according to the amount of fuel injection.
As for values from the T.sub.IL and T.sub.IH maps, the T.sub.IL value which
is to be used in the sixteenth step 216 consists of a value searched from
the T.sub.IL map when a command to open the solenoid value 16 was not
present in the previous cycle, and of a value obtained from the T.sub.IL
map less a prescribed amount of hysteresis .DELTA. T.sub.I when a command
to open the solenoid value 16 was present in the previous cycle. A similar
process is executed in regards to the arithmetic process in the
seventeenth step 217 for determining a T.sub.VT value. The T.sub.VT value
which is to be used in the seventeenth step 217 consists of a value
searched from a T.sub.VT table when a command to open the solenoid value
16 was not present in the previous cycle, and of a value obtained from the
T.sub.VT table less a prescribed amount of hysteresis .DELTA.T.sub.VT when
a command to open the solenoid value 16 was present in the previous cycle,
whereby a hysteresis is given to the switching property of the amount of
fuel injection at the point of valve timing switch over.
In the subsequent eighteenth step 218, a comparison is made between this
T.sub.VT value and the amount of fuel injection T.sub.OUT in the previous
cycle. If the T.sub.OUT is found to be less than T.sub.VT, a comparison is
made in the nineteenth step 219 between a rotational speed upper limit
N.sub.E2 (which may range between 5,900 and 5,700 rpm including
hysteresis) above which the engine output in high speed valve timing
condition is always higher that that in low speed valve timing condition,
and the current rotational speed N.sub.E of the engine. If N.sub.E is less
than N.sub.E2, a comparison is made in the twentieth step 220 between the
T.sub.IL value and the T.sub.TH value obtained in the sixteenth step 216,
and, if T.sub.IL is found to be larger than T.sub.IH, a valve close
command is supplied to the solenoid valve 16 in the twenty first step 221
or, in other words, low speed valve timing is selected.
If it was found in the thirteenth step 213 or in the fifteenth step 215
that the T.sub.TH map was selected in the previous cycle or a low load and
low rotational speed condition is produced after experiencing a high speed
running condition, the system flow advances to the twentyfirst step 221.
On the other hand, if its was found in the eighteenth step 218 that
T.sub.OUT is equal to or larger than T.sub.VT, if it was found in the
nineteenth step 219 that N.sub.E is equal to or larger than N.sub.E2, or
if it was found in the twentieth step 220 that T.sub.IL is equal to or
less than T.sub.IH, a valve open command is issued to the solenoid valve
16 or, in other words, a high speed valve timing condition is selected.
Thus, it can be seen that a point of switch over between high speed valve
timing and low speed valve timing is determined from the rotational speed
of the engine and the demanded amount of fuel injection.
An adjustment is made so as to have a relatively rich mixture in high load
range, and a high speed valve timing operation is more desired for
increasing the engine output in high load range. However, if the point of
switching over valve timing is determined in a definite fashion, a hunting
may occur in a boundary region and a shock may be produced due to an
abrupt change in the torque output at the point of switch over. Therefore,
according to the present embodiment, an optimum switch over control action
is obtained by carrying out the composite steps of the eighteenth through
twentieth steps 218 through 220.
Following selection of high speed valve timing operation, it it determined
in the twenty third step if zero value is placed in a flag F.sub.LVT to
indicate that low speed valve timing operation is not selected in the
turbocharger control routine which will be described hereinafter. If it is
determined that the turbocharger end presupposes low speed valve timing
operation, the system flow advances to the third step 203. Otherwise, the
system detects a signal from the oil pressure switch 22 in order to
monitor the operating condition of the switch over control valve 17. If it
is found that the oil pressure switch 22 is off or that oil pressure is
acting upon the coupling switch over unit 51, the remaining time of the
delay time T.sub.DHVT following the activation of the coupling switch over
unit, which was previously set up in the fourth step 204, is determined in
the twenty-fifth step 225. If T.sub.DHVT =0, in the twenty-sixth step 226,
a preparation for actuation of a delay timer is made following the setting
up of a timer for the elapsed time T.sub.DLVT (for instance 0.2 seconds)
following a switch over to low speed valve timing operation. Then, a fuel
injection amount map T.sub.IH and an ignition timing T.sub.IGH
corresponding to high speed valve timing operation are selected in the
twenty-seventh step 227, and a revolution limiter value N.sub.HFC is
changed to a value N.sub.HFCH for high speed valve timing in the
twenty-eighth step 218.
Following the issuing of a valve close command to the solenoid valve 16 in
the twenty-first step 221, presence of an oil pressure switch signal
O.sub.P is detected in the twenty-ninth step 229. If the oil pressure
switch 22 is turned on or no oil pressure is being applied to the coupling
switch over unit 51, the time remaining from T.sub.DLVT which was set in
the twenty-sixth step 226 is read out, and, if T.sub.DLVT =0 the system
flow returns to the fourth step 204.
If the oil pressure switch signal O.sub.P is not turned off in the
twenty-fourth step 224 even though a switch over from low speed valve
timing operation to a high speed valve timing operation is effected, the
system flow advances to the thirtieth step 230 and the low speed valve
timing operation is maintained until the oil pressure switch signal
O.sub.P is turned off. C | | |
